Engineering data contents Components Accumulators .......................... 5 Actuators linear.................................. 54 rotary ............................... 152 Air compressors ....................... 11 dryers ................................. 15 filters.................................. 19 logic ....................................21 motors................................ 26 regulators ........................... 30 Airline lubricators ..................25 Boots & bellows.....................32 Carriers for hose & cable .......34 Cartridge valves .................... 35 Clamps for pipe, hose, tubing...37 Control networks................... 39 Cylinders ............................... 54 Directional-control valves..... 62 Electrohydraulic proportional valves ............ 68 servovalves ........................ 68 Fittings ...................................75 Flow-control valves............... 78 Flow meters..........................143 Flushing procedures ...............84 Heat exchangers .................... 87 Hydraulic filters.................................. 91 fluids................................ 103 hose.................................. 110 manifolds ......................... 136 pumps .............................. 119
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Hydrostatic transmissions ... 128 Intensifiers........................... 133 Motors air....................................... 26 hydraulic .......................... 114 low-speed/high-torque..... 134 Pressure-control valves ....... 138 Pressure gages ..................... 143 Pressure switches ................ 146 Reservoirs ........................... 148 Rotary actuators ...................152 Seals & packings ................. 154 Shock absorbers .................. 163 Transducer technology ........ 167 Vacuum technology............. 191
Basic circuits Index of basic circuits ......... 196 Accumulator circuits . . . . . . 197 Air lubrication circuits . . . . 199 Cylinder locking circuits . . 200 Deceleration circuits . . . . . . 202 Decompression circuits . . . 204 Electrical control circuits . . 206 Filter circuits . . . . . . . . . . . . 208 Hydraulic filter circuits . . . . 208 Hydraulic motor circuits . . . 209 Hydraulic speed control . . . 222 Intensifier circuits . . . . . . . . 211 Locking circuits . . . . . . . . . . 200 Meter-in circuits . . . . . . . . . 222 Meter-out circuits . . . . . . . . 222 Motor braking circuit . . . . . 209 Parallel motor circuit . . . . . 210 Pneumatic speed-control . . 212
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Pressure-control circuits . . . 215 Pump-unloading circuits . . . 217 Regenerative circuits . . . . . 219 Safety circuits . . . . . . . . . . . 220 Sequencing circuits . . . . . . . 221 Series motor circuit . . . . . . . 210 Speed-control circuits Hydraulic . . . . . . . . . . . . . 222 Pneumatic . . . . . . . . . . . . . 212 Synchronizing circuits . . . . 225
Fluid power graphic symbols Accumulators . . . . . . . . . . . 242 Air motors . . . . . . . . . . . . . . 230 Compressors . . . . . . . . . . . . 229 Directional-control valves . 234 Cylinders . . . . . . . . . . . . . . . 231 Filters . . . . . . . . . . . . . . . . . . 242 Flow-control valves . . . . . . 239 Hydraulic motors . . . . . . . . 230 Intensifiers . . . . . . . . . . . . . . 233 Proportional valves . . . . . . . 236 Pressure-control valves . . . . 237 Pumps . . . . . . . . . . . . . . . . . . 229 Sources of energy . . . . . . . . . 24
Logic symbols . . . . . . . . . 247 Glossaries Compressed air terms ............ 12 Fluid power terms ............... 249 Hydraulic fluids................... 108
Organizations with interest in fluid power ................. 255 General index ................ 256
Accumulators
A
ccumulators usually are installed in hydraulic systems to store energy and to smooth out pulsations. Typically, a hydraulic system with an accumulator can use a smaller pump because the accumulator stores energy from the pump during periods of low demand. This energy is available for instantaneous use, released upon demand at a rate many times greater than what could be supplied by the pump alone. Accumulators also can act as surge or pulsation absorbers, much as an air dome is used on pulsating piston or rotary pumps. Accumulators will cushion hydraulic hammer, reducing shocks caused by rapid operation or sudden starting and stopping of power cylinders in a hydraulic circuit. There are four principal types of accumulators: the weight-loaded piston type, diaphragm (or bladder) type, spring type, and the hydro-pneumatic piston type. The weight-loaded type was the first used, but is much larger and heavier for its capacity than the modern piston and bladder types. Both weighted and spring types are infrequently found today. Hydro-pneumatic accumulators, Figure 1, are the type most commonly used in industry. Functions Energy storage — Hydro-pneumatic accumulators incorporate a gas in conjunction with a hydraulic fluid. The fluid has little dynamic power-storage qualities; typical hydraulic fluids can be reduced in volume by only about 1.7% under a pressure of 5000 psi. (However, this relative incompressibil-
Charging valve
Charging valve
Gas cap Shell
Body Bladder
Poppet Piston
Spring Hydraluic cap
Hydraulic cap Bladder
Gas
Piston
Fig. 1. Cross-sectional views of typical of bladder and piston-type accumulators.
ity makes them ideal for power transmission, providing quick response to power demand.) Therefore, when only 2% of the total contained volume is released, the pressure of the remaining oil in the system drops to zero. On the other hand, gas, the partner to the hydraulic fluid in the accumulator, can be compressed into small volumes at high pressures. Potential energy is stored in the compressed gas to be released upon demand. Such energy can be compared to that of a raised pile driver ready to transfer its tremendous energy upon the pile. In the piston type accumulator, the energy in the compressed gas exerts pressure against the piston separating the gas and hydraulic
fluid. The piston in turn, forces the fluid from the cylinder into the system and to the location where useful work will be accomplished. Pulsation absorption — Pumps, of course, generate the required power to be used or stored in a hydraulic system. Many pumps deliver this power in a pulsating flow. The piston pump, commonly used for its high pressure capaFor more information about using accumulators in hydraulic systems, please refer to the Basic Circuits section which appears elsewhere in this handbook.
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A C C U M U L AT O R S
(a)
(b)
(c)
(d)
(e)
(f)
Gas
Fig. 2. Six stages of operation for bladder- and piston-type accumulators: stage (a), the accumulator is empty -- no gas charge; stage (b), the accumulator has been precharged with dry nitrogen; stage (c), system pressure exceeds precharge pressure, and hydraulic fluid flows into accumulator; stage (d), system pressure peaks, maximum fluid has entered the accumulator, and system relief opens; stage (e), system pressure drops, precharge pressure forces fluid from the accumulator and into the system; and stage (f), system pressure reaches minimum needed to do work.
bility, can produce pulsations detrimental to a high-pressure system. An accumulator properly located in the system will substantially cushion these pressure variations. Shock cushioning — In many fluid power applications, the driven member of the hydraulic system stops suddenly, creating a pressure wave that travels back through the system. This shock wave can develop peak pressures several times greater than normal working pressures and can be the source of system failure or objectionable noise. The gas cushion in an accumulator, properly placed in the system, will minimize this shock. An example of this application is the absorption of shock caused by suddenly stopping the loading bucket on a hydraulic front end loader. Without an accumulator, the bucket, weighing over 2 tons, can completely lift the rear
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wheels of a loader off the ground. The severe shock to the tractor frame and axle, as well as operator wear and tear, is overcome by adding an adequate accumulator to the hydraulic system. Supplementing pump flow — An accumulator, capable of storing power can supplement the hydraulic pump in delivering power to the system. The pump stores potential energy in the accumulator during idle periods of the work cycle. The accumulator transfers this reserve power back to the system when the cycle requires emergency or peak power. This enables a system to utilize a much smaller pump, resulting in savings in cost and power. Maintaining pressure — Pressure changes occur in a hydraulic system when the liquid is subjected to rising or falling temperatures. Also, there may be pressure drop due to leakage of hydraulic fluid. An accumulator compen-
sates for such pressure changes by delivering or receiving a small amount of hydraulic fluid. If the main power source should fail or be stopped, the accumulator would act as an auxiliary power source, maintaining pressure in the system. Fluid dispensing — An accumulator may be used to dispense small volumes of fluids, such as lubricating greases and oils, on command. Operation When sized and precharged properly, accumulators normally cycle between stages (d) and (f), Figure 2. The piston will not contact either cap in a piston accumulator, and the bladder will not contact the poppet or be compressed so that it becomes destructively folded into the top of its body. Manufacturers specify recommended precharge pressure for their ac-
A C C U M U L AT O R S
Table 1–Relative outputs, 10 gal accumulator
Compression ratio 1/2
Fig. 3. Horizontally mounted accumulator can cause uneven bladder wear and trap fluid away from the hydraulic valve.
cumulators. In energy-storage applications, a bladder accumulator typically is precharged to 80% of minimum hydraulic system pressure and a piston accumulator to 100 psi below minimum system pressure. Precharge pressure determines how much fluid will remain in the accumulator at minimum system pressure. Correct precharge involves accurately filling an accumulator’s gas side with a dry inert gas, such as nitrogen, while no hydraulic fluid is in the fluid side. Accumulator charging then begins when hydraulic fluid is admitted into the fluid side, and occurs only at a pressure greater than the precharge pressure. During charging, the gas is compressed to store energy. A correct precharge pressure is the most important factor in prolonging accumulator life. The care with which precharging must be accomplished and maintained is an important consideration when choosing the type of accumulator for an application, all else being equal. If the user tends to be careless about gas pressure and relief valve settings, or adjusts system pressures without making corresponding adjustments to precharge pressure, service life may be shortened, even if the correct type of accumulator was selected. If the wrong accumulator was selected, premature failure is almost certain. Mounting position The optimum mounting position for any accumulator is vertical with the hydraulic port down. Piston models can be horizontal if the fluid is kept clean. When solid contaminants are present or expected in significant amounts, horizontal mounting can result in uneven or accelerated seal wear. Maximum service life can be achieved in the horizon-
System pressure, psi
maximum 1
Recommended precharge, psi
minimum 2
bladder 3
piston 4
Output, gal
bladder 5
piston 6
1.5 2.0
3000 3000
2000 1500
1600 1200
1900 1400
2.53 3.80
3.00 4.41
3.0 6.0
3000 3000
1000 500
800 –
900 400
5.06 –
5.70 6.33
tal position with multiple piston seals to balance the piston’s parallel surface. A bladder accumulator also can be mounted horizontally, Figure 3, but uneven wear on the bladder as it rubs against the shell while floating on the fluid can shorten life. The amount of damage depends on fluid cleanliness, cycle rate, and compression ratio (defined as maximum-system-pressure/ minimum-system-pressure). In extreme cases, fluid can be trapped away from the hydraulic end, which reduces output or may elongate the bladder to force the poppet closed prematurely. Sizes and outputs Available sizes and capacities also influence which accumulator type to choose. Piston accumulators of a particular capacity often are supplied in a choice of diameters and lengths, Table 1. Furthermore, piston designs can be built to custom lengths for little or no price premium. Bladder accumulators are offered only in one size per capacity, with fewer capacities available. The inherently higher output of the piston accumulator may make it the best alternative when space is tight.
Table 1 lists outputs for 10-gal piston and bladder accumulators operating isothermally as auxiliary power sources over a range of minimum system pressures. The differences in precharge pressure, columns 3 and 4, (determined by 80% of minimum system pressure for bladder models, 100 psi below minimum for piston) lead to a substantial difference in outputs, columns 5 and 6. To prevent excessive bladder deformation and high bladder temperatures, also note in Table 1 that bladder accumulators should be specified with compression ratios greater than 3:1. Multiple components Although bladder designs are not available in sizes over 40 gal, piston designs are currently supplied up to 200 gal in a single vessel. Economics and available installation space have led engineers to consider multiple component installations. Two of these can cover most high-output applications. The installation in Figure 4 consists of several gas bottles serving a single piston accumulator through a gas manifold. The accumulator portion must be
Gas manifold
Fluid
Fig. 4. Piston accumulators used in conjunction with gas bottles.
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sized so the piston does not repeatedly strike the caps while cycling. One drawback of this arrangement is that a single seal failure could drain the gas system. Because gas bottles often are less ex-
Fluid manifold
Fig. 5. Several accumulators may be manifolded to provide large system flows.
Fig. 6. A small accumulator may do the job if it is remotely connected to an auxiliary gas bottle.
LVDT
pensive than accumulators, one advantage of this setup might be lower cost. Several accumulators, either piston or bladder design, can be mounted on a hydraulic manifold, Figure 5. If using piston accumulators, the piston with the least friction will move first and occasionally could bottom on the hydraulic cap. In slow or infrequently used systems, this is insignificant.
Gas bottle installations Remote gas storage offers flexibility in large and small systems, Figure 6. The gas bottle concept is generally described with this simple formula: accumulator size minus required fluid output equals gas bottle size. For example, an application that calls for a 30-gal accumulator may only require 8 to 10 gal of fluid output. This application, therefore, could be satisfied with a 10-gal accumulator and a 20-gal gas bottle. An accumulator used with remote gas storage generally has the same size port at the gas end as at the hydraulic end to allow unimpeded flow of gas to and from the gas bottle. The gas bottle has an equivalent port in one end and a gas charging valve at the other. These twopiece accumulators can be configured or bent at any angle to fit available space. The gas bottle concept is suitable for either bladder or piston accumulators. Note that bladder accumulators require a special device called a transfer barrier at the gas end to prevent extrusion of the bladder into the gas bottle piping. Again, a piston accumulator should be sized to prevent piston bottoming at either end of the cycle. Bladder designs should be sized to prevent filling to more than 85% or discharging to more than 85% empty. The flow rate between the bladder transfer barrier and its gas bottle will be restricted by the
neck of the transfer barrier tube. Because of these drawbacks, bottle/bladder accumulators should be reserved for special applications. Flow rates and response times Table 2 suggests maximum flow rates for representative accumulator sizes and types. The larger standard bladder designs are limited to 220 gpm, although the rate can be boosted to 600 gpm using an extra-cost, high-flow port. The poppet controls flow rate; excessive flow causes the poppet to close prematurely. Multiple accumulators mounted on a common manifold are needed to achieve flows that are greater than 600 gpm. Allowable flow rates for piston accumulators generally exceed those for bladder designs. Flow is limited by piston velocity, which should not exceed 10 ft/sec to avoid piston seal damage. In high-speed applications, high seal contact temperatures and rapid decompression of nitrogen that has permeated into the seal material can cause blisters, cracks, and pits in the rubber. Bladder accumulators respond more quickly to system pressure variations than do piston types for two reasons: 1. Rubber bladders do not have to overcome the static friction which a piston seal must, and 2. The piston mass does not need to be accelerated and decelerated. In practice, though, the difference in response may not be as great as commonly believed, and is probably insignificant in most applications. Shock suppression Tests at the University of Wisconsin, Madison, indicate that shock control does not necessarily demand a bladder accumulator. With system flow at a
Table 2–Maximum recommended accumulator flow rates
Tubing length makes pumpto-servovalve distance 118 ft
Fig. 7. Test circuit to generate and measure shock waves in system.
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Piston bore, in.
Gpm at 3000 psi Bladder capacity
Bladder Piston
Standard
High-flow
2 4 6
1 qt 1 gal 2.5 gal
100 400 800
60 150 220
– – 600
7 9 12
larger than 2.5 gal
1200 2000 3400
220 220 220
600 600 600
A C C U M U L AT O R S
4000 4000
Trace A
Trace A Trace B
Pressure—psi
Pressure – psi
Trace B
Trace C
2000
0
1000
2000
Time – ms
Fig. 8. Graph indicates results of shock wave tests.
nominal 30 gpm in the test circuit, Figure 7, an internally piloted directional control valve, 118 ft away from the pump, closes to generate a shock. As the shock wave travels from the valve back through the hydraulic lines and around corners and various restrictions, some portion of its energy is consumed while accelerating the mass of fluid in the lines. With 11/4-in. tubing, a 2750-psi relief valve setting, and no accumulator in the circuit, oscilloscope trace A, Figure 8, shows a pressure spike of 385 psi over the relief valve setting. Adding a 1-gal piston accumulator at the valve reduces the transient to 100 psi over relief valve setting, trace B. Substituting a 1-gal bladder accumulator cuts the transient to 78 psi over relief valve setting, trace
(a)
(b)
Fig. 10. Starburst rupture in end of bladder, (a), could indicate loss of elasticity of bladder material due to embrittlement from cold nitrogen gas during precharge. If bladder is forced under poppet, (b), bladder could sustain C-shaped cut from poppet.
2000
0
Trace C
1000 Time—ms
2000
Fig. 9. Results of second test using smaller-diameter tubing.
C, only 22 psi better than the pistontype protection. A second, similar test with 5/ 8-in. tubing and a relief valve setting of 2650 psi results in a pressure spike of 2011 psi over relief valve setting without an accumulator, trace A, Figure 9. A piston accumulator damps the transient to 107 psi over relief valve setting, trace B, while a bladder accumulator damps the transient to 87 psi over relief valve setting, trace C. The difference between accumulator types in shock suppression again was negligible. Servo equipment Another common misconception says that all servo applications require a bladder accumulator. Experience shows that only a small percentage of servos require response times of 25 ms or less, the region where the difference in response between piston and bladder accumulators becomes material. Bladder accumulators should be used for applications requiring less than a 25-ms response, and either type when response of 25 ms or greater is adequate. Setup and maintenance: precharging On newly repaired bladder accumulators, the shell ID should be lubricated with system fluid before precharging. This fluid acts as a cushion, and lubricates and protects the bladder as it unwinds and unfurls. When precharging begins, the initial 50 psi of nitrogen should be introduced slowly. Neglecting these precautions could result in immediate bladder failure.
High-pressure nitrogen, expanding rapidly and thus cold, could channel the length of the folded bladder and concentrate at the bottom. The chilled brittle rubber expanding rapidly could rupture in a starburst pattern, Figure 10(a). The bladder also could be forced under the poppet, resulting in a C-shaped cut in the bladder bottom, Figure 10(b). The fluid side of piston accumulators should be empty during precharging so that gas-side volume is at a maximum. Little damage, if any, can take place during precharging. Too high a precharge pressure or reducing the minimum system pressure without a corresponding reduction in precharge pressure may cause operating problems or damage to accumulators. With excessive precharge pressure, a piston accumulator will cycle between stages (e) and (b), Figure 2, and the piston will range too close to the hydraulic end cap. The piston could bottom at minimum system pressure to reduce output and eventually cause damage to the piston and its seal. The bottoming of the piston often can be heard; the sound serves as a warning of impending problems. Too high a precharge in a bladder accumulator can drive the bladder into the poppet assembly when cycling between stages (e) and (b), Figure 2. This could cause fatigue failure of the spring and poppet assembly, or a pinched and cut bladder if the bag gets trapped beneath the poppet as it is forced closed. Too high a precharge pressure is the most common cause of bladder failure.
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A C C U M U L AT O R S
tor may be thrown to one side of the shell, displacing the bladder and flattening and lengthening it, Figure 12. With this distortion, fluid discharge could cause the poppet valve to pinch and cut the bladder.
Acceleration
Fig. 11. If pressure fluid is allowed into uncharged accumulator, bladder could be crushed or extruded into gas valve and punctured.
Fig. 12. Forces applied perpendicular to bladder accumulator vertical axis can distort bladder shape and risk bladder puncture.
Too low a precharge pressure or an increase in system pressure without a compensating increase in precharge pressure also can cause operating problems, with possible accumulator damage. With no precharge in a piston accumulator, the piston likely will be driven into the gas end cap and probably will remain there. A single contact is unlikely to cause damage. For bladder accumulators, too low or no precharge can have severe consequences. The bladder may be crushed into the top of the shell, then may extrude into the gas valve and be punctured, Figure 11. One such cycle is sufficient to destroy a bladder. Pis-
ton accumulators, therefore, are more tolerant of improper precharging.
Pressure transducer
Pressure transducer
Fluid
Fluid
(a)
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External forces Any application subjecting an accumulator to acceleration, deceleration, or centrifugal force may have a detrimental effect on accumulator operation. Forces along the axis of an accumulator’s tube or shell normally have little effect on a bladder model but may increase or decrease gas pressure in a piston type because of the mass of the piston affects the force. Forces perpendicular to an accumulator’s axis should not affect a piston model, but fluid in a bladder accumula-
(b)
Pressure transducer
Hall proximity sensor
Sonar absolute position sensor
Fluid
Failure prediction Several methods can be used to monitor the precharge pressure of piston accumulators: With the hydraulic system shut down — A pressure transducer or gage located in the gas end cap, Figure 13(a), indicates the true precharge pressure after a working hydraulic system has cooled, and the accumulator does not contain fluid. With the hydraulic system operating — On request, accumulator manufacturers will install a piston-position sensor in an accumulator’s hydraulic end cap, Figure 13(b). This sensor can be connected to a number of electronics packages. With an accurate precharge and after enough system operation for thermal stability, the electronics can be calibrated to provide continuous readout of precharge pressure that corresponds accurately to the true precharge. With the accumulator coupled to a gas bottle — A ferrous or nonferrous sensor can be installed in the accumulator gas end cap, Figure 13(c), to detect when the piston comes within 0.125 in. of the cap. This warning indicates that precharge pressure has dropped, and the system should be shut down and checked.
(c)
Fig. 13. With pressure transducer mounted in cap of piston-type accumulator, (a), actual precharge will be indicated after working system has become dormant and cooled. Piston-position sensor, (b), can provide continuous readout of precharge when connected to proper electronics package. With Hall-effect sensor installed, (c), close proximity of piston to end cap can be indicated. Note: pistons in (b) and (c) must be flat for use with sonar or Hall-effect sensors.
Air compressors C
ompressed air has become one of the most important power media used in industry. What kind of equipment makes the compression process tick? Two basic types of machines compress ambient air for industrial use: positive-displacement and dynamic air compressors. In positive-displacement compressors, ambient air is isolated in a volume that subsequently is mechanically reduced to increase the air’s pressure. The action may use a crankshaft and reciprocating pistons — much like the familiar internal combustion engine — or rotary elements. The most common rotary elements are sliding vanes that move radially, and male and female rotors that mesh as they turn. In dynamic compressors, the mechanical action of rotating impellers accelerates ambient air as it passes through the machine. The additional kinetic energy is converted into pressure energy downstream. Dynamic compressors are identified as centrifugal or axial — depending on the manner in which air flows through them. Thermodynamics The various compression processes are based on the ideal gas laws of thermodynamics. Neither air nor other gases meet all the assumptions implied in these perfect gas laws. However, some knowledge of these laws can be combined with information gained from experimentation to permit an engineering analysis of the compression process. Compression efficiency in any compressor is compared with two theoretical standards —isothermal and adiabatic. (Neither type occurs in an actual compressor because of the unavoidable losses of the real world.) Isothermal compression would occur if the air temperature were kept constant as pressure increases. To keep temperature constant, the heat of compression would
have to be removed continually. This perfect cooling cannot be accomplished in actual practice. The isothermal equation is a statement of Boyle’s Law: P1V1 = P2V2 Adiabatic compression would occur if no heat were transferred to or from the outside during the process. True adiabatic compression also is not attained in practice. The adiabatic equation is: P1V1k = P2V2k where k is the ratio of the specific heat at constant pressure to the specific heat at constant volume, cp/cv. For dry air, k ≈ 1.4. Dynamic compressors follow a polytropic compression cycle in which the relationship of pressure and volume is held constant. The polytropic equation is: P1V1n = P2V2n where n is a constant, determined experimentally for the particular type of machine involved. Significant power can be saved in any type of compressor by using the multi-stage principle. In these compressors, the output from the first stage is fed to the inlet of the second stage, and so on. Cooling between stages saves more power, because the compression process — adiabatic or polytropic — is non-linear with an exponential constant greater than unity. Drivers Electric motors are the most widely used compressor drivers for industrial facilities. (Steam and natural gas engines are other possibilities with limited applications.) Internal-combustion engines serve portable compressors. In almost every case, the compressor manu-
Some basic information about the machines that produce industrial compressed air.
facturer provides the driver and compressor as an assembly according to these general technical considerations: ● voltage and frequency requirements ● any current restrictions — particularly kVA inrush during starting ● motor-to-compressor speed match ● torque requirements for starting and running ● power factor considerations ● cycling considerations ● proper protective enclosure ● ambient temperature range ● desired efficiency, and ● anticipated service factor. Economic considerations involving the drive motor are motor cost, maintenance cost, operating cost, and — most important — an analysis of accessory equipment needed to operate the motor. Along with the starter and motor controls, these include any special transformer requirements and cost of additional power lines. Reciprocating compressors The original industrial air compressors were reciprocating machines and they still are offered in more models and sizes than perhaps any other type. Horsepower ratings today range from fractional to around 600; available pressures can be 6000 psi and higher. This variety makes it possible to find a reciprocating compressor small enough to operate a single function on a machine or large enough to supply a small manufacturing facility. A number of different physical configurations, Figure 1, may even make it possible to match a compressor to the particular space where it must be installed. Reciprocators are the most efficient compressors for the majority of applications. They can be fitted with control systems which match their output almost exactly to operating demands. Modern electronic pressure sensors join sophisticated computer-control systems to se-
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AIR COMPRESSORS
quence a gang of different-size air compressors for maximum efficiency. Reciprocating compressors have one or more cylinders, each fitted with a piston driven by a crankshaft through a connecting rod. Each cylinder also has intake and discharge valves, and some means for cooling the mechanical parts. Ambient air is drawn into the cylinder during its suction stroke. At the end of the suction stroke, the crankshaft reverses the piston’s direction and the air is compressed and expelled during the discharge stroke. When only one end of the piston
contacts the air, the compressor is identified as single-acting. when both ends of the piston act on the air, the compressor is double-acting. Obviously, a double-acting compressor discharges approximately twice as much air per cylinder per cycle as a comparable-size single-acting machine. As in other high-cycling machines, lubrication and cooling are important to the operation of reciprocating compressors. Depending on compressor size, splash lubrication, pressurized crankcases, or pumped lubrication may provide the former of these functions. Wa-
Glossary and gas laws Adiabatic compression occurs when no heat is transferred to or from the air during compression. Aftercooling — the cooling of air after it has been compressed to lower its temperature and precipitate condensed vapors. Boyle’s Law states that the absolute pressure of a fixed mass of gas varies inversely as the volume, provided the temperature remains constant. Charles’ Law states that the volume of a fixed mass of gas varies directly with the absolute temperature, provided that pressure remains constant. Clearance is the volume of a reciprocating compressor’s cylinder not swept by the piston’s movement. It includes space between the piston and the head at the end of the compression stroke, and typically is expressed as a percentage of cylinder displacement. The clearance may be different at the two ends of a double-acting cylinder. Compression efficiency is the ratio of the theoretical work required (in a given process) to the actual work required to be done to compress and deliver the air. Expressed as a percentage, compression efficiency accounts for fluid-friction losses, leakage, and thermodynamic variations from the theoretical process. Compression ratio is the ratio of the absolute discharge pressure to the absolute intake pressure. Density is the weight of a given volume of air, usually expressed in lb/ft3 at standard temperature and pressure. Displacement is the net volume
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swept by the moving parts of a compressor per unit of time. This term applies only to positive-displacement compressors. Free air is air at the atmospheric conditions at a specific location. Ideal gases follow the perfect gas laws without deviation. There are no real ideal gases, but they provide a common starting point for calculations and corrections. Intercooling is the process of cooling air between stages of compression to liquefy condensed vapors and save power by reducing the temperature of air entering the next stage. Isothermal compression occurs when the temperature of the air remains constant during compression. Mechanical efficiency is the ratio of the indicated horsepower to the actual shaft horsepower. Polytropic compression occurs when heat is transferred to or from the air at a precise rate during compression so that PVn is constant. Pumping or surge is the reversal of flow within a dynamic compressor. It takes place when insufficient pressure is generated to maintain flow. Standard pressure and temperature is generally defined as 68º F and 14.70 psia with a relative humidity of 36%. Volumetric efficiency is the ratio of the actual volume of air admitted (at a specified temperature and pressure) to the full piston-displacement volume — obviously for reciprocating compressors only.
Fig. 1. Reciprocating compressors. Upper image: Kaeser Compressor’s oil-less Airbox, in sound-dampening enclosure with anti-vibration frame and padding, generates noise levels only 66 to 69 dB(A). Models from 4 to 15 hp have TEFC motors. Lower image: Thomas Industries’ Model HP-1500 11⁄2-hp piston air compressor can flow 5.2 cfm at 100 psi.
ter is the most-common coolant for aircompressor cylinders, intercoolers, and aftercoolers, although some smaller models may be air-cooled. Rotary-vane compressors The basic elements of this type of compressor are a cylindrical case with an internal eccentrically mounted rotor, Figure 2. Vanes fit into radial slots in the rotor, and move centrifugally outward until they meet the case’s inside surface as the rotor turns. When the vanes pass inlet ports in the case, they form pockets to trap air. These pockets decrease in size due to the rotor’s eccentric location. The trapped air is compressed, and then expelled when the pocket reaches a discharge port. Because of sliding friction in the slots and tip wear against the case, the vanes are the most wear-prone parts of this type compressor. Vane length, vane rubbing speed, and bending forces on the extended vanes all tend to limit
AIR COMPRESSORS
Fig. 2. A typical rotary-vane compressor has oil injected during compression to absorb some heat of compression. Air exiting from vane and screw compressors is delivered to a separator where liquid oil is removed.
the flow capacity and pressure ratings of rotary-vane compressors. Oil frequently is injected into rotaryvane compressors to help with sealing, provide lubrication, and absorb some of the heat of compression. In older installations, the combination of high heat and oil sometimes formed sludge and varnish deposits. Modern lubricating oils have eliminated that situation. Rotary-screw compressors Rotary-screw compressors, Figure 3, have become the dominant design because they literally offer more horsepower per dollar. They can run for ex-
tended periods of time with little maintenance required, particularly when matched with the sophisticated control and warning systems available today. These machines are built in two versions: wet (or oil-flooded) and dry. In either one, air is compressed by the action of two intermeshing rotors which turn inside a housing. Filtered ambient air is drawn into the voids formed as male and female rotors unmesh, then trapped and sealed as these voids pass the intake ports. As rotation continues, the volume of the voids decreases and pressure increases as the rotors remesh. In oil-flooded machines, cooling oil is sprayed into the housing during compression. This oil absorbs the heat of compression as it is generated, lubricates all dynamic contact surfaces, and also forms a seal between rotors and between rotors and housing walls. The oil then is separated from the compressed air, cooled, filtered, and returned to the injection point for re-use. Dry screw compressors are machined to extremely close tolerances so that the two rotors do not touch. Therefore lubrication of the compression chamber is unnecessary. These compressors can operate at high rotational speeds. A cooling system still is needed to remove the heat of compression. Centrifugal compressors Centrifugal compressors develop pressure within themselves, indepen-
Fig. 3. TS Series tandem air compressor — available in 100- to 600-hp sizes — has two sets of rotors that divide compression process equally between two stages to develop flow capacities from 515 through 3100 cfm.
dent of load — but the load determines the flow to be handled. This general statement is, of course, limited by the physical size of the machine and the power of its driver. In its simplest form, a centrifugal compressor is a single-stage, singleflow machine with its impeller overhung on its drive, Figure 4. Air enters the unit through the inlet nozzle, which is proportioned so that the air arrives at the impeller with a minimum of shock and turbulence. The impeller receives air from the inlet nozzle and dynamically compresses it. The impeller also sets the air in motion, achieving a velocity somewhat less than the tip speed of the impeller. A diffuser chamber surrounds the impeller and receives air leaving the impeller. The diffuser serves to gradually reduce the velocity of the air and convert its velocity energy to a higher pressure level. A volute casing surrounds the diffuser and repeats the procedure, collecting the air, reducing its velocity further, and recovering additional velocity energy. The stresses that are permissible in the impeller limit the maximum discharge pressure that may be obtained from this single-stage unit. For higher pressures, the answer again is multiple staging— with multiple impellers and passages to take air from each diffuser
Fig. 4. Cutaway view of single-stage, single-inlet centrifugal compressor with closed-type impeller. Electric drive motor is at left center.
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AIR COMPRESSORS
to the inlet of the succeeding stage. Axial compressors accelerate the air in a direction generally parallel to the rotating shaft. Each pair of rotating and stationary blades form a stage, but pressure rise per stage is small, so the usual axial compressor must have multiple stages to produce typical shop air. While centrifugal machines deliver practically constant pressure over a considerable range of capacities, axial compressors have a substantially constant flow delivery at variable pressures. Note that these characteristics also mean the flow from a centrifugal compressor must be greatly reduced to increase the pressure ratio, while an axial compressor can develop a substantial increase in pressure with a modest reduction in flow rate. In general, centrifugal machines have a wider stable range than axial compressors. Because of their more-or-less straight-line flow, axial machines tend to be smaller in diameter than centrifugal machines, and
Installation suggestions Industrial air compressors are rugged machines that will perform under adverse conditions, but it always is advisable to provide proper operating conditions to maximize reliability at minimum operating cost. Traditionally, compressors have been located in separate rooms to isolate their noise. Such locations are almost mandatory today to meet OSHA requirements. However, it still is important that the compressor room have an adequate foundation (particularly for reciprocating machines) as well as ample space so that the machine is easily accessible for inspection and maintenance. Stairways and catwalks can assist these procedures on larger compressors. The compressor room ideally should be clean and dry. Auxiliary equipment, piping, and wiring should be arranged so that it does not interfere with routine inspections. Instruments should be located within easy view of operators.
A/14
are apt to be longer. Their efficiency also tends to be slightly higher. Both types of dynamic compressors are subject to a stable maximum flow which is limited by an inlet choking phenomena. They also exhibit non-stable minimum flow which is affected by an outlet surge phenomena. Flow rates and pressure Regardless of basic design, the role of any industrial air compressor is to produce a rate of flow at a level of pressure satisfactory to meet the demands of the facility in which it will operate. For many industrial plants, this is a dynamic situation, with fluctuating loads every day and a high probability of increased demand — due to new equipment and operations — in the future. Obviously, an informed guess about future compressed-air requirements will be valuable when considering a new or augmented compressor installation. Assuming that most industrial facilities want compressed air in a pressure range from 100 to 125 psi, a good rule of thumb is that central, multi-stage compressors in the 100-hp size range will deliver about 5 scfm/hp. Delivery improves slightly as compressor size and horsepower rating increase, and conversely,decreases for smaller units. Another aspect of pressure: if only one section of a facility requires relatively high air pressure, consider a separate, high-pressure compressor to supply that area. Or install a booster compressor to increase shop pressure to the higher level in that section. Compressor selection Those who purchase plant air compressors today generally buy a completely packaged unit augmented by many other components besides the basic compressor. Evaluating these supporting components and their ability to help the compressor match output to operating conditions is, perhaps, more important than the choice of the type of compressor. When selecting a compressor package for a particular application, users should always consider every facet of the package, including: ● compressor capacity — pressure and
flow ratings, including performance at the low or high ends of the operating pressure range, and at the altitude where it will be installed ● provision for air- or water-cooling ● electric motor drive ● electric motor starter ● type of control ● serviceability ● maintainability ● type of enclosure and its function ● required auxiliary equipment ● operating noise level ● special accessories, such as unit or compressor prefilter; outdoor enclosure; cold or hot weather package; heat-recovery package; aftercooling, reheating, and filtration of discharge air; even skid or wheel mounting for portable units; and, of course ● cost and warranty. A word of caution on capacity ratings: be sure the compressor manufacturer provides data about the actual volume of air the package will deliver under your operating conditions. Some manufacturers rate their machines in delivered air, others in flange-to-flange or bare compressor performance. To compare bare-compressor ratings, they must be reduced to account for inlet filtration, cooling, and other losses. Such losses can range as high as 5 to 7%. Power cost At one time, many users of industrial compressed air considered it free — or at least so inexpensive that its cost was negligible. Steeply rising energy costs over the past twenty years have put that misconception to rest. Power cost is the largest single item in the operating cost of an air compressor. Over the typical long life of the machine, power costs will be large multiples of the initial cost of the compressor itself. Consequently, more efficient compressors can pay dividends in power savings over their lifetimes, even though their initial installed cost may be higher. In general, for flow requirements of 6000 cfm or less at typical shop-air pressures, reciprocating machines are the most efficient, followed by centrifugal (in relatively small sizes), rotaryscrew, and rotary-vane machines.
Liquid water in compressed air systems can escalate a plant’s operating expenses. Water can damage production machinery, resulting in downtime and spoiled product.
Air dryers W
hen components wear or become corroded as a result of moisture, they consume more compressed air — a sure sign of lost energy efficiency. When this wear or corrosion becomes great enough, components must be repaired or replaced — adding more costs to operating expense. The cost of replacement parts, labor, standby inventory, and downtime can have a devastating effect on the plant’s bottom line. Eliminating even one of the above by drying a system’s compressed air will offset the cost of installing and operating the equipment to do the job.
Types of dryers Dryers remove moisture from the air, which lowers its dew point — the temperature to which air can be cooled before water begins to condense from it. In broadest terms, there are four basic types of industrial compressed air dryers: deliquescent, regenerative desiccant, refrigeration, and membrane. Deliquescent dryers contain a chemical desiccant which absorbs moisture contained in the air, whether the moisture has already condensed or is still a vapor, Figure 1. The desiccant is consumed in the water-removal process and must be replenished periodically. Solution drained from these dryers contains both liquid water and the deliquescent chemical, so disposal may be a problem. Deliquescent dryers reduce the dew point of the air 15° to 25°F below the inlet air temperature. If the incoming air has a dew point of 90°F, it will leave a deliquescent dryer with dew point of about 65°F. Depending on operating conditions, some deliquescent dryers can produce dew points as low as -40°F; new deliquescent chemical may produce even lower dew points. Two important points: desiccant level should not be allowed to fall below that recommended by the dryer manufacturer and inlet temperature should be limited to 100°F or less to prevent excessive desiccant consumption. Regenerative desiccant dryers remove water from air by adsorbing it on
Dry air out Wet air in
Desiccant
Desiccant fill extends into vessel to prevent overfilling
Relief valve
Heater
,,, ,,, ,,, ,,, ,,,
Desiccant
Coalescing screen Collection chamber Drain solution
Fig. 1. Deliquescent chemical dryer takes moisture-laden air into collection chamber, passes it though support screens into desiccant chamber where part of the water vapor is removed.
the surface of a solid desiccant, usually silica gel, activated alumina, or molecular sieve. The desiccant does not react chemically with the water, so it need not be replenished. However, it must be dried or regenerated periodically. Heatless regenerative dryers use two identical chambers filled with desiccant. As wet air moves up through one chamber, a portion of the dry discharged air is diverted through the second chamber, reactivating its desiccant. The moistureladen purge air is vented to atmosphere. Some time later, air flow through the chambers is reversed. Heat regenerative dryers also use two identical chambers: one is shown in Figure 2. In this type, however, air flows through one chamber until its desiccant has adsorbed all the moisture it can hold. Then air flow is diverted to the second chamber. Heated outside air or an external source of heat (steam or electricity) then dries the desiccant in the first chamber. Because desiccant’s adsorption capacity decreases as temperature increases, the desiccant bed must be cooled from the temperature it reaches during regeneration. The regeneration cycle in these dryers usually lasts several hours — 75% heating and 25% cooling. Regenerative desiccant dryers can pro-
Pressure gauge Thermoswitch Heater in tube Heatconducting fins
Thermometer Tubular bed support and gas diffuser Perforated stainless steel support
Desiccant drain
Purge gas inlet
Fig. 2. Cutaway view of one tower of heated regenerative desiccant dryer shows electrically heated fins used to dry saturated chemical desiccant.
duce pressure dew points as low as -50°F. The type of desiccant used has a definite effect on the final dew point. Refrigeration dryers condense moisture from compressed air by cooling the air in heat exchangers chilled by refrigerants. These dryers can produce dew points of 35° to 50°F at system operating pressure. Many refrigeration dryers reheat the cooled air after it has been dried, usually by routing it through heat exchangers in contact with the hot incoming air. Reheating the cooled air prevents condensation from forming on the exterior of air lines downstream from the dryer and also precools incoming air. Standard refrigerating dryers should not be used where ambient temperature can drop below 40°F because lower temperature can freeze condensate, which blocks air passages and could damage the
1998/1999 Fluid Power Handbook & Directory
A/15
AIR DRYERS Wet air in
Wet air in
Precooler reheater
Dry air out
Water-to-air exchanger
Separator & drain
Heat exchanger Drain
Refrigeration unit
Liquid refrigerant
Fig. 3. Tube-in-tube refrigeration dryer uses refrigerant evaporator to cool wet, hot incoming air. Air-to-air precooler, top, allows heat from incoming air to warm cool, dry outgoing air. This precooling/post-warming process boosts overall dryer efficiency. Separator collects moisture condensed from the air, drain discharges it.
dryer’s evaporator. Dryers may be equipped with heat tracing packages for operating in ambient temperatures as low as 50°F. Refrigeration dryer types Refrigeration dryers can be further classified into three types: Tube-in-tube refrigeration dryers, Figure 3, operate by cooling a mass of aluminum granules or bronze ribbon that in turn cools the compressed air. As the tube-to-tube refrigeration dryer cycles, a thermometer in the granule mass senses its temperature. As the temperature rises, a switch turns on the refrigeration unit. When the temperature drops to a cut-off point, refrigeration stops. These dryers are designed to produce dew points of 35° or 50°F. Water-chiller refrigeration dryers, Figure 4, use a mass of water for cooling. An extra heat exchanger is necessary to maintain chilled water flow through the condenser, as is a water pump. Dew points can be between 40° and 50°F. Water-chiller dryers cycle as they operate. Direct-expansion refrigeration dryers, Figure 5, use a refrigerant-to-air cooling process to produce dew points that are 35°F below standard operating conditions (100°F temperature at compressor inlet, 100 psig, 100°F ambient — from NFPA standard). No recovery period is necessary, so direct-expansion refrigeration dryers run continuously. The cost difference between cycling and continuous operation is difficult to calculate. Differences in power consumption for frac-
A/16
Dry air out
Water chiller
Pump
Refrigerant compressor
Condenser
Fig. 4. Water-chilled refrigeration dryers use three heat exchangers. Precooler/heater performs same function as in tube-in-tube dryer; second water-to-air heat exchanger pumps chilled water through the exchanger counter to air flow; third heat exchanger uses refrigerant to chill water recirculating from second to third.
tional horsepower dryers are negligible. The cost of starting an electric motor larger than 1 hp at the onset of every cycle becomes greater than letting it run continuously — but unloaded — through much of the cycle. Typically a refrigerant compressor will cycle once or twice every two or three minutes. Thus, in-rush loads are likely to be higher than if the motor runs continuously. Membrane -type dryers Membrane-type dryers are gas-separation devices. They consist of permeable membrane surfaces that have been specially tuned to block nitrogen and oxygen molecules (air), but allow water vapor molecules to pass right through. They work as your lungs do, venting water vapor each time you exhale. Typically this membrane material is formed into bundles of thousands of individual fibers from one end of the dryer to the other. Water vapor escapes through the walls of the fiber to a sweep chamber from where it is continually vented to atmosphere as a gas. A fraction of the dried air is routed through the sweep chamber to continuously purge and exhaust moisture vapor. Industrial-grade membranes can be used for years to dry air continuously. They respond spontaneously to any change in inlet conditions. They perform at temperature between -40° and 150°F (ambient or inlet), and handle pressures from about 60 to 300 psig. They will deliver a consistent outlet dew point reduction anywhere between these extremes.
The inlet flow rate and pressure determine the outlet dew point depression. In other words, membrane air dryers deliver a consistent layer of drying protection that follows the rise or fall of the inlet dew point temperature, and can easily be sized to follow the ISA recommended 20°F pressure dew point suppression below ambient. Outlet pressure dew points can also be selected as low as -50°F. Currently, flow capacities to a relatively low 50 scfm are available, but modules can be installed in parallel for higher flows. Prefilters mounted immediately upstream from the dryer keep out liquids and solids to allow an almost unlimited service life. Because water vapor passes right through the membrane material, it does not accumulate there, so membranes do not become saturated and never need to be refrigerated. Membranes have no moving parts to wear out. They are non-electric and suitable for most hazardous locations. Wet air in Refrigerant-toair exchanger
Dry air out
Separator & drain
Compressor Condenser
Fig. 5. Direct-expansion refrigeration dryer uses two heat exchangers: air-to-air precooler/heater and refrigerant-to-air evaporator.
AIR DRYERS
100 90
30 ig
ig
s 0p
0
10
60 0
40
sig
p 00
5
20
00
ps
70 60
ig ps
50 40 30
ig ps
20 10
sig
p
0 -10
g
si 5p
ig
s 0p
40 ig ps
-20 -30
ic
2
er sph
-40
o
Atm
-50 -60 -70 -80 -90 -100
-90
-80
-70
-60
-50 -40 -30 -20 -10 0 10 Given dew point at atmospheric pressure
20
30
40
0.001 0.003 0.006 0.014 0.021 0.044 0.088 0.166 0.285 0.481 0.776 1.23 1.93 2.84 Moisture content, grains/ft 3
2
Importance of dew point As pointed out earlier, wet air adds to plant operating expenses through: ● repair parts ● repair labor ● product damage, and ● production downtime. The economic advantages of reducing or eliminating these detriments of moisture build a strong case for installing a dryer. Once the decision to install a dryer has been reached two questions arise: how dry must the air be and what type of dryer should be used. The most important criterion in choosing an air dryer is the pressure dew point that it must produce. The required dew point of an air system determines how dry the air must be and to a great extent, which type of dryer to use. Dew point varies with pressure. For example: the dew point conversion chart, Figure 6, shows that air at atmospheric pressure with a dew point of -12°F has a pressure dew point of 35°F at 100 psig. Dryer manufacturers may specify the dew point that a particular model can attain at atmospheric pressure or at a typical system pressure, such as 100 psig. If performance is specified at atmospheric pressure, use a chart like Figure 6 to find what the minimum dew point will be at the system’s operating pressure. The required dew point varies with each application. If preventing condensation in compressed air lines is the main
ps
20
Pressure, psi
They require no RF shielding or protection. They use no refrigerant gas or potentially dusty desiccants. They vent gas, not condensate so they can’t freeze up and don’t require heat tracing at very low ambient temperatures. They make no noise. They can be mounted in any orientation. Their lowmass components are inherently vibration-resistant. Because they are static, inert devices, they never need service or adjustment and don’t require monitoring devices. Made of plastic and aluminum, they do not rust or corrode and don’t need painting. They have almost no pressurized volume, so most pressure code restrictions do not apply. Note: membrane gas separators will remove other gasses too. Membrane-type compressed air dryers can reduce outlet oxygen concentrations and must not be used for breathing air.
00
80
ig
0 10
Dew point at elevated pressure
00
ig ps
6
11
26
40 83 167 316 543 916 1478 2344 3678 5412 Moisture content, ppm (by weight)
Fig. 6. Dew point conversion chart assists in determination of dew point of air at variety of pressures. Moisture-content scales chart quantity of moisture contained in atmospheric air at indicated dew points.
concern, then the lowest ambient temperature to which air lines will be exposed will be the controlling factor. However, for some applications, dew point requirements will be more severe, possibly as low as -4°F at line pressure. An example might be the air used for spraying a powdery substance. Even the slightest trace of moisture in such air could condense and cause particles to stick together. If all the compressed air will be used inside a building where temperature is maintained at a stable level, then the required dew point can be fixed within a few degrees. But if some or all of the compressed air is subjected to outdoor temperature variations, the required dew point can change from day to day, or even hour to hour. Do not be too aggressive and estimate and unjustifiable margin for error. Stating a dew point much lower than that actually required wastes money. A rule-of-thumb margin for error is about 20°F maximum. Extremely low dew points may be required at only a few isolated locations. If this is the case, consider using individual dryers at each low-temperature point of use to attain these low dew points. A less expensive dryer to dry the air to less-strin-
gent requirements can then be installed for the rest of the air system. Evaluating flow capacity An air dryer not only must dry compressed air to the required dew point, but also must be able to handle the required air flow without causing excessive pressure drop. Flow capacity of a dryer depends on: ● operating pressure ● inlet air temperature ● ambient air or cooling water temperature, and ● required dew point. When any of the above conditions changes, flow capacity of the dryer also changes. Dryer manufacturers can supply performance curves that show the relationship of their dryer’s flow capacity to these four factors. Table 1 compares dew point, flow capacity, inlet temperature, and ambient temperature for several types of dryers. Evaluating characteristics of the different types of dryers will help indicate which is best for a particular application. This is where cost finally can be considered. Purchase price of the dryer is only one factor to evaluate when choosing an air dryer. A deliquescent chemical dryer, for example, has a relatively low initial cost, but its chemical
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AIR DRYERS
must be replaced periodically, adding to the operating cost. This cost is offset somewhat because the deliquescent chemical dryer requires no external power source. Other dryer types may cost more initially, but have lower operating costs because they can run for long periods with little or no maintenance require. It should be clear then, that cost analysis should be conducted based on manufacturers specifications as they relate to an individual application’s physical and economical requirements. Installation and maintenance Location can affect how well an air dryer performs. The site for an air-cooled dryer should be well ventilated so heat can be carried away and readily accessible to aid maintenance. The maximum ambient temperature for a refrigeration dryer is about 100° to 100°F. Higher temperature prevent the dryer from exchanging heat with its surroundings and keep it from operating properly. Dryers with water-cooled condensers can tolerate higher ambient temperature because they transfer heat to the cooling water instead of to the surrounding environment. Refrigerant dryers, whether air- or water-cooled, should not be exposed to ambient temperature below 0°F unless optional low-ambient-temperature controls are installed. If a deliquescent dryer is used in a central compressed air system, bypass piping should be installed around the dryer to maintain air supply whenever the dryer is taken off line to add desiccant. There should also be no set of operating conditions that permit system pressure to drop low enough to allow high, turbulent air flow through the dryer that might carry chemicals into system air lines. Refrigeration and deliquescent dryers should be drained regularly, depending on the volume of liquid accumulated. Most refrigeration dryers have automatic drains, as least as an option. Most, if not all, refrigeration dryers require a prefilter to remove oil and dirt from incoming air that would otherwise coat the inside of the dryer’s heat exchanger and lower its heat-transfer capability. Regenerative desiccant dryers require equipment that removes compressor-oil carryover that can coat the desiccant and render it useless. These dryers also require afterfilters to prevent downstream migration of desiccant particles.
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Dewpoint needs and applications Refrigeration
Dual tower regenerative Pressure dewpoint
Direct Tube expan- -toHeated Heatless sion tube
Application
Deliquescent Combination units desiccant Water RefrigHygro- chilled erated Water scopic desiccant desiccant chiller Urea salts (Urea) (Urea)
General plant air for tools and fixtures when ambient temp remains at 707F or higher
U
U
U
U
U
U
507F
Same as above, but ambient temp as low as 607F
U
U
U
U
U
U
457F
Same as above, plus general instrumentation & controls, air conveying, air gages, food & chemical processing and ambient temp as low as 557F
U
U
U
U
U
U
1
U
U
U
U
U
2
U
607F
357F
Same as above, but ambient temp as low as 457F
U
U
207F
Same as above, but ambient temp (air lines) to –207F to prevent freeze-ups
U
U
07F and lower
Some aerospace applications and other exotic requirements for exceptionally dry air
U
U
1
2
With 607F cooling water
U
U
U
U
U
U
With 507F cooling water
Inlet temperature and air flow demand 0-1000 SCFM, inlet 907F maximum, intermittent load, 40-80 hours per week
U
U
U
U
U
U
U
U
U
Same as above but, but 70% capacity or more constant air flow
U
U
U
U
U
U
U
U
U
U
U
to 1507F
0-1000 SCFM, inlet 917F or higher, intermittent load, 40-80 hours per week
U U
Same as 1000 SCFM, inlet 907F maximum, intermittent load, 40-80 hours per week
U
U
U
Same as above, but 70% capacity or more constant air flow
U
U
U
3
1107F to 1307F
U
U
4
U
U
U
3.4
U
U U
cooling water to 707F 4 Temperatures to 1207F
Ambient temperatures–limitations Ambient temperature to 1007F Ambient temperature to 1107F Ambient temperature to 1307F 5
U
U
U
U
U
U
(1207F)
Water cooled only
(1207F)
U
U
U
U 5
U
U
U
U
U
U
U
U
5
U
Inlet temperature–limitations
Inlet temperature to 907F
U
U
U
U
U
U
U
U
U
Inlet temperature to 1007F
U
U
U
U
U
U
U
U
U
Inlet temperature to 1107F
U
U
U
U
U
U
U
U
U
U
U
U
U
Inlet temperature to 1507F Inlet temperature to 2007F
5
U
Table 1. Guide for selecting compressed-air dryers outlines major application factors which will affect specification of an approximate unit. (Note that relatively new membrane dryers do not appear on this chart.) Check marks indicate where dryers will perform.
Reliability is one of the strong suits of pneumatic systems — and proper filtration is a key to reliability and longevity. Here are some basics that explain the whys and hows.
Air filters I
t is estimated that the air we breathe contains about 1/2-million particles of dust per cubic inch. This means that there are billions of contaminating particles in the four-million cubic inches of air the average 10-hp compressor ingests per hour. At high concentration and high speed, the larger of these particles can be extremely harmful to any compressed air system by blocking orifices, eroding components, and clogging clearances between moving parts. In addition, when ambient air is drawn into a compressor, its relative humidity can, depending on weather conditions, reach 100%. As air is compressed and cooled, some water vapor condenses out as free water, and even with an aftercooler, part of this liquid is swept downstream into the air system. This frequently results in rusted pneumatic tools and components, destroyed lubricants, and frozen air lines during
Wet air enters
,
Dry air exits
Cleanable porous bronze element
Polycarbonate bowl Metal guard
Multi-stage baffle spins air to separate moisture
Quiet zone eliminates carryover
Fig. 1. Cutaway drawing shows internal workings of typical airline filter. Baffles direct air into circular motion where centrifugal action separates contaminants, which collect in quiet zone, from where they can be removed.
low-temperature periods. Other types of foreign matter in air lines include: ● construction and assembly debris ● carryover oil from the compressor ● impurities generated within the compressed air line — such as wear particles, pipe scale, rust, and ● contaminants ingested into the air system during maintenance or through leakage passages. All contaminants large enough to cause air system problems should be removed by filtration. Therefore, the first step in filter selection is to determine the filtration requirements of the most critical components in the system. Contamination particle size is measured in micrometers — abbreviated mm. A mm is one-millionth of a meter, or 0.000039 in. Particle-removing filter elements are rated according to the particle size they will trap. For most industrial applications, filter elements rated at about 40 to 60 mm are adequate. When necessary, filtration down to 20 or even 5 mm or finer can be provided. Remember, however, that finer filtration increases the pressure drop through the element. As micrometersize rating varies, so does the size and type of filter. Most oils entrained in a compressed air stream are in the form of tiny mist or aerosol droplets that can pass through standard industrial filter elements. A coalescing-type filter can be used to remove these aerosols. The sub-micrometer oil particles that escape a coalescing filter should have no detrimental effect on industrial pneumatic components. However, if these particles must be removed for applications such as food processing or breathing air, a third oilremoval-type element is available. Filter operation When pressurized air enters a typical filter bowl, Figure 1, a curved inlet and deflector direct the incoming air in a
downward whirling pattern. Centrifugal force hurls the larger solid and liquid water particles outward, where they collect on the inner surface of the filter bowl. The particles spiral down past a baffle into a quiet chamber. The baffle prevents turbulent air in the upper bowl from re-entraining liquid contaminants and carrying them downstream. The dry, cleaner air then follows a convoluted path through the filter element, where finer solid particles are filtered out. Finally, filtered air passes up the center of the element and out the discharge port. Thus, air line filters remove impurities in two operations; dynamically by centrifugal force, throwing out heavier particles and entrained water; and statically through the filter element, straining out smaller particles. High-efficiency coalescing filters operate on a somewhat different principle from air line filters. The key difference is in the element, where a fiber network is narrowly spaced to trap smaller contaminants. The special fibers hold any liquid particle that contacts them. Pre-filtered air enters the cylindrical element at the center, Figure 2. As air flows through the element, particles are captured by three different mechanisms: direct interception as particles impinge on the fibers, inertial impaction as particles are thrown against fibers by the turbulent air stream, and diffusion as smaller particles vibrate and collide with fibers and other particles. As a result, coalescing elements can capture particles smaller than the nominal size of the flow passages through the element. Collected liquid migrates to the crossing points of the fibers where larger drops form or coalesce. Pressure differential through the element then forces these drops to the downstream surface of the element, where they gravitate downward to the sump. The filtered air then exits through the outlet port.
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A I R L I N E F I LT E R S
Air inlet Body
Air outlet Center mounting stud
Coalescing element
Bowl Collected liquid
Drain
Fig. 2. Cut-away drawing shows internal workings of coalescing-type filter. A narrowly spaced fiber network allows capturing contaminants smaller than the nominal size of flow passages through the element.
Regardless of the type, it is important to inspect or change filter elements on a regular basis. Saturated elements restrict air flow, causing excessive pressure drop across the filter, which degrades pneumatic system performance and efficiency. Most manufacturers offer an option that automatically provides a visual indication or electronic signal when the filter needs attention. Typical construction Most pneumatic filters consist of two basic elements: a die-cast body, into which the inlet and outlet piping is connected, and a sealed, removable bowl, which contains collected contaminants. The bowl is fitted with a drain mechanism to remove liquids before they rise to the baffle level. The drain system usually operates while the filter is pressurized, so the unit must be exhausted to remove the bowl for cleaning and element service. However, the piping need not be disturbed. Generally, a transparent plastic bowl is used for convenience because it provides easy visual inspection of the sump level. However, a hostile environment, higher pressure, or
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higher temperature may require a metal bowl for safety. The most common plastic used for bowls is polycarbonate, a highstrength plastic offering crystal-clear visibility. This material performs well at air pressures below 150 psi and temperatures between 408 and 1208 F, but polycarbonate can be attacked by several chemicals. A metal bowl guard with many viewing holes adds a degree of safety. For pressures between 150 and 250 psi, and temperatures between 408 and 1508 F, all-metal bowls are recommended. Metal bowls provide superior protection when installed in an environment containing chemicals that are incompatible with polycarbonate. A metal bowl can be fitted with a sight gage as long as no chemicals are present that would attack the bowl sight glass. Sight gages should not be used when pressures may exceed 300 psi or temperatures below -408 F or above 1808 F will be encountered. Thus, the environment determines the choice of bowl. Polycarbonates offer great strength and visibility, but can be attacked by certain chemicals. Metal bowls offer strength at higher pressures and temperatures, and provide superior protection when installed in an environment containing chemicals that are incompatible with polycarbonate. Drains There are many manual and automatic drains to remove collected liquids from filter bowls. Where size is a consideration, miniature filters may use a tire valve-type drain with an external push-pin to act as an actuator. Larger filters often have metal or plastic drain cocks actuated by wing nuts or knobs. Some manufacturers equip their manual drains with a flexure tube that opens the drain whenever the tube is pushed away from its normal vertical position. The automatic filter drain is a valuable option when filter location makes servicing difficult or dangerous, or when filters are hidden and may be inadvertently overlooked. If low temperatures are encountered during long shutdown periods, water may condense in air lines. At startup, slugs of condensed water can be pushed into downstream
Fig. 3. Typical airline filter with integral automatic float-type drain. As water accumulates at bottom of bowl, float rises to open drain valve, which allows condensate to escape.
filter bowls, overloading them immediately. Drains can be an integral part of the filter bowl, Figure 3, or can be located externally with connecting piping. The two basic types of automatic drains are pulse-operated for intermittent air flow and float-operated for continuous flow. The less-expensive pulseoperated drain usually has the drain valve linked mechanically to a piston or diaphragm, which fits across the bowl above the surface of the sump. When fluid pressure is equal on both sides of the piston, the larger top area holds the drain valve seated. As flow starts through the filter, pressure above the piston temporarily drops, and the piston unseats the drain valve to empty the sump. Pressure above the piston then reseats the drain valve. Flow through the filter must be intermittent for the pulseoperated drain to work. In float-operated drains, accumulated liquids in the sump raise a float, which either opens the drain mechanically or pilots a valve which opens the drain. Bowl pressure forces the liquid out rapidly: the float subsequently lowers, and the drain closes.
In this day of microchips and personal computers, air logic can still provide an effective, efficient, and inexpensive means of machine control for certain applications.
Air logic E
lectrical and electronic devices normally control fluid power circuits. Relay logic circuits, programmable controllers, or computers are common control methods. Another way to control fluid power systems is with air logic. Air logic controls perform any function normally handled by relays, pressure or vacuum switches, time delays, counters, and limit switches.
A
A
B
B Logic symbol
Fig. 1 Passive AND element
A
A B B
Logic symbol
Fig. 2 Active AND element
The circuitry is similar, but compressed air is the control medium instead of electrical current. High dust or moisture environments are excellent places for air logic controls. Practically no danger from explosion or electrical shock is possible even in these atmospheres. Water can splash on the controls with no effect on the operation. If there is danger of explosion, air controls cannot ignite the materials involved. Another place to use air logic is on machines that have cylinders or fluid motors, but no type of electrical device. It requires two services on a machine powered by air and controlled electrically. Another reason for this might be where electrical and mechanical maintenance come under different labor ANSI symbol grades. Since there are no electrical devices involved, one craft works on the circuit and machine parts. Disadvantages of using air logic control are lack of understanding of how the components work and how to read the schematic drawing. If an air controlled machine fails, very few people can work troubleshoot it. Also, air logic with long control lines will have a noticeANSI symbol ably slower cycle. Control lines over ten to fifteen feet, fill and exhaust slowly when compared to electrical
signals. In addition, air quality must be above average to ensure long life. Air logic controls are basically miniaturized 3-way and 4-way air valves. The actions of the valves give on or off functions like relays or switches, plus exhausting of the spent signal. The symbols used for air logic are similar to electronic symbols. Some manufacturers use modified electrical symbols and ladder diagrams to show circuitry. Basic logic elements Following are explanations of the basic logic components using standard ANSI logic symbols and ISO graphic symbols of a comparable directional control valve. Figures 1 and 2 show two types of AND elements. An AND element must receive two inputs before there is an output. This ensures that two functions are complete before there is a command to continue the cycle. In other words, input A and that input B must both be present before an output action occurs. When using more than two inputs, connect AND elements in series. The first AND receives signals 1 and 2, and the output of this element hooks to an input of the second AND. The other input of the second AND receives a third signal, making three inputs necessary before an output action occurs. Some manufacturers supply passive and active types of elements but designate the passive AND simply AND, whereas the active AND is designated YES. The difference in the elements is, the passive AND element uses the lower of the two inlet ports as an output. In contrast, the active AND element has two inputs to achieve an output, but the designer has the choice of which input goes to the output. Using this feature can amplify a weak signal.
1998/99 Fluid Power Handbook & Directory
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AIR LOGIC
A
A
A
N
B Supply
B Supply B Logic symbol
ANSI symbol
Fig. 3 OR element
Logic symbol
ANSI symbol
Fig. 4 NOT element
The weak signal pilots the valve open while the through signal comes from a full pressure supply. The YES is an active element. Figure 3 shows The symbol for an OR element. A shuttle valve serves the same purpose as an OR element. Either input to an OR element gives an output. A pilot signal from two different sources can pass through to start the next function. Another way of saying this is, signal A or that signal B gives an output. An OR element differs from an in-line tee because an OR passes either input to the output but does not allow the inputs to pass to each other. Stack OR elements to allow for more than two inputs. Use an extra OR for each input after the first two signals. Figure 4 shows the symbol for a NOT element. A NOT logic element is a normally open three way valve. An input signal or pressure to SUPPLY will go through the valve until there is a
pilot signal at port (A). Pressurizing port (A) blocks supply and exhausts the output signal to atmosphere. NOT elements will block a signal or supply as long as there is pilot pressure on the (A) port. The NOT always returns to a normally open condition without a pilot signal. Use a NOT element to replace a limit switch to indicate a cylinder is at the end of stroke. Pressure from the cylinder port goes to port (A) of the NOT, holding it closed. As the cylinder moves to the work, pressure is maintained because of the meter out flow control. When the cylinder contacts the work, the signal on port (A) drops, the NOT opens and sends a signal to start the next operation. The cylinder can stop at any position and the NOT’s output signal will indicate its non movement. This will
always happen whether the cylinder stopped where it should be or is stalled by some other means. Since this is the case, take care using a NOT to replace a limit switch. In contrast, this feature can be advantageous when clamping different sized parts. Use a NOT element for applications where different work locations stop the cylinder. Most manufacturers supply a differCAUTION! Any pressure-control valve only shows a pressure buildup. When a positive location must be made, use limit valves. ent pilot ratio for a NOT element used as a limit switch. The valve function is the same but the pressure that it shifts at is much lower. Some manufacturers
S
S
R
FF
MEM
Supply
Supply Supply
Supply R
Logic symbol
Fig. 5 FLIP-FLOP element
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S
R ANSI symbol
Logic symbol
Fig. 6 MEMORY element
ANSI symbol
AIR LOGIC
A
A
A
A
Logic symbol
ANSI symbol
Fig. 7 ONE SHOT element
make a special NOT element that mounts directly to a cylinder port. A port mounted meter out flow control used in conjunction with this special NOT makes a compact installation. Figure 5 shows a FLIP FLOP element symbol. A FLIP FLOP is a double piloted 5-way valve that sends SUPPLY air to either outlet port with a signal at pilot ports (S) or (R). SUPPLY can be system pressure or air from another logic element. The main use for a FLIP FLOP is to eliminate the first pilot signal to a directional control valve. This allows a second signal on the directional valves opposite pilot port to shift it back. FLIP FLOPs, sometimes called MEMORY elements, stay in the last shifted position even with no air supply. Whether the signal maintains or drops out, output from the FLIP FLOP stays the same. The (S) and (R) signals stand for
Logic symbol
ANSI symbol
Fig. 8 ON DELAY TIMER element
SET and RESET. The SET signal shifts the FLIP FLOP for a function and whether the signal continues or not, the element stays shifted. The RESET signal puts the FLIP FLOP back to its original position until the next cycle. Another use for a FLIP FLOP is to set up a new cycle allowing the operator to momentarily push the start buttons. Use this same FLIP FLOP to eliminate unwanted signals and set up the circuit for cycle completion as required. Figure 6 shows another valve actually called a MEMORY element. A MEMORY element is a normally closed three way valve with a built in shuttle valve. The shuttle valve uses the MEMORY’s output air to hold it shifted once it receives a SET signal. A momentary SET signal gives continuous pilot output. A RESET signal shifts the MEMORY element to normally closed and exhaust’s output air. Also, turning SUPPLY pressure off returns a A MEMORY element to the start position. In air logic control there are A three different Supply types of time delays. Fixed or adjustable time Supply delays are common in both normally closed Logic symbol ANSI symbol and normally open configuraFig. 9 OFF DELAY TIMER TIMER element tions. Some
time delays use an orifice and accumulator chamber for delays up to one minute. Some manufacturers use air actuated diaphragms and orifices that eliminate system pressure fluctuation inaccuracies. Figure 7 shows the symbol for a ONE SHOT timer, sometime called an IMPULSE TIMER. A ONE SHOT timer takes a signal and passes it on to the circuit. At the same time the input signal is going through an orifice to an accumulator tank. The setting of the orifice and size of the accumulator gives a certain time delay before the normally open 3-way valve closes. After a ONE SHOT times out and closes, it remains closed as long as it has an input signal. Figure 7 shows an adjustable time delay before losing the output. Leaving off the sloping arrow in the symbols makes it a preset time delay. Times range from 1⁄2 second to two or more seconds on valves with preset time delays. Many circuits use ONE SHOTs to eliminate opposing signals. When a valve receives a signal to extend a cylinder, it resists a return pilot signal to itself until loss of the first pilot. Using a ONE SHOT element drops the extend signal shortly after initiation. However, when the short duration signal meets a hard to shift valve, the time may not always be long enough to move the valve spool. When the valve does not have time to shift, the cycle stalls. For best results use a FLIP FLOP to drop unwanted signals after it performs its task. Figure 8 shows an adjustable, normally closed TIME-ON, time delay sym-
1998/99 Fluid Power Handbook & Directory
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AIR LOGIC
A
A
N
B
Supply Supply B Logic symbol
ANSI symbol
Fig. 10 NAND element
A
A
N
B
Supply Supply B Logic symbol
ANSI symbol
Fig. 11 NOR element
bol. A TIME-ON delay passes a signal through the element after timing stops. A TIME-ON delay is a preset fixed timer without the sloping arrow. Most anti-tie down circuits use a fixed time delay. This forces the operator to actuate both palm buttons concurrently. The symbol in Figure 8 shows an input (A) going to the blocked port of a 3-way directional valve. Signal (A) also goes to a meter in flow control to fill an accumulator. After the accumulator fills, pilot pressure shifts the 3-way valve, which allows air to pass on to the next operation. As long as the input signal stays on, the time delay stays open.
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Some brands of TIME-ON delays use shop air to the normally closed port (A) of the 3-way valve while the signal to the timing section comes from another logic element or limit valve. This allows a strong passing signal to travel long distances or to quickly shift several other logic elements. With a built-in accumulator tank the time delay length is usually under one to one-and-one-half minutes. With added external accumulators, time delays up to five minutes are possible. The repeatability of long time delays using accumulators is poor. Often diaphragm type timers go to three minutes with good repeatability.
With a normally open 3-way in place of a normally closed 3-way, the delay is TIME-OFF. Figure 9 shows the symbol for a TIME-OFF delay timer. A continuous input to the SUPPLY gives an output until a set time after receiving a signal at (A). When (A) receives a signal, the time delay starts and continues timing. When the accumulator fills, it closes the normally open 3-way valve and exhausts the signal. As before, a preset, nonadjustable time delay is available. TIME-ON and TIME-OFF delays often are identical in appearance. The part number may be the only way to tell these units apart. To get different functions, connect air logic elements together like the examples in Figures 10 and11. These are two common pairs that might be familiar to anyone using air logic. A NAND element, Figure 10, uses an AND to signal a NOT. As long as there are not signals at (A) and (B) air passes. If signals (A) and (B) are present the NOT closes and exhaust’s the output signal. The term NAND means, not this and this. A NOR element, Figure 11, uses an OR to signal a NOT. As long as there is NOT a signal at (A) or (B) air passes through the NOT. If a signal is present at either (A) or (B), the NOT closes and exhaust’s the output signal. The term NOR means, not this or this. Some other commonly used air logic elements are: ● Amplifiers —Detect a low pressure signal (down to three inches water column) and sends it on as an 80 PSI signal. ● Pressure or vacuum sequence —Elements that shift after reaching a set pressure or vacuum. ● Programmable controllers — Combination elements used to design complex circuits with minimum knowledge of circuit design. ● Air-operated indicators — Indicators to show circuit condition and/or function. Several colors are available but none emit light.
Airline lubricators
Valves, cylinders, and air motors can be lubricated by injecting oil into the air stream powering them. Properly locating the lubricator in the circuit is important to ensure proper lubrication.
M
any pneumatic system air passes through the venturi Feed-rate components and almost where it mixes with metered oil Fill plug adjustment all pneumatic tools redroplets. Under higher flow conVenturi Sight glass quire oil lubrication in the airditions, the spring-loaded bypass supply line for proper operation valve opens and the excess flow and long service life. This lubribypasses the venturi, then blends Inlet Outlet cant is carried by the air stream. with the lubricated air at a downToo little oil can allow excessive stream point. A manual adjustBypass wear and premature failure. Too ment in the housing sets the oil valve much oil is wasteful and can bedrip-rate into the air stream; a come a contaminant in the ambisight glass enables the operator to ent area, particularly when carmonitor that rate. Capillary tube ried over with air exhaust. Fill plugs at the lubricator top Intermittent lubrication may be provide access to refill the reserthe worst condition of all because voir with oil. The bowl can be rethe oil film can dry out to form moved for cleaning. sludges and varnishes on internal surfaces during periods when the Typical construction system is not operating. Bowls are available in polycarAirline lubricators meter oil bonate and metal — subject to the from a reservoir into the moving same constraints discussed under air stream. In popular terminol- Pressure in lubricator bowl forces oil up capillary tube to filter bowls. Transparent polycarogy, if the oil droplets carried low-pressure venturi section, where fast-moving air bonate simplifies inspection of downstream are relatively large, stream breaks oil into droplets. Bypass valve opens at the oil level and checking for dirt they are termed a fog; smaller high flows. Air streams reunite downstream from venturi. and liquid condensate in the oil. droplets form a mist. Fog-type luNote that the system must be exbricators serve best when the down- up the oil into droplets. hausted before removing the bowl. In a typical lubricator, pictured stream flow path is straight and short. In some models, the system also Mist-type lubricators are used when oil above, filtered and regulated air enters must be exhausted before opening the must be carried for longer distances or the lubricator housing and is channeled fill plug to recharge the lubricator. in either of two directions depending Other designs automatically bypass the through an obstructed path. on flow rate. At low flow rates, all the air during refilling. Lubricator operation Some lubricators, as in the drawing, Almost all lubricator designs include are equipped with a needle valve in the The Basic Circuits pages in this a high-velocity venturi section in the tube to set oil feed rate. Others have a Engineering Section of the air flow path which creates a low preswick at the end of the capillary tube Handbook include samples of sure area to draw oil from the reservoir from the reservoir, and feed rate is conair lubrication circuits. through a capillary tube to the point of trolled by adjusting the distance that injection. There, the air stream breaks the wick extends into the venturi. 1998/1999 Fluid Power Handbook & Directory
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Air motors A
ir motors are used to produce continuous rotary power from a compressed air system. They boast a number of advantages over electric motors: ● because they do not require electrical power, air motors can be used in volatile atmospheres ● they generally have a higher power density, so a smaller air motor can deliver the same power as its electric counterpart ● unlike electric motors, many air motors can operate without the need for auxiliary speed reducers ● overloads that exceed stall torque generally cause no harm to air motors. With electric motors, overloads can trip circuit breakers, so an operator must reset them before restarting equipment ● motor speed can be regulated through simple flow-control valves instead of expensive and complicated electronic speed controls ● motor torque can be varied simply by regulating pressure ● air motors do not need magnetic starters, overload protection, or the host
of other support components required by electric motors, and ● air motors generate much less heat than electric motors. As one would expect, electric motors do possess some advantages over air motors: ● if no convenient source of compressed air exists for an application, the cost of an air motor and its associated support equipment (motor-driven compressor, controls, filters, valves, etc.) will exceed that of an electric motor and its support equipment ● air motors consume relatively expensive compressed air, so the cost of operating them will probably be greater than that of operating electric motors ● even though electronic speed controls escalate the cost of electric motor drives, they control speed more accurately (within ±1% of desired speed) than air motor controls do, and ● air motors operated directly from a plant air system are susceptible to speed and torque variations if flow and pressure fluctuate.
Common designs of air motors include rotary vane, axial piston, radial piston, gerotor, turbine, V-type, and diaphragm. Rotary vane, axial- and radial-piston, and gerotor air motors are most commonly used for industrial applications. These designs operate with highest efficiency and longevity from lubricated air. Of course, specific designs are available for applications where lubricated air proves undesirable. Turbine motors are used where very high speed but low starting torque are required. V-type and diaphragm air motors are used primarily for special applications and will not be covered here. Piston motors Piston air motors are used in applications requiring high power, high starting torque, and accurate speed control at low speeds. They have either two, three, four, five, or six cylinders arranged either axially or radially within a housing. Output torque is developed by pressure acting on pistons that reciprocate within the cylinders.
Fig. 1. Cut-away view of axial-piston air motor. High starting torque is a key benefit of both axial- and radial-piston air motors.
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AIR MOTORS
Air in
Air ports Hole(s) to bottom of rotor slot
Air out
Vane
Body Kidney shaped port (both end plates)
Rotor
Fig. 2. Diagram of vane-type air motor operation. Air flows through the vane motor body to end plates and then to open kidneyshaped ports, where it enters rotor slots and pushes vanes against the housing. Air then passes into the main motor chamber through holes drilled in the rotor, where air directly pressurizes exposed portions of vanes to turn the rotor.
Fig. 3. Gerotor-type air motor delivers high torque at low speed. The gerotor element is shown at right.
Motors with four or more cylinders provide relatively smooth torque at a given operating speed, because power pulses overlap. This is because two or more pistons are undergoing a power stroke at any time within a revolution. Motors designed with overlapping power flow and accurate balancing are vibration-free at all speeds — particularly at the low end. Power developed by a piston motor depends on the inlet pressure, the number of pistons, and piston area, stroke, and speed. At any given inlet pressure, more power can be obtained from a motor that runs at a higher speed, has a larger piston diameter, more pistons, or longer stroke. Speed-limiting factors are the inertia of the moving parts (which has a greater effect in radial- than in axial-piston motors) and the design of the valve that controls inlet and exhaust to the pistons.
Radial- and axial-piston motors have one significant limitation: they are internally lubricated, so oil and grease supplies must be checked periodically and replenished. They must be mounted in a horizontal position to provide proper lubrication to bearing areas. However, at least one manufacturer offers a radial-piston motor with the shaft vertically-down as a standard configuration. Other mounting positions from any manufacturer require special lubrication configurations. Radial-piston motors feature robust, oil-lubricated construction and are wellsuited to continuous operation. They have the highest starting torque of any air motor and are particularly beneficial for applications involving high starting loads. Overlapping power impulses provide smooth torque in both forward and reverse directions. Sizes range to about 35 hp for speeds to 4500 rpm. Axial-piston motors, Figure 1, are more compact than radial-piston motors, making them ideal for mounting in close quarters. Their design is more complex and costly than vane motors, and they are grease lubricated. However, axial-piston motors run smoother and deliver maximum power at much lower speeds than vane motors can. Smaller and lighter than electrical gear motors of the same power rating, axial-piston motors also tolerate higher ambient temperatures. Maximum size is about 31/2 hp. Vane motors Rotary vane motors normally are used in applications requiring low- to medium-power outputs. Simple and compact vane motors most often drive portable power tools, but certainly are used in a host of mixing, driving, turning, and pulling applications as well. Vane motors have axial vanes fitted into radial slots running the length of a rotor, which is mounted eccentric with the bore of the motor’s body housing, Figure 2. The vanes are biased to seal against the housing interior wall by springs, cam action, or air pressure, depending on design. The centrifugal force that develops when the rotor turns aids this sealing action. Torque develops from pressure acting on one side of the vanes. Torque at the output shaft is proportional to the exposed vane area, the pressure, and the moment arm (radius from the rotor centerline to the center of the exposed vane) through
which the pressure acts. In a multi-vane motor, torque can be increased at a given speed by increasing the air pressure at the motor inlet to increase the pressure imbalance across the motor vanes. However, there are tradeoffs: increasing inlet air pressure increases air supply costs and generally leads to faster wear and shorter vane life. Output power at a given speed determines air consumption. A small motor producing 1 hp and operating at 2000 rpm using 80-psi air consumes the same volume of compressed air as a larger air motor producing 1 hp at 2000 rpm using air at a lower, more economical pressure. Rotary vane air motors are available with three to ten vanes. Increasing the number of vanes reduces internal leakage or blow-by and makes torque output more uniform and reliable at lower speeds. However, more vanes increase friction, cost of the motor, and decrease efficiency. If, in a 3-vane design, one vane sticks in a retracted position, it can prevent the air motor from starting under load. Spring-biasing the vanes against the housing wall, porting pressure air to the base of the vanes, or camming the base of the vane prevents this problem, as does using a motor with four or more vanes. Vane motors operate at speeds from 100 to 25,000 rpm at the rotor — depending on housing diameter — and deliver more power per pound than piston air motors. Because the vanes slide against the housing wall, many vane motors require lubricated air, particularly if short duty cycles are followed by long inactive periods. However, more and more motors continue to be designed to operate on non-lubricated air to serve critical applications and environmental concerns. Operation of ungoverned vane air motors with no load at high speed should be avoided. When a multi-vane motor operates ungoverned under no load, its high speed can heat and char the vane tips as they rub against the cylinder wall. Abnormal wear and damage to other motor parts should also be expected. Vane-type air motors are available in four basic mounting configurations: base, face, hub, and NEMA-flange. Basemount models simply bolt onto a subbase, and the load is belt-driven or directly coupled. Face and hub mounts are used when the motor must be mounted through a bulkhead or as an integral part
1998/99 Fluid Power Handbook & Directory
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AIR MOTORS
90
12
Maximum horsepower
5.0 4.0 3.0
Starting torque
8 6
60
2.0
4
30
1.0
2
0
50
100
150
200
250
300
350
Total flow Energy-producing flow Leakage flow
10
Flow — cfm
120
Brake horsepower
Torque — ft.-lb
150
14
Torque and power at 90 psi Torque and power at 60 psi
Stall torque
6.0
180
400
450
0 500
1500
Fig. 4. Air motor torque and horsepower characteristics plotted against speed for 60- and 90-psi supply pressures.
of a driven device. NEMA-flange mounts enable air motors to directly replace NEMA-frame electric motors. Gerotor air motors Gerotor air motors deliver high torque at low speed without additional gearing. When coupled with a 2-stage orbital planetary gear train, gerotor power elements provide torque at speeds down to 20 rpm. These motors are well suited to hazardous-environment applications where relatively high torque is needed in limited space. Low-speed/high-torque gerotor air motors can deliver torque exceeding 250 lb-in. within a speed range of 20 to nearly 100 rpm from a 90-psi supply of compressed air. They are rated for continuous operation at supply pressures to 150 psi. Low rotating inertia of the gerotor design produces instant starting, stopping, or change in direction when the valve supplying the motor is shifted. Furthermore, the design prevents the motor from coasting or being backdriven, which can eliminate the need for external brakes. Like vane motors, they are much less sensitive to mounting orientation than piston motors are. Total envelope of the motor, less shaft projection, is 5-25-29-in. Turbine motors Efficiency of an air motor is defined as the ratio of the actual power output to the theoretical power available from the compressed air for the expansion ratio at which the machine is operating. Turbines convert pneumatic power to mechanical power at about 65 to 75% efficiency. Turbine efficiency is higher than other air motors because sliding contact of parts does not occur to cause internal
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2500
3500
4500
Speed — rpm
Speed — rpm
Fig. 5. Leakage through air motor obviously reduces flow available to transmit energy and drive loads.
friction. As a result, there is no need for extensive lubrication. The absence of lubricating oil dramatically improves cold-weather performance. Until recently, turbine air motors typically were used for applications requiring very high speed and very low starting torque — dental drills and jet aircraft engine starters being most typical. Now, however, turbine technology is being applied to starting small, medium, and large reciprocating engines. Turbine technology offers simple, highly efficient pneumatic starters that require no lubrication of their supply air, tolerate contaminants in the supply air, and need little maintenance. Turbine starters include a planetary gear reduction to bring the turbine’s high rotor speed down to normal engine cranking speeds. Turbine motors are relatively compact and light for their power-delivery capability. Higher gear ratios — from 9:1 through 20:1 — provide high stall torque and versatility for a variety of engines. Turbine horsepower is easily changed by limiting air flow through the motor. Operation of a turbine air motor involves a nozzle that directs and meters air to a turbine wheel or rotor. It changes high-pressure, low-velocity air flow to low-pressure, high-velocity air flow. The mass-flow rate of air passing through a turbine determines its horsepower. Changing the number of nozzles or nozzle passages changes power output proportionally. If a 16-nozzle starter is reduced to 8 nozzles, the altered starter will produce half the power of the original. Therefore, within the same basic starter configuration, many models can be designed that have a wide range of inlet pressures, cranking speeds, and cranking or stall
torques. This capability, combined with various gearboxes, allows production of low-cost starters for a wide variety of applications. For example: turbine starters are available to crank engines with displacements from 305 to 23,800 in.3 at pressures from 40 through 435 psig. Performance characteristics Power characteristics of air motors are similar to those of series-wound DC electric motors. With a constant inlet pressure, the brake horsepower of an air motor is zero at zero speed. Power increases with increasing speed until it peaks at around 50% of free speed (maximum speed under no-load conditions). Figure 4. At the peak point, torque decrease balances speed increase. Power decreases to zero when torque is zero, because all the inlet air power is used to force the volume of air required to maintain this speed through the motor. Torque output for an air motor of given displacement theoretically is a function of the differential pressure and a constant that depends on the physical parameters of the motor. Therefore, regardless of speed, torque should be constant for a given operating pressure. Actually, this is not the case, because as air flow increases through the motor, pressure losses in the inlet and outlet lines consume a greater portion of the supply. In practice, torque reaches its greatest value shortly beyond zero speed, Figure 4, and falls off rapidly until it reaches zero at free speed. Starting torque is the maximum torque the motor can produce under load. It is about 75% of stall torque. It takes more torque to start an air motor than to
AIR MOTORS
Direct drive 3:1 Gear ratio 9:1 Gear ratio Torque — ft-lb Power — hp
Torque — ft-lb Power — hp
Motor 1 Motor 2 Motor 3
Power
Power
Torque Torque 2000
4000
6000
8000
10,000
12,000
Speed — rpm
Free speed — rpm
Fig. 6. Three motors producing the same maximum horsepower — but with different torque characteristics — can exhibit substantially different speeds under varying loads.
keep it running. Do not confuse stall and starting torques. If the air motor load exceeds its starting torque, the motor will not start. Stall torque, the maximum torque of an air motor, is about twice the torque at rated horsepower, and can be determined from information on power and speed given in manufacturers’ literature. The relationship between torque and rated power is: T=5250P/n, where T is torque, ft-lb P is power, hp, and n is speed, rpm. Because stall torque is about twice torque at rated power, if n=525 rpm, and P=0.03 hp, then T=3 ft-lb, and starting torque is 2.25 ft-lb. Rated power generally refers to maximum horsepower at 90 psi. Although air motors typically can operate at pressures from 20 to 150 psi at the intake, usual practice limits operating pressure to between 30 and 100 psi. To compare motors rated at different inlet pressures, use this rule of thumb: reduce horsepower 14% for each 10-psi reduction in air pressure. Conversely, a 10psi reduction in air pressure will cut motor efficiency by 14%. Obviously, this relationship directly affects productivity. Again, this is only a rule of thumb and does not apply exactly to any particular motor model. Be sure to measure supply pressure at the motor inlet. It is not enough to determine that there is 90-psi supply pressure at the compressor — line losses usually reduce that pressure before it reaches the
Fig. 7. A motor with a steep torque curve is less sensitive to a drop in speed from a higher load than a motor with a flatter curve. Reduction gearing decreases the influence of the load by increasing the slope of the torque curve.
air motor. There must be 90 psi at the motor inlet for the motor to perform at rated torque and horsepower. Controlling air pressure supplied to the motor is the simplest and most efficient method of changing the motor’s operating characteristics. Conversely, not maintaining the required supply pressure at the motor inlet certainly degrades operating characteristics. There is no direct relationship between power and speed; that is, the lowest horsepower does not indicate the highest speed or vice versa, Figure 4. Free speed is the maximum speed of the motor under no-load conditions. For a governed motor, the term free speed actually means free governed speed, or the maximum speed at which the motor will run while the governor is operating. Design speed is that speed at which rated horsepower is reached. It is about half the free speed of a non-governed motor, and 80% of free governed speed of a governed motor. An air motor operates most efficiently at design speed. Because air motors are constant-displacement devices, their speed, theoretically, is directly proportional to air flow rate. This is true if there is no leakage, but leakage certainly affects motor speed. Leakage increases with pressure, and is nearly constant at any given pressure. Thus, at fixed speed, air consumption increases as supply pressure increases; at low speeds, a much higher proportion of total flow is lost through leakage. A typical air motor performance curve, Figure 5, shows that the additional increment of flow per rpm is nearly con-
stant. Notice, though, that total flow per revolution decreases as speed increases. Leakage also decreases slightly as speed increases, because less time is available for leakage. When the load on an air motor increases, speed decreases until motor torque meets that load requirement. Opening the throttle to the motor to increase inlet air pressure may bring the motor up to rated speed. For applications involving varying loads, the major consideration is whether the motor can provide enough power for all operating conditions. Motors producing the same maximum horsepower but with different torque characteristics can exhibit substantial differences in speed, depending on load, Figure 6. On the other hand, if you wish to reduce change in speed with varying load, select a motor with a steep torque curve, Figure 7. This is because the steeper the torque curve, the less speed changes with load. The influence of the load can be reduced by installing speed-reduction gearing between the motor and the load. This decreases output speed while retaining the same power to increase the slope of the torque curve. Remember, maximum power usually occurs at 50% of free speed, so reducing free speed also reduces design speed, Figure 6. Gearing also reduces efficiency. Another good rule of thumb is to choose an air motor that provides the required horsepower and torque at about 2/3 of available air supply pressure. Full line pressure then can be used for starting and overloads.
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Air regulators P
neumatic pressure controls fall in the category of pressure reducing valves, commonly referred to as air line regulators. It is also essential, once a system pressure has been selected to perform a task, that air be supplied at constant pressure to the actuator, regardless of variations in flow and upstream pressure. Thus, it is important to add to a pneumatic system a pressure regulator that: ● supplies air at constant pressure regardless of flow variation or upstream pressure ● helps operate the system more economically by minimizing the amount of pressurized air that is wasted. (This happens when the system operates at pressures higher than needed for the job) ● helps promote safety by operating the actuator at reduced pressure ● extends component life because operat-
ing at higher-than-recommended pressures increases wear rate and reduces equipment life ● produces readily controlled variable air pressures where needed, and ● increases operating efficiency
Air regulators are special valves that reduce supply pressure to the level required for efficient operation of downstream pneumatic equipment.
Main spring
Aspirator Types of regulators Unbalanced poppet, nonpilot operated — Figure 1 shows the simplest type of unbalanced poppet regulator. Normally, supply pressure enters the regulator and flows Supply Reduced around the poppet, which is pressure pressure seated, blocking flow. Turning the adjustment Poppet screw to compress the adjustStem ment spring forces the diSpring aphragm down. It pushes the stem down and the pop- Fig. 2. Regulator has an adjustment spring and an pet uncovers the ori- unbalanced poppet. It has a separate diaphragm fice. As downstream chamber which contains an aspirator tube connectpressure rises, pressure ing to the reduced-pressure port. It is self-relieving. Ajustment air acts on the underscrew Ajustment side of the diaphragm, balanc- main air flow to help reduce the effects of spring ing the force exerted by the the abrasive air on the diaphragm. An aspirator tube connects the diadjustment spring. The poppet Diaphragm throttles the orifice to restrict aphragm chamber and the outlet chamflow and produce the desired ber. As flow through the regulator inpressure. As downstream flow creases, the tube creates a slightly demand varies, the regulator lower-than-outlet pressure in the diStem automatically repositions the aphragm chamber. The power pressure poppet in relation to the ori- under the diaphragm deflects it downOrifice fice. The spring under the pop- ward, forcing the poppet farther away pet assures that the regulator from the orifice. The adjustment spring Output Supply extends to open the poppet orifice withpressure will close at no-flow. This regpressure out significantly decreasing outlet presulator is non-relieving. Unbalanced poppet, non- sure. The effect is the same as increasing pilot operated with di- the adjustment setting and thus reducing Poppet aphragm chamber — The reg- droop at higher flow rates. Poppet spring This regulator’s much larger diulator in Figure 2 is larger (and Fig. 1. Simplest regulator incorporates an adjustment more expensive) than the model aphragm area produces greater forces and spring (that is, it is not pilot-operated) and an unbal- in Figure 1. It also has a di- thus displaces the poppet more with a anced poppet. It does not have a separate diaphragm aphragm chamber which iso- given change in reduced pressure. Larger lates the diaphragm from the diaphragms increase regulator response chamber, and it is non-relieving..
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A I R R E G U L AT O R S
nisms are then separated. A small air pilot line connects the regulator (in the air line at the point of use) to the remote setting mechanism, which can be mounted at any convenient location. The remote setting mechanism is a small regulator that produces a control air signal. The signal is sent to a pilotoperated, balanced-poppet regulator, similar to the previous regulator except that the top is replaced with a short, pressure-tight bonnet to receive the control signal from the small remote setting regulator. Instead of working against a force created by a compressed spring, the pilot-operated regulator works against a force created by air pressure — that is, an air spring. The air spring maintains a constant force on the upper side of the diaphragm of the pilot-operated regulator because the remote setting regulator holds the control signal at constant pressure. Thus, Fig. 3. Regulator has a balanced poppet and a separate diaphragm chamber with an aspirator. It is non-pilot operated.
and sensitivity. Balanced poppet, non-pilot operated, with diaphragm chamber — This regulator, Figure 3, has the same general internal construction as the previous type. However, it has a considerably larger orifice to allow for greater flow. In addition, to maintain good stability, the poppet is pressure-balanced. That is, the poppet sees the same reduced pressure on both top and bottom surfaces. Thus, the effects produced by reduced pressure fluctuations cancel out, and sensitivity and response are greatly improved. This very-large-capacity regulator has low droop. With an exactly balanced poppet, the system-pressure/reduced-pressure ratio has no effect because the unbalanced resultant forces on the poppet caused by supply pressure are zero. However, these poppets are generally designed with a slight tendency to close. Therefore, a large increase in supply pressure forces the poppet closer to the orifice, throttling flow. This causes the reduced pressure to drop slightly. Remote controlled, pilot operated, balanced poppet — In some applications, the regulator must be installed where it cannot be easily adjusted. The regulation and pressure setting mecha-
droop in this regulator is small. Internal, pilot-operated, balanced poppet — This regulator also uses the pilot-operated principle to produce a precision regulator. Both the pressure setting regulator and the pilot-operated regulator are combined in a single housing. The same force-balance principle applies as in the previous regulator. With this regulator, some supply air bleeds into the cavity over the lower diaphragm and escapes through the nozzle. As increased air pressure on the upper diaphragm opens the flexible seat, the pressure above the lower diaphragm drops and causes the poppet to approach the primary orifice, reducing flow, and thus pressure. A bleeding-type relief seat vents through the center of the diaphragm. This regulator also has a safety-type relief valve above the pilot mechanism, which rapidly exhausts any very high overpressure.
Electronic pressure control As the trend toward automation continues, electronic control of air pressure is a logical progression. Such electronic control of pressure has, in fact, become commonplace — both to allow automatic control of pressure for production machinery and for controlling pressure from a central location, even if the regulator is placed in an inaccessible or hard-to-reach location. Instead of requiring operators to set pressure manually, a regulator may contain an electric motor that turns an adjusting stem until the desired outlet pressure is reached. Another design — sometimes referred to as an I/P or E/P transducer — accepts an electronic input signal (either E for voltage input, I for current) and produces an output pressure that is proportional to that signal. If the command signal (typically 0-5 or 0-10 V DC or 4-20 mA) calls for greater pressure, a valve in the regulator shifts to expose the outlet connection to the higher incoming pressure. If lower pressure is called for, a valve shifts to open and bleed the outlet to atmospheric pressure. With both designs, a pressure transducer senses outlet pressure and produces an electronic signal representative of that pressure. A processor onboard the regulator then com- Line-mounted, electronically pares this value with the command signal and controlled regulator produces acts accordingly to either decrease or increase output pressure proportional to outlet pressure. If needed, the regulator can an analog command signal. also supply an electronic signal representative The regulator shown offers fast of the outlet pressure for process control, data (40 msec or better) response, acquisition, etc. Versions that interface with low power consumption (5 W or less) ±1% repeatability. digital controls are also available.
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Rod boots and bellows
Extend the life of cylinder rods, seals, and other components by protecting them from abrasive, corrosive, and downright dirty environments
W
hen dirt, grit, metal chips, scale, and other substances impinge on hydraulic or pneumatic cylinder rods, these contaminants not only shorten the life of the seals, but often can damage the rods themselves. For almost all applications where contaminant-caused wear is a problem, installing the proper rod boot can make a big difference in rod and seal life. One manufacturer of automated welding equipment noticed a decline in service problems and achieved longer service life on hydraulic cylinders equipped with bellows-type boots fabricated of Hypalon-coated polyester material. Although simple conicalshaped rod boots are sometimes used to protect short-stroke cylinders, a round bellows made by one of several construction methods is generally the most suitable design. Rod boots originally were made by sewing discs of leather or rubberized fabric together to form a bellows. The outer circumference of one disc was sewn to the outer circumference of a second disc. The inner circumference of this second disc was then sewn to the inner circumference of a third disc. The outer circumference of the third disc, in turn, was sewn to the outer circumference of a fourth disc. This alternating pattern continued until the required length was attained. This process is still widely used today, but other methods of attachment in addition to sewing are often used. Today’s suppliers can furnish a wide variety of configurations and materials to protect cylinder rods and seals from most types of contaminants. Selecting the right boot The selection process, of course, requires weighing the interrelated design and operating parameters to pick the best solution for the application at
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Fig. 1. Boots of standard stitched construction meet demands of most applications cost effectively, especially in small to medium quantities.
hand. The following variables should be considered: ● space available for the boot in retracted and extended positions ● cross-sectional area available ● interference points along extended length ● stroke length ● rate of travel and frequency ● ambient and operating temperatures ● type, source, and amount of contaminants, and ● possible need for venting due to rapid travel speeds. Although motion in most applications is linear, some types of mounting or operation with other mechanisms may cause twisting or result in a nonlinear travel path, such as an arc. Some rod boots accommodate such variances better than others. Despite all these variables, custom rod boots are available to handle most common protective requirements, usually with little or no tooling charge. Different boot constructions offer different extended-to-retracted (E/R) length ratios. In general, as boot OD increases while holding boot ID constant, the extended-to-retracted ratio increases. If radial clearance is limited around the cylinder rod, the boot may have too long a retracted length due to OD limitations. To overcome this, the boot can be mounted over the cap end of the cylinder so it encompasses the
entire cylinder. This essentially doubles the retracted length, which effectively cuts the required E/R ratio in half to allow a smaller OD for a given ID. For rodless designs, the entire cylinder usually is enclosed by the boot. Boot construction, design Standard stitched-construction boots, Figure 1, protect internal components from damage by common contaminants and nominally severe environments. Assembled from rings of elastomer-coated fabric sewn along alternating inner and outer circumferences, these boots will exclude most dust, dirt, chips and coolants. For added sealing or abrasion protection, seams can be coated after stitching. This construction offers the widest choice of materials for all types of service requirements and allows the greatest E/R length ratios. End collars or flanges for mounting can be any size or shape, whether larger or smaller than the boot. Stitched bellows cannot be completely sealed, so adherence to air-, light-, or dust-tight requirements cannot be guaranteed. Coating helps, but a sewn boot still tends to “breathe” through its seams. For areas where fine contaminants require more complete exclusion, it may be necessary to seal the boot so tightly that it may collapse if the travel speed is too
BOOTS AND BELLOWS
Ranges of E/R ratios and stroke-toretracted-length (S/R) ratios for the various designs and material thicknesses are available from manufacturers.
Fig. 2. Common method of mounting boot uses hose clamp to secure boot collar to rod and backing plate to secure boot flange to cylinder.
rapid. Realize that the boot is actually a bellows, so a vacuum is generated during rod extension, and internal pressure develops during retraction if adequate venting is not provided. Methods will be described for providing the necessary makeup air while still excluding contaminants. For more complete sealing and improved appearance, a vulcanized bellows of Hypalon-coated polyester material can be used. These are similar to the stitched bellows, except that seams are vulcanized instead of stitched, making them air-, light- and dust-tight. Because these seams are wider, the E/R ratio is slightly lower than with standard stitched construction. This factor becomes critical primarily if space for the boot is limited when the cylinder rod is retracted. Injection-molded bellows configurations offer complete sealing but present some disadvantages. Molds are expensive, so volume requirements for costeffectiveness may be too high for most applications. In addition, the unreinforced material lacks strength and can tear, and the rounded convolution shapes usually exhibit a poor E/R ratio. In most cases, a suitable custom boot can be produced by other processes using standard tooling and components. An excellent alternative is a bellows-type cover expanded from a tube of fabric-reinforced elastomer. It provides the same tight-sealing advantages of a molded boot, but with additional strength and low tooling charges, if any. These boots are cured or vulcanized after forming and can be made with a variety of end mounting configurations.
Materials, mounting Many different materials can be used to produce rod boots, depending on the operating conditions and the manufacturing process selected to meet the needs of the application. Base fabrics include nylon, fiberglass, and Kevlar. Elastomers commonly used for boots include butyl, nitrile, neoprene, silicone, synthetic rubber, fluoroelastomer, PTFE, and polyurethane. Other specialized materials also are available, such as aluminized fiberglass for welding applications. Even though the potential damage caused by abrasive debris is a legitimate design concern, fabric-reinforced elastomer materials used in these covers will withstand a fair amount of abuse. When conditions are extreme, such as very sharp or hot particles, even more durable materials can be specified, such as Hypalon for greater abrasion resistance than neoprene. Other types of contamination could require different engineering approaches. Covers for food processing machinery, as a case in point, generally use special material that will not contaminate and can withstand frequent washdowns. A formed or molded construction usually is used to provide smoother contours than a stitched bellows. Two basic methods are used for mounting most rod boots; ends are held in place either by a collar and hose clamp or by a flange and backing plate. Most often, a hose clamp fastens a collar on one end of the boot to the rod end or clevis. At the other end, a flange at the end of the boot may be bolted to the face of the cylinder with a large washer-type backup plate that has appropriate mounting holes, Figure 2. The boot also can be furnished with a collar, sized to fit the cylinder body or a boss, if it is long enough, and fastened in place with a hose clamp. Certain characteristics can be added to most boot designs. In addition to flanged or collared ends for mounting, typical accessories include screen vents or breathers and tie strips or tapes to ensure uniform extension. Longitudinal zippers, Figure 3, accommodate retrofit
Fig. 3. A longitudinal seam allows rod boots to be installed on existing equipment without having to disassemble components. Zippers, snaps, Velcro, or other fasteners can be sewn into boot to aid assembly and provide access for inspection, adjustment, relubrication, etc.
installations where disassembly would be difficult or impractical. This longitudinal slit also allows access to internal components for inspection, relubrication, adjustment, etc. Application considerations When considering the various rod boot choices and the specifics of your application, don’t overlook the little details that can make the difference between effective protection and consternation. Many times, what would be a troublesome condition can be accommodated with one of the optional designs available from the manufacturer. For example, installing several screen breathers may allow sufficient venting for applications if contaminants are course enough. Finer contamination may require a breather hose routed to a filter or a cleaner intake location. Potential interference throughout the length of the proposed boot’s travel is often overlooked. Methods of accommodating some obstructions include extra-long mounting collars or special flanges. Thoroughly checking and documenting any obstructions will enable the manufacturer to suggest the most effective way of handling them. A concern that usually can be neglected is the load imposed on a system by the boot. Negligible force is needed to operate most rod boots through the cylinder stroke when compared to the high cylinder forces generated in most fluid power applications. In other than the most sensitive laboratory equipment, load from the boot can be neglected.
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Carriers for hose and cable
When long sections of hose or cable must follow motion of a machine element, carriers can keep individual members from becoming kinked, tangled, or damaged.
W
hen hydraulic or pneumatic lines connect to a moving machine component, there are several ways to accommodate the motion. While the easiest is to simply allow slack hose, this method can lead to excess flexing, premature hose wear, and damage from pinching or exposure to other hazards. Hose reels are another possibility, but generally they are useful only for single or dual lines and may not accommodate longer travel distances or larger hose sizes easily. Another approach that might prove to be better is to use a cable and hose carrier that forms a flexible shield for multiple lines, preventing them from tangling, and paying out the hose with a smooth, rolling action. Hose carriers can be divided into four basic types: metal link, metal tube, non-metallic link, and non-metallic tube. Each type is best suited to a different combination of application parameters, although the choice is not always clear cut. Metal-link carriers are simply parallel sidelinks joined by crossbars that accommodate the hoses. Pivot pins and stops allow the links to travel through a predetermined arc. These designs are best for low-speed, heavy-load applications, such as machine tools. Enclosed metal carriers are formed with small convolutions to provide better protection against chips, weld spatter, and other particles. Because carriers of this design are not limited by individual links and with their distinct travel arcs and stops, they also can operate at higher speeds, an advantage in applications such as robotics. Where greater weight- and hosecarrying capacity for longer travels are needed, non-metallic link carriers can handle higher speeds and provide greater width and carrying capacity. This type leaves hoses open for easy inspection, but does not protect them from abrasion.
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Four basic types of hose carriers include: metal link (upper left), metal tube (upper right), non-metallic link (lower right), and non-metallic tube (lower left).
Non-metallic tube carriers are the best choice for applications that require fast travel speeds and also need protection from outside elements. Few of the lines between parameters are clearly drawn, so it is best to begin working with a hose carrier manufacturer early in the design of any equipment with hydraulic or pneumatic lines that connect moving parts. This allows greater flexibility in choosing the hose carrier design that provides the best combination of features for the application. Here are a few general guidelines for making a preliminary selection. For size decisions, the practical maximum width for an enclosed metal tube carrier is about 6 in. Very small sizes, conversely, are easily handled in non-metallic designs. Small sizes, however, are not practical in the metal-link types.
Metal link carriers are molded from fiberglass-filled nylon, which provides high strength per unit weight. However, at about one-fourth the strength of steel, such materials are not practical for carriers combining large widths, the weight of multiple lines, and long lengths. These factors all interact to determine the final choice. For example, a lowspeed application with moderate particle exposure probably calls for a metal link design. A combination of higher speed and hot welding or grinding spatter points toward a metal tube. A highspeed application with the need to monitor hoses for leaks calls for an open non-metallic design, particularly with a large number of hoses. A high-speed application that also needs corrosion resistance, such as a car wash or machine with heavy coolant spray, would be a logical place for non-metallic tube.
Valve cartridges can be thought of as “bodyless” valves — valves without an integral housing — because cartridges are only the guts or internal moving elements of the valves.
Cartridge valves A
fter a cartridge is inserted into a cavity, such as a manifold with appropriate flow passageways, the resulting valve performs like any conventional valve. Slip-in cartridges are held in the cavity by a cover plate, Figure 1; screw-in-type cartridges mate with threads in the cavity, Figure 2. Another type of insertable cartridge has circumferential grooves. After it is inserted into the cavity, it is held in place by swaging internally with a tapered pin that expands the cartridge diameter into interference contact with the bore. Versatility A wide variety of cartridges allows engineers to find almost any hydraulic control function in cartridge form — with very few limitations. Here are the control functions readily available in cartridge configuration today: ● relief valves ● sequence valves ● pressure-reducing valves ● check valves
Fig. 1. Slip-in cartridge for pressure reducing consists of biasing spring, poppet, sleeve, seals, and unshown cover plate.
● pilot-operated check valves ● load-control and counterbalance valves ● flow-control valves (pressure- and non-pressure-compensated, fixed, priority, proportional divider) ● solenoid valves in 2- or 3-way poppet or spool type, and 4-way, 2- or 3-position versions, Figure 3 ● electro-proportional directional, flow, and pressure controls, and ● specialty valves, such as shuttle valves and velocity fuses. One innovation in cartridge valve technology today is the incorporation of two or more functions into a single cartridge housing, e.g. check and flow valves, dual crossover reliefs, solenoidoperated relief, etc. Custom valve packages A manifold with at least one cartridge valve — but more commonly with two or more — may be thought of as a valve package. Often, the manifold is custom designed for a specific application or function. The valve package may contain some or all of the control valving for a hydraulic system. Custom-designed valve packages offer many advantages in mobile and industrial equipment. A custom valve package can be: ● a single cartridge installed in a manifold, designed to fit a specific mounting configuration, or ● a manifold containing multiple cartridge valves performing some or all functions in a complete hydraulic control system, Figure 4. System advantages The cartridge-design approach can offer important advantages over systems plumbed with individual components: 1. Light weight and compact size. Combining several functions into a single manifold may save valuable space and weight, often occupying only 10 to
Fig. 2. Pilot-operated relief cartridge screws into manifold cavity, seals with O-ring.
20% as much volume as conventional line-mounted systems. This space and weight reduction may provide substantial energy savings, particularly in mobile applications. 2. Economy. On many systems, the cost of a custom valve package — if built in sufficient quantities — can be equal to or less than that of a system with individually plumbed components. This premise is especially valid if one compares the total installed cost, considering all components and labor. 3. Leakage prevention. Because a valve package eliminates so many connections, system leakage potential is reduced drastically. In addition to fewer connections and potential leakage sources, custom valve packages re-
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C A R T R I D G E VA LV E S
duce the amount of attached conventional plumbing that may work loose from machine vibration. 4. Efficiency. Valve packages which eliminate interconnecting lines and fittings lead to higher overall system effi-
Fig. 3. Several types of single- and doublecoil solenoid cartridge directional control valves with integral manual overrides.
Fig. 4. Multi-function manifold block incorporates ten 3-position, 4-way and two 2way cartridges, plus pump-pressure control and unloading function. This manifold interfaces with a computer to operate an industrial wood-sawing system.
Fig. 5. Manifolded control system for asphalt-paving machine contains 27 cartridge valves.
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ciencies. Not only is pressure drop through lines and piping reduced, but overall machine weight and space requirements also are reduced. Higher energy efficiency can be achieved for a given machine, particularly saving fuel in mobile applications. Many times, a valve package can be mounted directly on an actuator or pump where it will be most effective in performing its functions. The close proximity of valve and actuator also will minimize system compliance for those systems — such as aircraft flight controls — where the position of a controlled surface is critical. 5. Convenience. A factor important to OEMs: the valve package that they receive from their supplier is normally already fully assembled and pretested. They simply mount the package on their machine, connect it, and the system is ready to run. Being more compact, valve packages sometimes can be mounted where they are less susceptible to external damage. 6. Serviceability. Machines not only must be reliable, but also should be easy to service and maintain. This is an important characteristic of valve packages because troubleshooting and servicing may be a simple procedure. Cartridges usually can be removed and replaced quickly without disturbing any external plumbing. 7. Aesthetics. Finally, eliminating piping, fittings, etc. normally leads to a better looking machine. A clean design will not affect overall performance of the machine, but a good-looking machine on the dealer’s sales floor in a competitive market can be very valuable to the OEM. Logic circuits The advent of hydraulic logic elements (basically 2-way , slip-in poppet cartridges), that can be controlled with low-flow pilot signals, has expanded the horizons for valve packages. These compact and cost-effective elements can handle flows from 5 gpm to more than 5000 gpm, at pressures to 6000 psi or higher. Individual logic elements can be arranged so that trapped fluid pressure is relieved before full connection to a tank line. Also, logic elements can be actuated at optimum points in the system’s
Application precaution When considering the use of poppet logic in a system, circuit designers must realize that they also are assuming the role of valve designer. In this capacity, they must address potential problems such as smooth shifting of elements and efficient manifold design to minimize system pressure losses. These are problems they may not have faced with conventional valving. cycle — much as a spark plug must fire just before top dead center in a gasoline engine for optimum performance. Disappearing disadvantages In the past, producers of cartridge components expanded their product lines by adding proprietary items. Cavities were randomly developed and manufacturer-to-manufacturer interchange was rare. Today, however, a series of industry-common cavities have evolved through competitive pressure — to the extent that now several sources typically can compete on a fit/installation basis. In addition, an NFPA/ISO effort is in process to develop a world standard for cartridge cavities. Dimensioned metrically, these proposed standards are already being implemented by some cartridge makers. Also, cartridge product lines continue to proliferate so that choices of hardware are available for most function requirements. In addition, the viability of complex manifolds has grown. Building manifolds for more than six or eight cartridges once was considered risky because of potentially high scrap costs if mistakes were found near completion of machining; and cleaning and deburring often were difficult. Now, CAD capability and CNC machines have greatly reduced the concerns associated with fabrication by minimizing the possibility of human mistakes. These techniques also shorten the product-development cycle. New thermal, ultrasonic, and chemical deburring techniques simplify manifold clean-up. As a result of these factors, manifolds with 20 or more cartridge cavities, Figure 5, are no longer unusual.
Clampsfor hose, pipe and tubing
Properly fastening hydraulic pipe, tubing, and hose in place is critical to the long-term success of fluid power systems.
U
nrestrained hymaterials — including draulic tubing wood, rubber, metal can emit sound and thermoplastics — as though it were a tunand designs to accoming fork. Fluid velocity, plish this task. The suppressure, and line size port systems generally all contribute to the high fall into two categories: frequency vibration, individual supports that line shock, pressure hold single tubes, and surge, and sound, which multi clamps to secure are prevalent in many two or more lines. Most hydraulic machine opersystems can accommoations. Shock, pressure date pipe and tube sizes surge, and vibration will from 3⁄16 to 24 in., with flex metal tubing. This others handle sizes up flexing can cold-work to 42 in. Most will stack the tubing, particularly for several tiers, and can around connectors. The be fastened down with a consequences are split Clamps on vertical structural members hold horizontal piping runs in place, limit weld base or to a vibration and shock, help control noise. fluid lines. mounting rail. Accordingly, the fastening components used have proExcessive noise for plant personnel Support material gressed from simple clamps to spe- is only one negative effect. More imWhile wooden blocks and rubber cially engineered pipe and tubing portantly, if pipes and hoses are not grommets are quite adequate for supsupports. A properly designed and in- properly supported, vibration may porting pipes and tubing, they do have stalled piping-support system can re- cause fittings and flanges to loosen. several disadvantages. Wood tends to duce noise pollution, risk to property This can lead to increased safety risks dry out, crack, and lose its absorption and personnel, and environmental for plant personnel, downtime, higher properties. Wood is also a relatively exdamage caused by the discharge of hy- maintenance costs, inefficiency of the pensive support material. Aside from its system due to air causing cavitation in lack of chemical and abrasive resistance draulic fluid from a damaged line. pumps and valves, and loss of hy- properties, rubber is generally only used draulic fluid. The economics of hy- in conjunction with some kind of steel Effects of vibration Pipe supports should control vibra- draulic leaks as a result of loose fit- strap or surround, adding to the cost and tion and absorb shock. Pumps and tings and flanges can be quite complexity of these supports. Supports made from such thermovalves are the greatest cause of vibra- revealing. While a dripping flange tion and shock because of the drastic might only be a messy nuisance, the plastics as polypropylene (PP), polypressure and fluid velocity changes that case of a northeastern steel mill which amid (PA), and Santoprene (SA) overoccur in these components during oper- lost approximately 20,000 gallons of come the drawbacks associated with ation. While intensive research has re- hydraulic fluid a month is alarming. other materials. Due to the high molecduced pump- and valve-induced vibraular structure of these thermoplastics, tion, it is still considered a problem for Resilient supports they readily absorb shock and vibration. Resilient supports have been devel- In addition, they are resistant to chemithe entire piping and tubing system. Through piping and tubing, vibration, oped to dampen the vibration in piping cals and abrasives, expand and contract shock, and related noise are transmitted systems. There are several types of sup- with changing temperature, and, at the and amplified. Note that pumps are the ports available on the market that ab- same time, exhibit high strength characgreatest source of noise, and reservoirs sorb shock and vibration to different teristics. Finally, supports made from degrees. These products utilize various thermoplastics are inexpensive. provide the strongest amplification.
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CLAMPS
Multi-tube support systems Generally, multi-tube clamping systems are manufactured from pressed steel. The tube or pipe is held by a split rubber or plastic grommet located in the steel clamp body; this grommet is held in a steel body by a retaining ring molded into it. This type of clamp is ideal for general hydraulic and pneumatic systems that do not generate excessive high-frequency vibration and is satisfactory for line diameters up to 2 in. and temperatures to 250°F. Several models offer the added versatility of interchangeable grommets to hold different size tubing while using the same clamp assembly. Some multi-tube systems allow fitting inserts or junction adapters to be placed between the clamping halves. These inserts permit the clamp to serve as a bulkhead connector plate or a junction block, tube-to-tube connection, tube-to-hose connection, or terminal. Another feature of some multi-tube systems is fixed centerline distances, both horizontally and vertically. This can be an advantage or disadvantage, depending on the installation. Individual tube clamps Some systems clamp individual tubes, which allows the design of a modular clamping system to accommodate as many tubes as necessary. Most clamping systems are for clamping individual lines. Single clamps are either fastened directly to a mounting surface at predetermined positions or located in a mounting rail, allowing for adjustment after mounting. Individual supports can be either molded plastic with an identical upper and lower half, or of the rubber grommet design with a molded grommet surrounding the pipe or tube. Steel retaining pieces hold the grommet tightly around the pipe and fasten to the mounting plate. The grommet design allows for quick and easy installation and adjustment, but is not suitable for large diameters or where heavy shock and vibration are present. This style of support cannot be stacked easily. Thermoplastic supports Manufactured by injection-molding processes, thermoplastic supports con-
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Table 1 Manufacturers’ recommendations Tubing diameter OD — in. 3 ⁄16 – 3⁄8 1 ⁄2 – 7⁄8 1 – 11⁄4
Distance between clamps — feet 3 5 7
sist of identical upper and lower halves that are held together by one or more steel bolts, often in conjunction with steel bases and cover plates. Using identical upper and lower support halves eliminates potential confusion during installation in the field. To attach the support, base plates are either tack-welded or, in the case of high-vibration applications, continuously welded to any available structure. To further simplify installation, multiposition rails have been developed to accommodate several individual supports side by side without wasting space. Additional advantages result from breaking down traditional multi-tube clamps into individual supports. Not only are they less expensive and easier to install, but maintenance or moving of pipes and fittings is greatly facilitated by providing access to one particular line without disturbing adjacent lines. Some multi-tube clamps solve this problem by cutting the upper half of the unit for convenient removal of several tubes for service. Individual supports are also ideally suited to compact and secure pipes and tubing in cramped areas. Several of the supports can be stacked on top of each other resulting in an uncluttered yet efficient use of available space. Routing After analyzing your hydraulic system, route your tubing by the most direct route. One manufacturer suggests routing lines along the bases of the machines. Beware of obstructions, moving parts, contours, and service openings. Plan around them. If this base routing is not economical, consider upright carriers for partial line suspension. If this is not practical, full overhead suspension is the last resort. One particular feature that saves trying to “measure in the air” is a clamp with standard centerline dimensions regardless of tubing size. This is helpful when laying out original tube runs.
How far apart? Manufacturers recommend spacing between clamps as shown in Table 1. All tubes should be clamped on both sides of a bend, as close to the bend radius as possible. This prevents tube movement if internal hydraulic shock occurs. This spacing permits a tubing system wall-thickness reduction from 0.095 in. to 0.065 in. with no splitting in the reduced thickness tubing walls. A well-designed fluid power system will include a tube and pipe support system. The fluid system may be for a single machine, a central system that serves an entire plant, or any size in between. Proper tube support pays dividends — for the designer, the maintenance crew, and plant personnel. How to mount clamps Methods of fastening clamps to supporting members vary. Some multi-clamp assemblies can be bolted on, can use a tapped hole, or even can be secured to a suspended column by bolting through special mounting adapters that use tube clamping holes in the clamp assembly. Power-driven studs are available for wood and masonry, too. Other multi-clamp sections, and some individual clamping models, have feet or brackets, which are welded to the supporting surface. The lower half of some are simply welded in place. Some individual clamp types can be anchored in channeltype metal pieces, which are welded or bolted to the support surface, and again allow designing the overall clamp assembly as wide as desired. Most support systems can be mounted on strut-type rails that already may be present in some installations, allowing light-duty hydraulic and pneumatic lines to run alongside water and gas lines. When stacking clamp assemblies, the usual holding method is with studs that thread into the bolts or studs of the lower clamp assembly. One variation provides a wider bottom clamp assembly; the bolts of the upper assembly thread directly into the top of the lower clamps. When stacking supports horizontally, take care that the weight of the assembly does not exceed the manufacturer’s recommendations.
Control networks for fluid power
I
n an industrial-control setting, a transducer measures some system output variable, say shaft speed. Then, based upon the desired or command(ed) speed, an error is calculated and used to bring actual speed into closer agreement with command(ed) speed. Such applications are called feedback control systems and their general form, Figure 1, indicates how the links of the measurement and control network is expanded into several subparts. The expansion is presented because it is necessary to investigate and understand the measurement problem in some detail. Measurement link 1: Measurand — Certainly, the primary consideration is the variable that must be measured and controlled. Selection of this variable dictates the nature of the output which, in turn, dictates system interfacing and signal-conditioning needs. If the variable to be measured is fluid temperature, for example, the nature of the signal from the temperature transducer most likely is of a different form and strength than if the measured variable is cylinder position. Measurement link 2: The physical
sensor — The heart of a transducer is its physical sensor. It might be a conductor that becomes deformed, as is the case of a strain-gage transducer, or the sensor might make use of the Hall effect as do some encoders. Whatever the selection, it affects the other links in the measurement and control chain. Measurement link 3: Sender signal conditioning — Here, a transducer sometimes will be called a sending device and it will be helpful to distinguish things on the sending end from things on the receiving end. Many transducers are equipped with integral signal conditioners or interfaced with nearby signal conditioning equipment. Such signal conditioners may be simple amplifiers, phase-sensitive demodulators, or special band-pass filters. There are hundreds of possibilities; only a limited number are included. Handshaking, indicated between boxes 3, 5, and 6, and between 6, 5, and 3 in the reverse direction, Figure 1, is a term the digital industry uses to explain that two digital devices that want to share data must coordinate the availability of the data at the digitization end with the interrogation process on the
Transducers are used in the laboratory for collecting data regarding the status or performance of a component or system. Transducers also are used in a control environment wherein the data collected from the transducer is used to control a machine or process that the transducer is monitoring. Although many problems and challenges are similar in both categories, some issues are unique to the industrial control environment.
receiving end. Measurement link 4: Data transmission medium — The data transmission medium is the method used to transmit information from the transducer to the receiving device. Again, there are myriad possibilities, but in an industrial-control application, wires are usually the medium. On occasion, fiber-optic cable could be used to advantage if some special noise-control methods were necessary. In the case of an aerial lift, the medium could be an FM radio signal. If the sending-end signal conditioner generates a 24-bit parallel digital signal, the data transmission medium must have 24 signal wires plus a ground wire. On the other hand, if the sending signal conditioner is an analog DC output device, the transmission medium can usually be a simple twoconductor shielded cable. Only a few of the many possibilities are considered because apparatus such as FM radio transmitters and FM receivers/demodulators and their designs are left to qualified experts in those fields. Practitioners of the electrohydraulic art are advised to purchase such equipment as
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CONTROL NETWORKS
turnkey operational packages should they be needed. Measurement link 5: Receiver signal processing — Signal processing on the receiving end may or may not be needed. The most common possibility is that the sending device transmits a 4to 20- mA current-loop signal but the receiver requires a voltage — that is the receiver has a high input impedance. The typical industrial servo amplifier is an example of a voltage-processing receiving device. If a servo amplifier is the receiving device, a resistor, usually in the range of 100 to 500 Ω, will convert the current signal into a proportional voltage that the receiver then can process. Should the transmission medium be a radio signal, the receivingend signal processor must be a radio receiver and a suitable radio-frequency (RF) demodulator. Measurement link 6: The receiving device — The form this device can take has many possibilities but two common ones are a proportional valve amplifier or a servovalve amplifier, most often called simply a servo amplifier. Given the popularity of programmable logic controllers in industry and the growing use of motion controllers, the receiving device of the future increasingly will be a specialpurpose digital computer. If this is the case, interfacing decisions take on
new dimensions. If the transducer generates an analog signal, and analog-to-digital converter must be somewhere in the measurement chain or in the receiving computer. If the physical sensing device is, say, an absolute-position encoder with parallel digital output, the encoder can be coupled directly to the parallel input port of the receiving computer. Control link 7: Controller — A controller is a component that receives the conditioned data that represents the measured variable, compares it to the command, and issues any necessary corrective signal to the power amplifier. The controller could be an analog proportional or servo amplifier or it could be a special- purpose digital computer. The controller could even be a general- purpose desktop or laptop computer if it is configured with the proper hardware and software. Control link 8: Power amplifier — For electrohydraulic control systems, the power amplifier is the driver stage of the servo/proportional amplifier. Motion controllers may have the power amplifier as an integral part of the controller. Amplifier output drives the valve coil. Control link 9: Power output — The power output component is the hydraulic actuator: a cylinder or rotary actuator.
Measurement links Handshaking
1 Measurand (variable to be measured)
2 Physical sensing device
3 Signal conditioning, sending end
4 Data transmission medium
5 Signal processing, receiving end
6 Receiving device
Handshaking 9 Power output
8 Power amplification
To load
Machine or process to be controlled
7 Analog or digital controller
Commands
Fig. 1. A transducer and its support equipment for several links in the chain of the measurement-control, closed-loop network. All measurement links may not be present or necessary in all systems.
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The loop formed in Figure 1 is the measurement-control closed-loop chain of which the measurement links form a vital part. It can be shown that the measurement device and its supporting equipment alone dictate the performance of a high-gain closed-loop system. That is, deficiencies in the measurement network chain propagate one-for-one into deficiencies in the control of final system output. Transducer performance is crucial and its selection is critical. Three elements Usually, three elements of Figure 1 are chosen early in the design process or are mandated by application requirements: ● the control problem at hand usually dictates the measurand. If one wants to control the speed of a motor, for example, it is unlikely that a temperaturemeasuring transducer would be called upon to do this task; a speed transducer is the more likely choice. Once the measurand is selected, the choice of transducer most often narrows considerably: there are no commercially available speed transducers in which strain gages perform the basic act of transduction; frequency-generation methods are more popular and can be cost effective. Suppose that a frequency-generation transducer has been selected; the signal conditioning equipment must then be able to process a frequency input as opposed to an amplifier for a strain gage bridge or thermocouple analog voltage ● application circumstances many times control the data transmission method. The transmission medium in most industrial hydraulic systems will be a multiplicity of wires, but if necessary to remotely control a farm tractor, for example, radio waves may be a real possibility. If the distance between sending and receiving devices is great, a 4- to 20- mA current loop may be the choice. If there is a distance problem as well as a ground-loop noise problem, then opto-isolation and fiber-optic cabling may be called for. Whatever constraints and requirements apply, the data- transmission medium will significantly affect the nature and extent of the signal processing/conditioning equipment on the sending and receiv-
CONTROL NETWORKS
ing ends of the measurement network chain, and ● the receiving device will affect interfacing decisions. If the receiving device can only receive digital signals, the receiving signal processor must produce digital outputs. If the receiver is an analog device such as a servo amplifier on the other hand, the receiving signal processor must output analog signals. Carrying this point further: if the sending-end signal conditioner generates a high-level DC output (more than ±1 V) and the physical distance between the transducer and servo amplifier is just a few feet, there is no need for receiving-end signal processing; the servo amplifier can do it all. Make note of this — the servo amplifier is at once the receiving device and the controller. What we find then is an extensive jargon and terminology of transduction and control along with a block diagram that shows function more than it does physical hardware. This is not unusual in the technology of control systems. That is, the block labeled Receiving device, Figure 1, may in one instance be a piece of real and separate hardware such as an analog-to-digital converter. But is also may be, for example, nothing more than the 500 Ω resistor in a servo amplifier that converts a 4- to 20mA current signal into the proportional voltage that the servo amplifier needs. All of this recognizes that a great deal of the problem for the fluid power practitioner is one of terminology, definition, and common usage. Be aware that the diagrams used by control-systems artisans probably will contain blocks that define function as opposed to depicting separate, identifiable boxes that one would buy in a store. The hardware needed to implement a function may come all in one box or in several boxes. It is impossible to predict without dealing with specific applications and choice of specific hardware. Selection questions When selecting actual hardware, always ask this question: of all the functions needed to perform the desired control action, how many are contained in this specific piece of hardware and how many must be obtained with other pieces of hardware?
Analog data format methods DC
±1v,±5v,±10v,0 to 5v, 0 to 10v, others. No standards for industrial controls
AC
4—20mA current loop
Pulse width Pulse position Frequency Phase Amplitude modulation modulation modulation modulation modulation Carrier
Carrier
No Carrier
Supressed carrier
+AM
The answer to this question will be enhanced if the fluid power practitioner can remember, appreciate, and understand the: ● overall picture suggested in Figure 1, and distinguish between function and hardware boxes. Eventually, the hardware boxes must be identified, but at planning and design time, it helps to consider only function ● basic physics of the transducer in question because its operation will help select the transducer ● variety of formats in which data can exist. For example, when dealing with analog or digital signals, will the information be the frequency or the amplitude? and etc. ● various data transmission media and what factors assist in the decision to select one over the other. If application circumstances dictate a particular transmission medium, understand the signal conditioning/processing implications attendant with that decision, and ● nature of the controller and whether the other elements in the measurement network chain require a separate receiving device or can the receiving-device function be adequately performed by the controller. Adding to this multitude of possibilities is the fact that there are so many different transducers that perform the same function. For example, there are at least 20 different ways to measure liquid flow. Each method has its own particular technology, generates its own output data form, and requires its own special signal- processing equip-
Fig. 2. The analog signal family tree of data indicates the different forms that data may take in an electrohydraulic system. Several forms may exist in different parts of a single system.
ment. Add to that the possibility that there are several different controllers and data transmission media, and the number of combinations becomes truly astronomical. What follows reduces that astronomy into manageable proportions by looking at things the practitioner of the electrohydraulic art may encounter in the work-a-day world. Forms and methods for data transmission In the electrohydraulic control situation, components are connected in a chain-like fashion to form the feedback-control loop. The purpose of the several connections is to transmit information or data from one point to another in the loop to achieve the desired degree of control. Obviously, the information exists in a variety of forms at various points within the loop; for example, the transducer output may be a voltage; at the valve, input usually is a current; within the actuator, information is in the form of a pressure and a flow, and so on. When expanding the purely electronic elements within the loop, there are a variety of forms within that medium alone. The transducer probably outputs an analog voltage but within the controller (depending upon the type), the data may exist as a digit in computer memory. In between, there may be pulse-position modulation, a frequency-modulation element, or even a double- sideband, suppressedcarrier, amplitude-modulated signal. The possibilities may be endless.
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CONTROL NETWORKS
The general subject of the forms and methods of data transmission is covered in detail in an electrical engineering course called Information Theory or Information Transmission. That course deals with the many possible forms in which information can be processed and transmitted. Fortunately, for the electrohydraulic engineer, there are only a few forms used for data transmission. There are two broad data-form categories that must be understood to cover most electrohydraulic applications: digital and analog. In the following, electronic counters will be thought of as components or black boxes in the control/measurement network chain mostly to show methods of converting data from analog to digital form when the data is not a simple analog DC voltage. Because all these data forms are common in electrohydraulic control systems, they should be understood at least in a conceptual sense.
speaking, the DC branch is really a subset of the AC branch, but the fluid power technology does not make this distinction. The fact of the matter is that DC, by definition, is absolutely constant and therefore cannot transmit information because it never changes. But information is transmitted in the fact that the DC level does change. For example, when the output voltage of a pressure transducer changes, one knows that the pressure must have changed commensurately. That change is the data; it is the information. But, the argument goes, if the voltage changes, then by definition it must be an AC system. The concept seems almost philosophical and it is treated that way here. But it is more than that for the electronic circuit designer because indeed, AC circuit theories must be used for designing this thing that is categorized as a DC data transmission system. In spite of these arguments, the tree of Figure 2 has a DC and an AC branch.
Analog data forms A family tree of the analog signal data forms likely to be encountered in electrohydraulic systems, Figure 2, shows the two major branches of the tree, namely DC and AC. Strictly
Sending end 1 Source of analog DC voltage signal
DC data transmission forms While no standards govern the voltage level in the generation or transmission of DC analog signals, actual practice is slightly different. There Receiving device options
Transmission medium V10V
R1
Digital multimeter
R2
A/D converter Optional attenuator
Source of analog 2 - 20 mA data signal
4
Current-tovoltage resistor
A
Computer
Zero offset
Fig. 3. Either a voltage or a current source, located at the sending end (left) of shielded and twisted-pair analog DC data- transmission lines 1, 2, 3, and 4, sends low-amplitude voltage (±100 mV) in line 1; mid-amplitude voltage (100 mV to 10 V) through line 2, and high-level amplitude (±10 V) through line 3 to the receiving device. The fluid power industry uses 250- or 500- Ω input resistors to convert the current signal to a proportional voltage. Any analog receiving device (right) can be used depending on needs of the application.
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are a large number of systems that use a 0- to 10-V signal or a ±10-V signal. Most commercial transducers come with options that often include those values, or the transducer can be equipped with signal conditioning modules that provide that output. Furthermore, receiving devices that accept and process 10-V signals are easy to design, economical, and commonly available. This also is true of 5V signals; 5- and 10-V signals are consistent with the normal 5- to 24-V range of DC power-supply voltages and each gives good noise immunity because the noise levels usually are well below peak signal levels. Some of the options that should be considered in a DC analog data transmission system are indicated in Figure 3. When the system uses voltage as the data-carrying variable with the socalled high-level signals between 100 mV and 10 V, the interconnection is very straightforward. A signal wire and ground connection are all that are needed when the interconnection distances are, say, below 20 ft. In any event, it is always good practice to use a shielded, twisted pair as the interconnecting method, understanding that shielding becomes more important as the separation distance increases. This shielded, twisted pair refers to the type of cabling that has two signal wires and a common conducting sheath usually made of braided wire called the shield. When transmission distances exceed 20 ft, line 2, Figure 3, it may be desirable to include a buffer amplifier at the sending end of the data link. This amplifier provides a high input impedance for the transducer to look into, and at the same time provides a low output impedance to drive the transmission line. Both impedances will work to enhance the noise immunity of the data link. When the voltage level at the sending end is less than 100 mV, amplification at the sending end is almost mandatory. This statement is made assuming that the receiving-end device, such as a proportional servo amplifier, will best be served by a 5- or 10-V maximum signal level. That may not always be the situation at hand, so the suggestion must be interpreted as a rule-of-thumb; each application has its own specific considerations.
CONTROL NETWORKS
Note that the correct position of the amplifier is at the sending end, especially if the distances are more than two or three ft. This will raise the voltage, probably to the 5- or 10-V range, with the amplification taking place in the noise-protected environment of the sending device. With this increased voltage level, noise induced in the more-vulnerable interconnecting cable will have less effect that if the voltage were transmitted at the millivolt level. That is, the signal-to-noise ratio of the system will be higher when using a sending-end amplifier. For signals that exceed 10 V (the exception rather than the rule), it probably will be necessary to attenuate the signal so that it is in the 5- or 10-V range. If the attenuator is placed at the sending end of the transmission line, it can pose a problem. Voltage divider R1 and R2, line 3, Figure 3, will undoubtedly increase the effective output impedance of the sending device and will reduce the noise immunity of the line. Another reason it may be necessary to include a buffer amplifier at the sending end is to reduce the effective output impedance of the sending end. It may be possible to get by without the buffer amplifier if the attenuator is placed at the receiving end. The 4-to-20 mA current loop, line 4, Figure 3, grew up in the process-control industries where transmission distances between senders and receivers often are measured in terms of miles instead of feet or inches. The 4-to-20 mA current loop, or simply the current loop, was invented to solve the noise problems attendant with such long lines. The current loop has found a niche in electrohydraulic control systems when noise immunity is desired and/or long distances are involved. A current loop uses the concept of duality compared to the voltage method of data transmission. Using duality, the sending- device output stage contains a current amplifier rather than a voltage amplifier. Current amplifiers have high output impedances, usually in the kΩ range. In contrast, the ideal receiving device has zero input impedance. The reader is encouraged to compare this scenario with that of the ideal sender and ideal receiver for voltage transmission.
V
Data
Time (a)
Fixed-frequency carrier
V
Time (b) Local oscillator
Sensor Transducer input stimulus (DC analog)
Phase-sensitive Analog demodulator
(c)
SCDSAM signal
In practical terms, zero impedance is an unachievable goal, so designers must settle for low impedance. In the case of servo and proportional valve amplifiers, the practice is to output the input stage of the receiving amplifier with a 250- or 500-Ω resistor which converts the incoming current signal into a proportional voltage — what the amplifier was designed for. Note then, that Ohm’s Law tells us that with a maximum incoming signal current of 20 mA, the effective input voltage at the receiving amplifier is 5 V with the 250-Ω resistor and 10 V with a 500-Ω resistor, just about ideal. On the other hand, the minimum input current by standardized practice is 4 mA, which converts to 1 and 2 V, respectively, for 250- and 500-Ω input resistance. One of the disadvantages of the current loop is that the incoming signal does not pass through zero although it may be necessary to have a condition that corresponds to zero. With the current loop, system zero must correlate to a non-zero input current. The usual practice is to split the current spread down the middle, making zero for the system coincide with 12 mA of input current. Now, the input signal is 12 mA ±8 mA. To compensate for this, the receiving amplifier can be equipped with a zeroing, or offset, or
Fig. 4. Note that fixed-frequency carrier signal (b) of suppressed -carrier double-sideband AM (SCDSAM) transmitter undergoes a 180˚ phase reversal each time modulating or data signal (a) undergoes a sign change. This phase reversal required a phasesensitive demodulator, (c), to correctly recover the algebraic sign of the original data. Local oscillator outputs fixed-frequency, fixed-amplitude AC; circuit outputs DC with reconstructed algebraic sign.
bias control. Then, when the incoming current is 12 mA, the receiving system, ie the input to amplifier A, Figure 3, is zero. Of course, the sending end must be set so that the 12 mA output corresponds to the system zero condition. This may be done with the sending-end currentgenerating source. The major disadvantage of the current loop is at the same time its major advantage. The data range is standardized from 4 to 20 mA so that if the signal should ever go outside that range, it is an indication of a fault. The most unreliable part of any electronic equipment is the interconnecting cabling and the terminal connectors. If the data link in the current loop breaks, for example, the current at the receiving end goes to zero which is outside the permissible range. A simple op-amp detector circuit can sense this and take action to shut down the critical parts of the circuit or system. This probably is the one most important characteristic that has kept the current loop in continuous use for more than half a century. Several devices can be on the receiving end. Certainly, if the purpose of the system is to merely measure a system variable such as shaft speed, the data may be observed on any one of several possible data display devices. The most
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CONTROL NETWORKS
To Encoder E2
Twistable shaft Encoder E1 Ti Fig. 5. Twisting the shaft causes a phase shift between two encoder signals and produces phase-modulated signals E1 and E2.
common ones, Figure 3, include the oscilloscope. analog voltmeter, of digital multimeter (DMM). On the other hand, if the system shown in Figure 3 is a subsystem of a control system, then the component on the receiving end may be a servo or proportional-valve amplifier or an A/D converter connected to a computer or programmable logic controller. Any device that listens to an analog voltage in the range that is being transmitted is a possible and viable receiving-end device. The transmission of data by means of AC voltage invariably brings up the subject of modulation and carriers. In over- simplified terms, the modulation contains the information or data while the carrier is the vehicle that carries the data. The AC branch of the family tree, Figure 2, shows the most common forms for data likely to be encountered in electrohydraulic applications. While the tree is not exhaustive, it covers those used most often: ● amplitude modulation (AM) ● phase modulation ● frequency modulation (FM) ● pulse position modulation, and ● pulse width modulation. The form of the modulation tells how the data is contained. Modulation form may be dictated by the kind of transducer on the sending end or it may be the result of necessary or optional signal conditioning. The form of the modulation dictates the nature of the receiving device because it must be appropriately designed/selected to recover the data. The data recovery process, in general, is called demodulation and its purpose is to separate or extract the data from the
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carrier. The specific circuits that perform the demodulation vary from a simple diode for certain forms of AM to phase-sensitive demodulators for other forms of AM, to FM detectors. The possibilities are legion, but when the concepts of data form are understood, the details of the circuits that process the signals become less important. These details can safely be left to the specialized electronic design engineers. Amplitude modulation, suppressed carrier, double sideband. Suppressed-carrier, double-sideband AM (SCDSAM) data transmission definitely is found in electrohydraulic control systems, Figure 4. There are a number of transducers that must use AC and many of them generate SCDSAM signals: synchros and resolvers, variablereluctance pressure transducers, LVDTs, variable-capacitance pressure transducers, strain-gage transducers when AC is used to power the bridge, and potentiometers when AC is used as a supply voltage. The SCDSAM signal is characterized by the fact that the data envelope, Figure 4(a), modulates the amplitude of E1
T
T
Time
(a) T
E2
T
0
0
(b)
Time
(c)
Time 0
E11
E22
0
0
(d)
Time
E0 0 T
0
(e)
T
0 Time
Fig. 6. Reference signal E1 (a) compares with phase- modulated signal E2 (b), can then be converted to pulse position modulated signals E11 (c) and E22 (d), and then to pulse width modulated signal Eo (e), using circuit in Fig. 7.
the carrier, Figure 4(b), as in the case of simple AM, but the envelope and the data pass through zero at the same time. Note that when the data passes through zero, the modulated signal undergoes a 180° phase change. Simple rectification fails to properly reconstruct the algebraic sign reversals in the data. A more sophisticated device, a phase- sensitive demodulator is required, Figure 4(c). This demodulator compares the raw transducer output to the local oscillator signal to maintain full sign sensibility. Of course, once the signal is demodulated, there is no evidence of the internal AC signal in the output. There are a number of commercial transducers that use internal AC, but with miniaturized onboard electronic signal conditioning, this is invisible to the user. Lest an incorrect impression be gained, it is emphasized that the synchro and resolver are potential candidates for phase-sensitive demodulation as described above, although they are not used this way in a commercial sense. The reason they are not is because imperfections in the demodulation process would destroy the otherwise excellent repeatability of the synchro family of transducers. Instead, they are used in synchro-synchro or resolver-resolver pairs in which the second of the pair is at once a receiving device and a demodulator. Additionally, resolvers have been applied in fluid power applications when the output is conditioned by a special resolver-to-digital converter. The result is a rotary positionmeasuring transducer whose performance equals or exceeds that of digital shaft encoders. Phase modulation. As its name implies, phase-modulation data forms are those in which the information is contained in the amount of phase shift between a reference signal and an output or data signal. Phase-shift modulation is used in certain ultrasonic flowmeters and torque transducers. In the torque transducer, imagine that there is a rotating shaft capable of producing a measurable amount of twist in the presence of torque, Figure 5. Output signals are shown as being nominally sinusoidal, Figure 6(a) and (b), but they need not be. In fact if the encoders are optical-incremental types,
CONTROL NETWORKS
Referance signal
E1
Phase modulated signal
E2
Zero crossing detector
Zero crossing detector
Counter is busy
D1
C R
E11
D2
C R
C1 S
Q
R
Q
R1 PWM AND
RS flip-flop
R2 D2
Reset Digital counter
1 MHz Parallel crystalcontrolled digital data out clock
E22
Fig. 7. This circuit converts phase-modulated signal E2 of Fig. 6 into PWM signal.
the outputs are essentially square waves. Assuming that as the shaft turns, the incremental encoders initially are aligned so their respective outputs are directly in phase when there is zero torque transmitted through the shaft, the signal from Encoder 1 will be in phase with the output of Encoder 2. Now, if there is a torque transmitted through the shaft, it will undergo some amount of wind-up or twist. The windup means that the output encoder gets slightly behind the input encoder and can be seen as a phase shift between E1 and E2. The degree of phase shift is directly relatable to the amount of transmitted torque. The output voltage of Encoder 2 is a phase- modulated signal because the data is contained in the amount of phase shift between reference E1 and the output. To extract the phase data, modulated signal E 2 must be compared to reference signal E1; by itself, voltage E2 is meaningless. Demodulation of the phase-modulated signal is easy once it is converted to a PWM signal. The first step in the conversion process is to put reference signal E 1 from the torque transducer just discussed into a zerocrossing detector, Figure 7. A zero-crossing detector is merely an op-amp comparator that produces a positive-going output anytime the input is greater than zero, plus an RC network at the output to differentiate the square wave. The differentiated signals are E11 and E 22, Figures 6(c) and (d), and appear as short-duration spikes. The E 11 spike, having been derived from the E1 reference signal, sets an RS flip-flop causing Q output Eo, Figures 6(e) and 7 to go high. Some time later, E 22 from the phase-shifted or modulated signal provides a spike that resets
D1
Ri
+ 0A
-Vcc R
Rf Data is ready when this voltage is positive
Digital display Computer bus
Fig. 8. PWM digitization is done with a counter that accumulates clock pulses when the PWM signal is high. Value of digital data is proportional to the PWM on-time. PWM is high or positive when on.
the RS flip-flop. The result is that output voltage Eo is high only for the interval between the time the reference goes through zero until the modulated signal goes through zero. The result, Eo, is a PWM signal in which the on-time is a measure of the torque transmitted by the torque shaft. Provided that the shaft is turning sufficiently fast and that there is a large enough number of output cycles from each encoder, the PWM signal will be a high frequency. It is not difficult to calculate that frequency for any given set of hardware and operational scenario. Now, given that the PWM signal has a sufficiently high frequency, any analog readout device such as an analog voltmeter can read the torque because the average or DC level of the Eo signal is directly proportional to the transmitted torque. This example has dealt with a torque transducer, but the principles apply to any phase-modulated signal. Note that the phase-shift torque transducer cannot be calibrated statically; its shaft has to be turning to produce a phase shift. This is a disadvantage that does not apply to straingage torque transducers. Digitization of a PWM signal. Digitization of a PWM signal is easily accomplished with a high-speed electronic counter, a precision stable clock pulse generator (clock), an AND gate, and some other analog elements to take care of handshaking, Figure 8. The clock is any stable, electronic oscillator of the type used in computers or electronic counters. The higher its frequency, the greater the resolution of the digitized output, but only if the counter has enough stages (digits or bits) to accommodate the resolution. The AND gate, Figure 8, is the basic
logic element. Assuming that the PWM signal is high when it is on and that the frequency of the precision clock is substantially higher than the frequency of the PWM signal, the counter will accumulate counts only when the PWM is on, precisely the result we want. At the conclusion of each of the on intervals of the PWM signal, the counter contains a count that is directly and linearly proportional to the width of the pulse. Note that if we would like to have, say, a resolution of 0.1% or 1 part in 1000, that specification fixes the minimum relationship between clock frequency and the frequency of the PWM voltage. That is, it will be necessary for the clock frequency to be at least 1000 times greater than the PWM frequency. By way of example, suppose we wanted to digitize the PWM signal that powers a certain proportional valve whose PWM frequency is 400 Hz, a common value. This requires a clock frequency of at least 400 kHz, not difficult with today’s electronics. Given these values, see that when the on-time of the PWM voltage approaches almost the entire period of the PWM cycle, the maximum total count in the counter approaches 1000 which gives the desired 0.1% resolution. The digitizer of Figure 8 is not atypical, in that it has to carry out some handshaking between it, the digitizer, and the thing that is going to interpret the counter. In the circuit at hand, during that time when the PWM signal is high, the counter is accumulating counts and the digitization is in progress. If the interrogation device (not specifically identified in Figure 8) were to read the counter before the PWM voltage went low, the reader would get an erroneous value. The only
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CONTROL NETWORKS
V
Time (a)
V
Time
(b)
V
Time High frequency
Low frequency
Center frequency (c)
Fig. 9. The fixed-frequency, fixed-amplitude carrier signal (a) shifts about the center frequency by an amount that depends on modulating data signal (b). Combined signal (c) is the modulated output signal.
time that valid data exists in the counter is the interval during which the incoming PWM signal is low or at zero. Note then, that the PWM signal itself contains all the information necessary to cause the handshake. If when ready to receive data, the reading device first looks at the Counter is busy line, Figure 8, and sees that it is high, the reader then knows the counter is being updated and data is not ready. On the other hand, if a poll of the Counter is busy line reveals a low or zero state, the reader knows it can read the counter to obtain a valid digitized value for the PWM input. In today’s control-system technology, the reading device is probably a digital computer or programmable logic controller, which in turn, is being controlled by software. The software would be the vehicle that polls the Counter is busy line. Polling is a term applied to the process of an interrogating device wherein it looks at a given data line and checks for a special condition — in this case the special condition is a single bit, but it could be an entire nibble or byte. The reader is looking for a
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single bit to be low to start pursuing the data. The reader looks and does not find the requisite condition for data acquisition and then looks again, and again, and again and keeps on. Whether the computer looks and then looks again, or optionally goes off and does something else (payroll, turn on the TV, or whatever else may be the proper domain of the computer) is strictly a programmer’s choice and the possibilities are all application specific. The circuit of Figure 8 only points out the concept of handshaking, and that handshaking is a requirement in all digital data communication. Occasionally the circuit may fail. For example, suppose that the reader polls the Counter is busy line and finds its low. This indicates that the counter now holds valid data. But if during that very short time interval between polling and subsequently accessing the counter, the reset pulse should occur, it will set the counter to zero and the reading will be in error. There are ways around this problem but the circuit immediately becomes more complex. For example, one ploy is to make the polling and reading processes take place simultaneously. Then the test can be made on the Counter is busy bit and if not low, the counter-reading discarded and the process continued until the busy bit does go low. Another way is to build a second register to hold the counter contents. The occurrence of the reset pulse shifts the current contents of the counter, bit-for-bit, into the second register, and then resets the counter. This simplifies the handshaking logic because now, the second register always contains a valid number; it is merely out of date by as much as one complete PWM cycle. A fatal flaw remains in the circuit that can be recognized when the reader attempts to read the second register at the same instant it is being updated. The solution for this problem is to use the reset pulse, the output signal from diode D1, Figure 8, as the Data-busy bit. When it is high, the data is invalid. Generally, the number of problems is extensive and their solutions are the rightful province of digital electronic designers. Frequency modulations. Using FM, the information is contained in the
frequency of the signal rather than its amplitude. The two versions of FM are: ● deviation from a center or carrier frequency, and ● carrier frequency is the data. The first is the method employed by the FM broadcast industry, and is not specifically used by the electric or fluid power controls industry. The second version, however, is very common in industrial controls. FM, deviation-from-center frequency. The typical FM radio transmission system, Figure 9, has a local oscillator at the transmitter site that generates a fixed-and-controlled frequency carrier between 88 and 108 MHz for the FM broadcast band. At the time the data (audio signal) is of zero amplitude, an FM modulator in the transmitter sends the carrier signal to the antenna at the same frequency as generated by the local oscillator. When the audio signal goes positive, the modulator causes the transmitted frequency to increase commensurately and conversely, when the data voltage goes negative, the transmitted frequency decreases below the basic carrier frequency. This is a difficult thing for some to visualize because the carrier is an AC voltage of very high frequency and the audio also is an AC voltage but of much lower frequency. Because the Unknown frequency AND
Data counter registers
Master reset
Million-to-one One-second counter/divider gate 1MHz crystal controlled clock
Fig. 10. Block diagram of a digital frequency meter. The Master reset forces the Data counter registers to all read 0. It also sets the Million-to-one counter/divider output high (1 level), gating the unknown frequency through the AND gate to the Data counter registers. One million cycles later, the One-second gate goes off with the Data counter registers holding the unknown frequency.
CONTROL NETWORKS
FM modulator causes the transmitted frequency to rise and fall as the modulation signal changes amplitude, it is the frequency of the transmitted signal that undulates rather than its amplitude. Furthermore, it is difficult to sketch an FM signal that is clear and yet near scale. Figure 9 portrays a typical square-wave data modulation signal rather than a typical audiomodulation situation. Note that when the data is at zero amplitude, the modulated signal is at the base-carrier or center frequency. When the data voltage is positive, the modulated signal frequency is some amount greater than the center frequency; when negative, the modulated signal is decreased in some amount from the center frequency. The demodulator in the receiver must extract the data by detecting and measuring the amount that the received frequency deviates from the center carrier frequency. It is unimportant to know how that is done. FM, carrier is the data. In a popular rotational speed- measuring scheme, the shaft is outfitted with an incremental encoder. Encoder type is unimportant here. Because the measured shaft and the encoder shaft are physically coupled, output frequency of the encoder can be directly related to the shaft speed by an integer amount. In this system, the data is in the frequency and when the encoder has 60 pulses/revolution, output frequency in Hz is exactly equal to shaft speed in rpm. Without doubt, this is an FM system of data transmission. A common demodulation method is to simply input the FM signal into a digital electronic frequency meter, Figure 10, and read the speed. Digitization of frequency. As stated earlier, frequency generation is a popular method of determining shaft rpm. The basic transducers are incremental encoders of the magnetic, Hall effect, or photo-optical encoding variety. The output of these devices are often referred to as being digital signals because they have only high or low values. This is erroneous, except to the extent that the signal is one bit of digital information. In speed or frequency measurement, it is the frequency value that is sought. Frequency is an analog quantity and if it is to be interpreted by
Master reset
S R
FF1
Q Q
Master reset flip-flop One-shot multivibrator
Schmitt
Unknown trigger input signal
One-shot multivibrator
One-shot multivibrator
Schmitt trigger
1MHz crystal controlled clock
AND 1
Main gate flip-flop R Q T FF2 Q S
AND 2
Data ready
Data counter/registers
a digital controller, the frequency must be digitized. Electronic counting technology is the usual method of digitizing frequency. The electronic, digital frequency meter, or electronic counter as it is often called, uses a precise, crystal- controlled 1-MHz oscillator as its clock. Clock output is fed into a counter that can count 10 6 and then reset to zero. Thus, the output of the million-to-one counter/divider produces a gate that is exactly 1-sec long. It also is called a frequency-divider circuit because it accepts the clock frequency and outputs a frequency that is one-millionth of its input. Commercial counters use other divider ratios that are usually panel selectable. The example of the millionto-one is convenient, because it produces a 1-sec gate. The Master reset, Figure 10 is a bit that is set (Usually from low to high) to start the digitization process. The Master rest all at once causes the counter/divider to go high and sets all the registers in the data counter to zero. The counter/divider output remains high until one million output cycles have been received which happens to coincide with 1-sec (1 MHz clock frequency). This precise, 1-sec gate is one input to a two-input AND gate, Figure 10. The one-second precision gate opens the AND gate so that unknown frequency cycles can be directed to the data counter. The count in the data counter goes up by one digit each time a pulse is received from the unknown frequency signal. After 1 sec has elapsed, the gate signal from the counter/divider goes to
Fig. 11. The period measurement function determines the number of clock cycles that occur in one cycle of the unknown frequency input. This strategy maintains resolution without increasing data acquisition delay when the unknown frequency is low.
zero and stops any subsequent unknown signal pulses from reaching the data registers. At that instant, the count in the data registers is frozen with the unknown number of signal pulses that have entered the data counter in 1 sec: the number of unknown cycles in 1 sec. The frequency is now contained in the counter registers. Commercial versions of electronic frequency meters usually provide means for changing the internal gate time. That is, a 1-sec gate and 1-MHz clock frequency are convenient for mental arithmetic because the combination produces a measure of frequency measured in Hz. An important disadvantage of this frequency measurement method, however, is that there is always a 1-sec delay between when the data acquisition process is initiated by the Master reset and when data is ultimately available. This may be prohibitively slow. In fact, in most electrohydraulic feedback control systems, a 1-sec data acquisition delay will render the control system useless. To get around this problem, a shorter gate time, say 0.1 sec instead of 1 sec is used. Now, there is only a 100-msec delay between Master reset and the availability of valid data. A shorter gate time produces further decreases in delay time, but resolution is lost unless clock frequency is increased. Note that with a 0.1- sec gate, if the unknown frequency were, for example, 5 KHz, the data counter would accumulate only 500 counts in the gate time interval. The resolution of the digitization process would be 0.2% of reading. It is
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CONTROL NETWORKS
Fig. 12. Block diagram, (a), of a frequency-to-voltage converter produces analog DC output that is proportional to the input frequency but is delayed in time.
ei Frequency input
e1 Schmitt trigger
e2
Low-pass filter
(a)
e3 Analog output voltage
High frequency
ei
Medium
Low frequency
frequency
easy to see that another decade of reduction in gate time would produce a tenfold increase in the resolution. This e1 constitutes a random error in the measurement of the frequency. The only way to maintain resolution is to e2 Tos increase clock frequency as gate time is reduced. But note that a 100-fold decrease in e3 gate time must be offset by a 100-fold increase in clock frequency to maintain digital resolution at a target value. Period digitization. An alternate method of digitizing a frequency signal is to use its period instead of its frequency. Many commercial electronics counters have the ability to switch from frequency to period measurement, but plug-in cards for digital computers and programmable logic controllers may not be so equipped without extensive software changes. It is true that as the frequency goes down, if one is to maintain a given digital resolution, the gate time must be increased if the clock frequency cannot be changed. In most counters, clock frequency cannot be changed, but gate-time change is commonly available. The result is that as the frequency to be measured and digitized gets lower, the delay time in data acquisition must increase; this quickly deteriorates into an untenable situation. To circumvent the problem. it may be possible to measure the period of the incoming signal wherein the reciprocal of the frequency is digitized. That is the number of seconds or msec per cycle are measured rather than the number of cycles per sec (Hz). To do so consists basically of swapping the unknown frequency of Figure 10 with the 1 sec gate, Figure 11. There, both the unknown in-
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One-shot multivibrator
Very low frequency
(b)
(c) Tos
Time
Time Tos
(d)
Time Average value of e2
(e)
Time
put and the onboard precision 1 MHz clock are passed through Schmitt triggers to square them off. For period measurement, it helps to realize that the clock frequency is much higher than is the frequency of the incoming unknown. In fact it is absolutely necessary that that be true. Now, during one cycle of the unknown, several hundred or thousand clock pulses will pass. Nothing happens until receipt of a Master reset. Master reset is latched up into its own flip-flop. The Q output is one input to AND1 and arms it for the next pulse to come from the unknown input. Note that the Master reset also goes to reset the data counter/register at the same time it arms AND1, and Master reset also resets the Main gate flipflop. With AND1 armed, the very next unknown input pulse causes the Main gate flip-flop to toggle from off to on. The second unknown input impulse toggles the Main gate flip-flop from on to off. In this manner, there is a voltage at the Q output which is high for one complete input period and then goes low thereafter. When FF2 switches from high to
low, note that its Q output switches from low to high. The leading edge of the Q signal is detected with the oneshot multivibrator whose output goes to reset the Master reset flip- flop. This disables the entire measurement process until another Master reset signal occurs. However, after Master reset, the very next unknown input pulse is directed to the Main-gate RTS flipflop, FF2. With its Q output high, AND2 is armed, allowing subsequent shaped clocked pulses to increment the Data counter/register. Eventually, another input pulse is received which can pass through armed AND1 and cause the Main-gate RTS flip-flop to toggle to its reset state. It is at that instant, as explained before, that the leading edge of the Qoutput is sent to reset the Master reset flip-flop which locks out any further action until another Master reset. At that time, the Data counter holds the number of microseconds that elapsed during one complete cycle of the unknown input wave. That, of course, is the period of the unknown. The interrogating device needs only to look at the Qoutput of FF2 and whenever that voltage is high, valid data is in the Data register. Note that the total elapsed time between the receipt of the Master rest and the availability of data is not more than two complete cycles of the unknown input signal. The disadvantage of this system is that the amount of data acquisition delay is a function of the period of the data. This can be a problem when attempting to improve the responsiveness of a speed-control system. Frequency-to-voltage conversion. If one is interested in remaining in the analog world instead of converting to the digital world, it is possible to convert the frequency of interest into a proportional voltage using the frequency-to-voltage conversion process. It is purely an analog process and like its digitizing cousin, there are time delays involved. These time delays may be invisible to the unwary, and therefore more insidious. An incoming frequency signal is first fed into a Schmitt trigger, Figure 12(a), to shape the arbitrary input waveshape into a square wave.
CONTROL NETWORKS
H
Ac power lines (high side) AC load In L Ac power lines (low side) CLA(parasitic) CHA (parasitic) Electronic amplifier A
+
+ -
Zi
ein
-
In
V0
Ae in C Amplifier circuit common
In E
CHE (parasitic)
Optional connection
Mother earth
Fig. 13. Parasitic capacitance exists merely because of circuit construction and proximity of one circuit to another. There is parasitic capacitance CHA between power line high side H and amplifier terminal A. Due to the potential difference, small noise current Iin travels through CHA, amplifier impedance Zin, and returns to mother earth and the power company through parasitic capacitance CCE. The voltage drop across Zin, due to Iin, is amplified and appears at the output of the amplifier as 60 Hz noise voltage. The amount of noise in Vo depends on physical distance, Zin, and amplifier gain A. Note that if the optional earth connection is used, total parasitic impedance diminishes, raising I in, creating a higher noise voltage.
In general, incoming signal ei, Figure 12(b), will vary in amplitude and frequency, but it is the frequency data that is of interest for conversion. Note that the Schmitt trigger not only squares the incoming signal but also generates a constant amplitude while the frequency is unaltered, Figure 12(c). The one-shot multivibrator is the device that really does the frequency-to- voltage conversion. Recall that the one-shot multivibrator receives an incoming pulse (it detects the leading edge of the incoming sig-
nal) and produces an output of exactly the same frequency, although on-time duration [T os, Figure 12(d), of the output pulse is the same from cycle to cycle. As subsequent pulses of fixed on-time get closer and closer, the average value of the voltage rises. As pulses get farther apart, the average voltage decreases. In this fashion, the value of analog output voltage e3, Figure 12(e), is linearly proportional to the frequency of ei. Of course, raw output from one-shot multivibrator e2 is filled with pulses. To smooth them, a low-pass filter must be used. This filter has substantial delay and can be detrimental in feedback control systems. There are a number of commercial frequency-to-voltage converters available on the market and before they are used in any feedback control system, the manufacturer should be consulted regarding time delays in the instrument. Published catalog data usually is evasive on this issue. There are models that often cost more money that are faster, but it is impossible to measure frequency without some amount of time delay in data acquisition. Frequency-measurement summary. It should be apparent by now that there is no way to use frequency data to measure speed without time delay. Furthermore, the amount of delay depends on the hardware and in some instances, on the data as well. In feedback control of speed or flow, for example, it may be necessary to use frequency measurement for speeds in one range and period measurement in another range. Regardless, frequency generation remains a popular and cost-effective way to determine speed. Some investigations into the suitability of frequency methods should be made. Note that the tachometric methods of speed measurement do not have these time delays. These methods should be evaluated in critical speed- or flowmeasurement situations. Noise control in electronic systems Electronic system noise control is more art than science, and its elimination is appropriately referred to as a witch hunt. Electronic noise is broadly defined as any unwanted variation in signal. The root of the
term noise harkens back to the early days of telephony and radio when noise literally meant audible sounds through the headset or speaker. Certainly, the term retains that interpretation today and all of us know of the background noise that is associated with certain conditions in telephones and radio. But now, noise has become a more generic term that refers to any unwanted signal regardless of whether it is audible or not. Control-system noise may be audible but it also may give evidence to its presence in an unexpected and/or unwanted twitch or vibration in a positional servomechanism, unexplained variation in manufacturing uniformity, or puzzling vibrations in some output member. The best approach to reducing noise to a tolerable level (theoretically it cannot be eliminated) is to identify its source. Generally, there are four common noise sources that should be considered at design time and at debugging time: ● electrostatic interference (ESI) ● electromagnetic interference (EMI) ● radio frequency interference , and ● ground loop. Identification of the type of noise is crucial because each method of control is different. Most non-electronic engineers who have had to deal with noise find that it is unpredictable and therefore unmanageable. The problem is that in one instance, the source may have been ESI that can be controlled with proper shielding and grounding, and in a later situation the source was EMI that cannot be controlled with shielding and grounding. Yet the symptoms, especially when associated with the 60 Hz power-line frequency, are identical. General rules for controlling noise These recommendations should be followed when constructing any electronic equipment for use in an industrial environment — an environment considered especially noisy: ● connect all electronic components so that the ground, low side, or circuit common all tie together. Never swap leads to change an algebraic sign, which may be necessary to correctly phase a servo loop, for example, unless the output is truly differential. Most
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CONTROL NETWORKS
H
Ac power lines (high side) AC load L Ac power lines (low side)
In
CLA(parasitic) CHA (parasitic)
Rs A Low value Signal ein source
Ins S
+ -
C
Electronic amplifier
+
Zin
- InA
Ae in +
-
Amplifier circuit common In
E
CHE (parasitic) Mother earth
electronic signals are not differential — the are single-ended. On the other hand, there are some situations where neither side of a receiving device is grounded, such as the coils of a servovalve. It is permissible to swap the leads of a servovalve in order the change the algebraic sign of the servo loop. Proportional valve-coil leads can be swapped but only if there is no spool feedback position transducer. ● do not run AC power lines in the vicinity of electronic data lines. The current levels in the AC power lines usually are sufficiently high to cause their magnetic fields to couple into the data lines. If that is the case, electronic devices such as servo and proportional valves and their amplifiers, most likely will respond ● run all signal (data) lines in twisted, shielded pairs ● run the AC power lines in twisted, shielded pairs, too ● consider a 4-20 mA current loop when data transmission distances exceed about 20 feet ● encase AC power lines in iron (not aluminum) conduit. Of course, the conduit is the shield, and ● connect the electronic circuit common to power ground (mother earth) only when necessary and only at one physical point in the electronic part of the measurement network chain.
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V0
Optional connection
Fig. 14. When amplifier input is connected to a signal source that has low source resistance R S , there is a tiny increase in noise current. A significant noise reduction benefit is achieved, however, because lowsource resistance R S diverts the major part of noise current Ins away from high-input impedance Zin. There is less noise voltage across Z in , and consequently less noise voltage at the amplifier output. Therefore merely by connecting a low-resistance signal source to the voltage amplifier of and by itself, can have a 1000:1 noise reduction component leaving the amplifier input open.
When this rule has to be violated because component manufacturers have provided built-in grounding, follow the recommendations in the later parts of this section. Parasitic capacitance and electrostatic interference Capacitance arises any time two conductors are separated by a dielectric. Of course, electronic equipment is built of conductors and those conductors most generally are separated by a dielectric - the insulation and the atmosphere between them. Therefore, all electronic equipment has parasitic capacitance. It is called parasitic because it exists merely because the circuit exists; there is just no way to build a circuit without it. As with hydraulic capacitance that arises because of fluid compressibility, capacitance is parasitic because no one has figured out a way to build hydraulic equipment without having an enclosed, compressible volume of fluid. Parasitic capacitance of electronic circuits can be the cause of induced noise, especially at power-line frequencies. Figure 13 shows a representation of how the power lines, the parasitic capacitance, and an electronic device (an amplifier) interact to create a noisy system. Note and understand that the capacitances in Figures 13, 14, and 15 represent real capacitances that exist
and can be measured, but that they do not represent capacitors that have been hard-wired into the circuit. These capacitances are not pieces of hardware; the circuit must be viewed as an analytical schematic rather than the usual hardware schematic familiar to the fluid power practitioner. The capacitances exist only because one has found a way to build the circuit without this type interaction. In Figure 13, the power company is represented by the AC alternator in the upper left-hand corner, the motherearth grounding connection in the lower left-hand corner, and the power lines and arbitrary load across the top. The amount of capacitance is very small, leading to a very high capacitive impedance, perhaps as high as 109 Ω. Amplifier input impedance, Zin, may be about 100 kΩ. Using the 109 Ω estimate for the parasitic impedance, and line voltage at 110-V RMS (about 311 V peak-topeak), there will be an AC current, the result of 311 V divided by 10 9 Ω, or about 0.311 µ A. That current, upon passing through the 100 kΩ input impedance will develop a voltage drop of about 31 mV (using Ohm’s Law). Depending on the gain of the amplifier and H Ac power lines AC load
In L
CLC
CHC C
C
CCA
CC0 Electronic amplifier Zin A
A
In
0
V0 M CMC
C
In
Optional amplifier-tocase shortening commection Mother earth
Fig. 15. When the amplifier is enclosed in a conducting case, noise current In, due to power line-to-amplifier parasitic capacitance is made to totally bypass Z in, thus perfectly nullifying its effects. In doing this, however, an unwanted feedback path is created through CCA and CCO. All capacitances are parasitic.
CONTROL NETWORKS
a number of downstream factors, the induced noise voltage may be palpable. This phenomenon is most dramatically demonstrated with an ordinary oscilloscope with the vertical sensitivity set to about 1 V/cm or maybe a little less. Connect a lead onto the high side of the scope input and hold the lead in hand while observing the trace. It is not unusual to see a deflection of several V. The amount of deflection depends on the wiring in the vicinity of the scope, and the induced noise voltage is definitely increased by the presence of fluorescent lights. In analyzing this situation, the lead and the human hanging on to it are sometimes referred to as an antenna. This is technically incorrect because an antenna picks up an electromagnetic wave, whereas this phenomenon is CCAI
CCA0
Electrically the case is here Electronic amplifier
A e0
ei Circuit-board common
Fig. 16. Because of the encasing shield, parasitic capacitance CCAI and CCAO create a feedback path that can cause the high-gain amplifier to break into oscillation. Deliberate connection
CCAI
Electrically the case is here
CCA0
Electronic amplifier
A ei
Circuit-board common
e0
Fig. 17. Connecting the circuit board common directly to the enclosing case kills the parasitic feedback and reduces the tendency for the amplifier to break into oscillation. It does not, however, kill parasitic capacitances CCAI and CCAO.
electrostatic. The huShielding man is one plate of a caCase Case case pacitor and touching Transducer Receiver/ Amplifier the wire puts that plate read out High in touch with the lead Load A and a human has more capacitor area than the Low wire. Therefore when common dropping the wire, the Twisted amount of scope deflecpairs tion decreases. This is Mother earth similar to the reaction when one touches the Fig. 18. Ideally, the electronic network chain is totally enantenna of an AM radio cased in conducting shields and enclosures. All shields and tuned to a weak station. cases must be electronically continuous and must be conThe audio level in- nected to the shield at one or more points. Finally, if a creases but in this case nother-earth connection is to be made, it must be made at the events have a differ- only one point. ent origin. To test whether the induced noise little voltage drop, and there will be is a capacitive (electrostatic) effect or no apparent noise on the scope signal. an electromagnetic effect, one need Even when you touch the high-signal only connect the low side of the oscil- side of the scope, the noise deflection loscope to mother earth. If ESI is the will not be measurable. Bear in mind culprit, the amplitude of the noise on that this would not be true if source the scope will increase, assuming that impedance R S were large, say of the the low side of the scope input is not same order of magnitude as the scope connected internally to mother earth. input impedance. This further argues This is because parasitic capacitance, in favor of low- impedance sources as C CE, has been shorted out, reducing being the best matchup for high inthe total capacitive impedance, in- put- impedance voltage amplifiers. creasing the current, increasing the voltage drop across Zin, and resulting Electrostatic shielding in a higher voltage. Were the effects To counteract the electrostatic efcaused by electromagnetic induction, fects of parasitic capacitance, the amgrounding the scope to mother earth plifier is placed inside a conducting would have no effect. case called a shield, Figure 15. The amTo carry this investigation a bit fur- plifier, literally, is placed inside a box. ther, in Figure 14, the scope (the elec- Note that the box need only be a contronic amplifier) has been connected to ductor; it does not have to be magnetic. some source. Like all sources, it has Immediately, the shielding case prooutput impedance R S. Recall that the vides a direct short around the ampliideal voltage source has low imped- fier so that noise current due to paraance; if that be the case, RS might be on sitic capacitance completely bypasses the order of say, 10 Ω. Immediately the the amplifier input terminals. The detriscope will show that all the noise volt- mental effect of power line-to-ampliage is gone! The reason can be seen fier capacitance is absolutely canlooking at the path that noise current In celled. That’s the good news. must take from the high side of the The bad news is that now there is a power-company line to mother earth. substantial parasitic capacitance beNote that in passing through CHA and tween the output terminal of the amentering the scope lead, the noise cur- plifier and the large area of the enrent can take one of two paths: it can go closing case. Plus there is a parasitic through the high impedance of the capacitance from that same case to scope (usually about 1 MΩ) or it can go the input terminal of the amplifier. through the low impedance of signal This capacitance forms a feedback source RS. path from output to input, Figure 16. Clearly, the current will go through High-gain amplifiers, of which the signal source, it will produce very servo- and proportional-valve ampli1998/1999 Fluid Power Handbook & Directory
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CONTROL NETWORKS
Power company alternator
User's appliance (a)
Mother earth
Mother earth
Power User's company appliance alternator Break in low-side line (b) Current path
Fig. 19. Because there are earth connections at the alternator as well as at the user end (a), a user’s appliances will still function if there is a break in the low-side line because of the earth’s conductivity (b). In that case, however, appliance operation will most likely be at less than optimum performance.
fiers are a part, are known to break into oscillation due to parasitic feedback. Oscillation frequency is several hundred to a few kHz. The symptoms are that the system responds sluggishly, and increases in gain do not improve responsiveness. The insidious part of this malfunction is that the effects cannot be seen with any DC measuring instruments; an oscilloscope is the only way to diagnose the problem. Adding to the insidiousness is that the response of the system mimics the performance of a contaminated servovalve. There is a way to prevent the oscillation problem, but the contamination problem is a matter for other study. Killing the parasitic feedback path. Addition of the enclosing case got rid of the line-to-amplifier parasitic coupling but it created a new parasitic path that can lead to amplifier oscillation. That problem can be eliminated by making a deliberate connection from the common side of the amplifier board (card) directly to the encasing shield, Figure 17. This connection couples the input and output parasitic capacitances that are shorted to circuit common. Note that this is not mother earth here; this only circuit-board common, the return point on the card, the center tap of the power-supply transformer. Now observe in Figure 17, that any current that enters CCAO due to output voltage Eo is
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directed to common, rather than into CCAI. This means that the parasitic feedback path has been effectively killed. The rule is simple: always place all electronic devices in a totally enclosing, conducting case and then connect the circuit board common to the case. Shielding the electronic network chain. To prevent undesirable effects of parasitic capacitance, it is necessary to enclose the entire electronic network chain within the shield, Figure 18. Each element in the network is placed within its own box or case and the cases connected together through the interconnecting shielded cable. The nature of the shielding is not critical from a technical point of view: it only is necessary that the shielding be a conductor. When each electronic component is in a separate case, a physical connection must be made from the cabling shield to the cases at both ends of the cable. Note that if a connection to mother earth is made, it should be at one point only.
somewhere between the alternator and the user’s site, current still can be delivered to the user’s appliance. However, it most likely will be reduced because of the impedance of the earth path. With good earth conductivity, the user may not even detect the fault. The aim here is to show that under not-so-rare circumstances, substantial current can be in the earth at a given point because of ground wire faults. A ground loop exists whenever an electronic circuit is constructed so that the earth-borne current can find its way into the electronic-shielding system, Figure 20. All it takes is connection of the electronic system shielding to mother earth at more than one point in the electronic network chain. Assume that this system is set up in a region where there are ground currents, I g . Note that circuit common is connected to the shielding at points A and B, and that there are also Earth connections 1 and 2 as well. Although there is current Ig in the earth path, the earth is an imperfect conductor. As a result, when the current arrives at the site of the first connection it is offered two paths: continue through the high impedance earth, or follow the very low resistance of the shield. The lower-resistance path will carry the bulk of the current. Regardless of how low the shield’s impedance may be, it is not zero; there is some resistance Rwire. Ground currents can be on the order of a few amperes to hundreds
Ground loops — cause and control Electric power companies always connect one terminal of their alternators to mother earth for safety reasons. The side of the alternator connected to earth is called the low side of the power line. The low side also is carried by a physical wire between the alternator and the user site, Figure 19(a). Because the earth is a conductor, the low side of the line actually has some redundancy to it. That is. it is possible for the Shielded twisted pairs low wire to break, Case Case Case Figure 19(b), and Receiver/ Transducer read out Amplifier yet a user’s electrical apparatus ea A Load will work beIg cause one side of +R the appliance is B wire A likewise conEarth Earth nected to mother connection 1 Ig connection 2 earth. Of course, Rg Ig Ig the electrical wiring codes require that the Fig. 20. When improper practice has been followed and two user’s site be con- connections have been made to earth, ground currents prefer the shielded path with its lower resistance rather than earth renected that way. Note that if the sistance Rg. Current through the shield’s resistance creates a low side of the small voltage drop that will be amplified, creating 60 Hz noise power-transmis- in the electronic signals. The electronic chain should never be sion line breaks connected to earth ground at more than one point.
CONTROL NETWORKS
CAL R0I L
A
OIT
OIR
Rwire Ig
Fig. 21. Opto-isolation eliminates the ground loop problem even when receiving and transmitting ends are grounded because the majority of ground current I g passes through the relatively low resistance of earth path Rg rather than very high optoisolator resistance ROI. The ground- loop problem is thus solved with opto-isolators.
of amperes depending on local conditions. This shield current will create a voltage drop that will be felt at the input of the amplifier and will be amplified and sent down the electronic network chain. The obvious correction to the ground loop problem is to eliminate one of the earth connections. Then any ground currents are forced to remain in the earth where they belong. Understand that electrostatic noise and ground-loop noise will generate signals in the electronics that are at line frequency, and it will not be apparent at troubleshooting time which type noise prevails in a given situation. Furthermore, if only one earth connection has been made but some manufacturers have internally grounded their equipment, another ground connection may be in place either through the electrical wiring or even through hydraulic and/or water plumbing. Do not assume that the only ground connections are those you have made or that are plainly visible. Opto-isolation Opto-isolation makes use of light waves to interconnect a sending and a receiving device. When ground-loop noise is severe and differential amplifiers are not practical, it may be necessary to optically isolate one component from another, Figure 21. For unstated reasons, the electronic network has two points connected to mother earth. OIT and OIR are the optical transmitting and receiving de-
vices, respectively. The transmitter usually is an LED whose intensity is affected by the amplitude of its current. The receiver is a photo transistor whose collector current is affected by the light intensity impinging on its exposed base materials. In this way it is possible to couple data from the sender to the receiver without a direct electrical connection. The advantage of this is that there is an extremely large impedance offered to the ground current so that it prefers to stay within the earth path and not be diverted to the shield. Electromagnetic interference Current in one circuit creates magnetic flux which links with the magnetic flux of another circuit to produce electromagnetic interference or noise. When the rate of change of flux is sufficiently great, a palpable noise can be felt in the linked circuit. There are no easy solutions to EMI. Only use of good design and fabrication practices will help minimize the problem. It is possible to control the receiving circuit to minimize the amount of flux linkage between it and whatever the source may be. This is done by always keeping signal lines away from power lines and always running the signal lines as twisted pairs. The twisting of signal wires: ● reduces the amount of cross-sectional area that the wires encircle and minimizes the amount of flux linkage that otherwise may exist, and ● alternates the direction of each consecutive loop in the wires so that if there is a magnetically induced voltage in one twist, it will be offset by a voltage of the opposite phase induced in the adjacent twist. It is emphasized that twisted pairing of wires is always the best practice. Attempts should be made to reduce the amount of magnetic radiation at its source. For example, the most common source of 60 Hz magnetic radiation is the field created around current-carrying conductors. The higher the current, the greater the amount of magnetic flux and the greater the likelihood that there will be an EMI voltage induced in any electronic equipment in the vicinity. Approaches to reduce this magnetic radiation include:
● twisting all power lines just like signal lines. This helps because the flux caused by the current in one conductor cancels the flux created by the opposite-going current in the mating conductor. This is true even for three-phase power lines as long as all three lines are twisted together, and ● encasing all power lines in iron conduits. This helps confine the magnetic field to the conduit, preventing its linkage with sensitive electronic circuits. Many times there are unencased three-phase power lines in the industrial environment. It is possible to locate them with an ordinary portable AM radio receiver. Tune the receiver to a frequency where there is no signal and then walk around, converging on those places where the 60-Hz noise from the receiver is loudest. Those places are the source of the powerline-based magnetic fields. Radio-frequency interference Radio-frequency interference (RFI) can come from a multitude of sources: ● bona fide radio signals radiated by a nearby radio, television, or radar station. Attempts to nip these generators at their sources are sure to be met with derision if not hostility, and ● RFI that is radiated when electrical contacts open or close with an arc. There is no specific frequency. Instead, the signals are broad band containing thousands of frequencies. Because RFI can be of short wavelength, twisting and shielding signal wires may alleviate the problem, but controlling the problem is not at all a systematic process. One method is to use arc-suppression devices on relay and starter contacts, and to put varistors in series with AC coils to reduce inrush current. In extreme cases, it may be necessary to run separate power lines and transformers from the electrical substation because the short wavelength RFI will sometimes find its way into the electronics through the power wiring. The wiring acts as a waveguide and directs the energy right to the place where it is not wanted. The symptoms of RFI are usually the twitching of a servoaxis or the sudden jump in a data display indicator that correlates with starting a motor or energizing an on-off solenoid.
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Cylinders
A
fluid power system can produce linear force and motion better than any other type of power medium. The fluid power actuator that produces this is a cylinder, sometimes called a linear motor. One of the oldest and most basic of fluid power components, cylinders have evolved over the years so that there is a lot more to them today than meets the casual glance. The principal parts of a fluid power cylinder, Figure 1, are a: Barrel — Made primarily of circular metal tubing with walls whose thickness depends on the fluid pressures that must be contained, recent innovations include new cross-sectional shapes as well as new materials. The tubing ID is finely finished to a microinch surface tolerance to prevent damage to the piston seals as the piston reciprocates within the barrel. Head — The head-end barrel closure has an opening through which the piston rod extends (unless the cylinder is a rodless type.) There usually is a cavity in the head around the rod hole to receive a cartridge that includes a bearing to support the rod. The pistonto-rod bearing distance is a lever arm that dictates how much side loading a cylinder can withstand. On long stroke cylinders, one method to ensure that this distance does not decrease below a preset valve is to include a stop tube in the cylinder. The stop tube, Figure 1, physically prevents a cylinder from extending beyond a preset length. The bearing cartridge, secured separately in the head, also contains rod seals and and rod wipers to control leakage and to stop contaminant ingres-
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sion as the rod retracts. This cartridge usually can be removed for seal/bearing replacement without having to disassemble the cylinder. The concentricity of this cartridge and the cylinder barrel is most important. Cap — The cap is the cylinder barrel closure on the non-rod end of a singlerod end cylinder. The cap and the head contain the cylinder ports, the cylinder cushioning activities (if any), and perhaps integral metering valves. Rodless cylinders often have the cushioning cavity in the piston faces. Piston — Cylinder piston design varies widely depending on factors such as operating medium, design life, and intended use. All have circumferential glands for seals to prevent cross-piston
leakage and many have engineered plastic bands called wear rings which extend cylinder operating life. Piston width and its bearing area also are important in the side loading problem. Rod — A cylinder rod connects the piston to the load. Usually made of high-tensile, chrome-plated steel bar stock, rods are securely locked to the piston at one end while the opposite end frequently is finished with rolled threads and wrench flats; many thread options are available. Some cylinder rods are hollow and act as the conductor for pressure fluid to enter and leave the cylinder. Others are internally splined and mesh with a mating spined rod secured to the cap. This ensures that the rod cannot rotate. Head
Barrel Cap
Piston
Cap end fixed clevis
Stop tube
Rod
Check valve Rod clevis
Cushion cavity Piston seal Cushion spear
Cushion nose
Variable orifice
Fig. 1. Cut-away drawing of double-acting cylinder illustrates basic parts. This particular example shows clevis mounts to accommodate load that moves in arc.
CYLINDERS
Another method to prevent rod rotation uses a rounded triangular cross section for the rod with matching-shaped bearing and seals in the head cartridge. Cushions — At high speeds, the moving mass of the piston, the rod, the load, and even the fluid must be decelerated as the cylinder approaches the end of its stroke. Without controlled deceleration or cushioning, the rapidly moving mass could easily damage the cylinder ends and/or the piston. Cylinder deceleration can be achieved with components external to the cylinder such as a mechanical rod brake or a hydraulic shock absorber. Most hydraulic cylinders are equipped with cushions built into one or both end closures. Internally cushioned pneumatic cylinders often have some sort of elastomeric element attached to Return spring
Inlet port
Vent
Piston and rod
Fig. 2. This single-acting short-stroke cylinder extends pneumatically; retracts because of spring compression action between rod bushing and piston. Other models can be spring extended or have spring exterior to barrel.
Fig. 3. Integrity of basic tie-rod cylinder is ensured by high-strength tie rods which hold head and cap onto barrel. This cylinder has centerline lug mounts which absorb forces along cylinder centerline and prevent torques in cylinder body.
the end closures or both faces of the piston to act as bumpers. Because of the compressibility of air, the pneumatic cylinder cushioning problem is distinctly different from that of a hydraulic cylinder. Hydraulic cylinders most often have a spear or nose on one or both sides of the piston (depending on whether cushioning takes place in one or both directions of piston travel) which fits into a mating cavity in the end closure. As the spear enters this cavity, fluid metering starts between the spear and the edge of the cavity, decreasing the flow of fluid leaving the cylinder. This metering is similar to that which occurs as the spool of a directional control valve begins to close, forming an orifice which meters fluid. The volume of fluid which easily can flow out of the cylinder decreases; pressure between the end closure and the piston increases to slow piston travel. When the orifice between the spear and the cavity closes, a separate passage with a small metering orifice is still available for fluid to leave the cylinder until the piston finally reaches full retraction. When the directional control valve shifts to extend the cylinder, pressure fluid flows through a second passage between the port and the piston face that is fitted with a check valve. This pressure fluid bypasses the cushion to affect the entire face of the piston. The stroke starts rapidly with full force and speed before the cushion spear leaves the cavity. Many spear shapes are available and the fit of the spear into the cavity varies, too. A few of these spear variations include straight, tapered, parabolic, piccolo (with each orifice closing as the spear advances into the cavity) and a spear with graduated steps. This variety attempts to provide rod travel vs. deceleration curves appropriate to the application. Operating principles Cylinder operation is single- or double-acting. Single-acting versions, Figure 2, accept pressurized fluid on only one side of the piston; the cylinder barrel on the other side of the piston is vented to atmosphere. This pressurized
Fig. 4. Mill cylinders are designed and constructed for extremely heavy-duty applications in hostile environments.
fluid may extend or retract the cylinder depending on whether the head or cap end receives the pressure fluid. In either case, the cylinder always returns mechanically to its original state when the air vents or oil returns to tank. This mechanical return mechanism is almost always a spring that has been compressed as the pressure fluid moves the piston. Another return mechanism is gravity acting on the loading cylinder rod. When the cylinder returns because of gravity, it must be mounted vertically. Double-acting cylinders, Figure 1, alternately receive pressurized fluid on one side of the piston, while fluid on the other side returns to tank (or air is vented). To change cylinder direction, the pressurized and vented sides of the piston are exchanged through valving. Double-acting cylinders operate while mounted in any position; they are by far the most commonly used type of cylinder. Because the area of the piston face with the rod is smaller than the full face of the piston, the forces the cylinder can exert when it extends are always greater than when it retracts, assuming the same operating forces. Because total cylinder volume is less with the cylinder fully retracted (because of rod volume) than when the cylinder is fully extended, a cylinder retracts faster than it extends with a given flow. Construction Cylinder manufacturers have devised several construction methods to close the ends of the cylinder barrel so it becomes a sealed, closed envelope. These methods have lent their names to the cylinder, and they are now known as the tie rod or JIC for Joint Industry Conference), mill, threaded end, and
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CYLINDERS
Cap Piston seals
Head
Brass rod bearing, mounting stud Wrench flat
Wrench flat
Rod packing
Fig. 6. One-piece cylinder has barrel permanently attached to ends using doublerolled construction. Cylinder also is example of single-rod configuration.
Fig. 5. Threaded-end cylinders are designed for lower-pressure hydraulic or pneumatic service. Often made of materials other than steel, cylinders of this variety may still be used in severe service conditions for extended periods.
one-piece or welded types. Tie-rod cylinders, Figure 3, have square or rectangular barrel closures held on to each end of the barrel by rods which pass through holes in the corners of the closures. Nuts and lock washers on each end tighten the closures to the barrel to maintain integrity. Static seals are placed in the gland of the barrelclosure interface. The majority of cylinders for industrial, heavy-duty applications use this construction. Variations of this design include use of more than four tie rods on a cylinder, or using a long bolt that threads into a tapped hole in one of the end closures. These cylinders are the only type that are covered by national standards. Mill cylinders, Figure 4, have circular flanges welded to the ends of the cylinder barrels with barrel closures of the same diameter as the flanges. Short bolts secure the closures to the flanges for cylinder integrity. On some large versions, the barrel wall is thick enough for the closures to be bolted directly into the barrel wall. As the name implies, these cylinders were originally designed for use in steel mills, foundries, and other severe-duty applications. There are no national standards that apply to mill cylinders. Threaded-end cylinders, Figure 5, have closures that thread into or onto the cylinder barrel. Most of these have the port in the closure piece. A variation of this uses a locking wire that
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keys the ends and barrel together. One-piece or welded cylinders, Figure 6, have the closures welded to the barrel and are throw-away components because they cannot be disassembled for repair or seal replacement. When used on applications which do not require the continual actuation of the industrial environment, these cylinders may be very cost effective. An alternative method of manufacture rolls the tube into a slot on the closure to mechanically lock the three pieces together. Another alternative design has a welded cap and a head locked into place using threads, lock rings, or some type of key. With the modification, the cylinder is no longer a throwaway because it can be disassembled for repair; this construction also raises the initial cost of the cylinder. Configurations The most common type cylinder is the single-rod end, Figure 6, in which the rod is nearly the length of the cylinder barrel. The rod is visible at only one end of the cylinder and transmits the generated forces to the load. A double rod-end cylinder, Figure 7, has a rod attached to both faces of the piston with the rods extending through both end closures. These cylinders are useful for moving two loads in reciprocity and they also solve the problem of different forces and speeds being generated during
Fig. 7. Double-rod end tie-rod cylinder had rods extending from piston through each end closure. Example also has a head rectangular flange mount.
Fig. 8. Tandem cylinder can double force output at same pressure, because two pistons areexposed to pressure fluid.
the extend and retract strokes. Another cylinder configuration is the tandem cylinder, Figure 8. This is really two single rod-end cylinders connected in line with each piston inter-connected with a common rod as well as a second rod which extends through the head. Each piston chamber is double acting to produce much higher forces without an increase in fluid pressure or bore diameter. These
CYLINDERS
Fig. 9. Duplex cylinder consists of multiple double-acting cylinder sections with rods of each protruding into next section to physically move the piston of that section to produce a number of fixed extension lengths.
Fig. 10. Telescoping cylinders provide long extension capability from short, retracted lengths. Used mostly on mobile applications where room for standard retracted cylinder length is not available.
Fig. 11. Ram cylinder rod has a uniform OD, and except for clearance, the same size as the ID of cylinder barrel.
cylinders necessarily occupy more linear space. Duplex cylinders, Figure 9, are similar to tandem cylinders in that both are cylinders connected in line, but the pistons of a duplex cylinder are not physically connected; the rod of one cylinder protrudes into the nonrod end of the second, and so forth. A duplex cylinder may be more than two in-line cylinders and the stroke lengths of the individual cylinders may vary. This results in a component that can achieve a number of different fixed stroke lengths depending on which of the cylinders and on which end the cylinders are pressurized. Telescoping cylinders, Figure 10, are nearly all single acting, although double-acting versions are available. They consist of as many as five sets of tubing with decreasing OD that nest inside one another. Each tube or stage is appropriately equipped with seals and bearing surfaces. Available for extensions in excess of 15 ft., most are used on mobile applications where available mounting space is limited. The collapsed length of telescoping cylinders varies between 20% to 40% of extended length. Their cost is several times that of a standard cylinder than can produce equal forces. Models are available in which all stages extend simultaneously or largest stage to smallest stage successively. Ram cylinders, Figure 11, are those cylinders with a rod OD the same diameter as the piston bore; the rod end within the cylinder is the piston. This style cylinder, used mostly for jacking purposes, is always single acting as
there is no internal cylinder volume to pressurize to retract the rod. The cylinders also are called plunger cylinders and are most often used for relatively short stroke applications. A pancake cylinder is another version of a short stroke cylinder. These often have a rod length that is a small percentage of piston diameter to provide substantial forces with little pressure. These cylinders lend themselves to pneumatically operated, automation machinery. In the past few years, rodless cylinders have become popular because their total length is approximately half that of a standard cylinder with rod extended for a given stroke. The two types of these cylinders are the cable cylinder, Figure 12, and the piston-lug version, Figure 13. Cable cylinders have a piston inside a barrel which has pulleys mounted at each end. Attached to one piston face, a cable travels around an end pulley to a metal bracket, on the other end of the bracket, a second cable travels around the second pulley and back to the other piston face. As the double-acting piston moves in one direction, the load bracket moves in the opposite direction because of the wrap around the pulleys. Options include automatic cable tensioning, single-acting models, cable tracks for greater load stabilization and capacity, a double-purchase arrangement which doubles the stroke length and stroke speed, caliper disc brakes on the cable pulley, and reed switches. The cable also can be wound around a drum to provide rotary motion. Load lug
Dirt shield
Pulley
Steel band
Barrel Piston
Fig. 12. Cable cylinder cables attach to piston faces while bracket moves load. Cylinder is one design method to shorten “extended” cylinder length.
Fig. 13. Piston-lug type of rodless cylinder is a close cousin of cable cylinder. Method used to transmit piston forces to external lug is chief difference of these cylinders. This version has built-in pneumatic brakes.
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CYLINDERS
Inlet Diaphragm
Spring Vent
Fig. 14. Single-acting flat diaphragm cylinder provides short stroke, large forces with low pressure because of large piston area.
Rolling Diaphragm
Spring
Vent
Fig. 15. Rolling diaphragm cylinder provides longer stroke than flat diaphragm version, has zero break-away friction.
Piston-lug pneumatic cylinders are of several designs, Figure 13. One type has a piston bolted to a lug; the connection between the two passes through a slit in the barrel to transfer forces to the load. Two steel bands are separated as the lug passes and act as a seal and a dust cover. Various widths of piston are available to meet any bending moments of an application. In this design, the exterior loadbearing lug moves the same direction as the piston. Stroke lengths of this cylinder can exceed 30 ft. Another variation is similar to the cable cylinder design except that the cable is replaced by a band running over pulleys at each end of the barrel and the upper lug yoke rides on the cylinder barrel for greater load stabilization and capacity. Some of these cylinders are fitted with a brake that stops the band anywhere during its stroke. This can be done with quite exacting repeatability.
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A third variation of piston-lug cylinders uses a piston in the tube with a yoke which rides on the barrel exterior. As the piston moves, the magnetically linked yoke provides linear motion. Breakaway forces of the magnetic field are in excess of 200 lb-ft. Diaphragm cylinders are either of the rolling diaphragm or the pancake type. Both use elastomeric diaphragms to seal the barrel-piston interface. The flat diaphragm type, Figure 14, uses an elastomer sheet secured at the parting halves of the pancake and are commonly used for truck and bus airbrake applications. The rolling diaphragm cylinder, Figure 15, has a hat-shaped diaphragm that rolls into the cylinder barrel as the piston advances. Both types have very low breakout forces, zero leakage, and are single-acting, spring returned. Flexible wall cylinders have evolved from designs that were originally made for vibration and isolation mounting. They consist of metal mounting plates fixed to elastomer bags that pneumatically extend and collapse as they are pressurized and vented. They have no lateral misalignment problems and are able to stroke through an arc without a clevis mount. Mounting Trouble-free use of fluid power cylinders and their ability to serve and remain leak-free depends, in large part, on properly mounting the component for the particular application. The designer must determine the loading the cylinder will experience and mount it accordingly. During the last 20 years, the NFPA has promulgated a number of standard dimensions for square-headed tie rod cylinders to promote cylinder interchangeability between manufacturers. Part of this standardization program includes cylinder mounting styles which generally provide: ● straight-line force transfer with fixed mounts that absorb force on the centerline of the cylinder ● straight-line force transfer with fixed mounts that do not absorb force on the centerline of the cylinder, and ● pivot force transfer with pivot mounts which absorb force on the centerline of the cylinder and allow the cylinder to
Fig. 16. Fixed cylinder mounts that provide straight-line force transfer are: (a) tie rods extended, and extended both ends (a + b); (c) head rectangular flange, (d) head square flange; and (e) rectangular head which provides same service as (c) but uses entire head rather than an added flange. Cap flange mounts are the same as (c) and (d) but bolt to cap (not shown). Rectangular caps also are available. See Fig. 3 for centerline lug mounts.
change alignment in one plane. Straight line, force absorbed — Cylinders with fixed mounts which absorb the force on the centerline of the cylinder are considered the best for straight line force transfer. Tie rods extended, flange, or centerline lug mounts are symmetrical and allow the thrust or tension forces of the piston rod to be distributed uniformly about the cylinder centerline, Figure 16. Mounting bolts are subjected to simple tension or shear without compound forces; when properly installed, cylinder bearing sideloading is minimized.
CYLINDERS
Cylinder tie rods are designed to withstand maximum rated internal pressure, and can be extended at either end and used to mount the cylinder. When the tie rods extend at both ends of the cylinder, one end can be used
Fig. 17. Side mounted cylinders include side lug (a), side end angle (b), side and lug (c), and side tapped (not shown). These mounts produce a turning moment as the cylinder applies force to the load.
Fig. 18. Pivot mounts absorb force along centerline and actuate loads that travel through arc. Cap trunnion (a), intermediate fixed trunnion (b) can locate anywhere between head and cap, and head trunnion (c) are versions of this style; only one of these versions is used at one time. Clevis mounting, Fig. 1, also is pivot mounting for loads that travel through arc.
for cylinder mounting and the opposite end can support the cylinder or be attached to the machine members. Flange mounts also are extremely good for straight line force transfer applications. Three styles available are head rectangular flange, head square flange, and a larger and thicker rectangular head with its own mounting holes; the same three versions are available for the cap. Selection of a flange mount depends partly on whether the major forces applied to the load result in compression or tension on the piston rod. Cap mounts are recommended for thrust loads while head mounts should be used where major loading puts the piston rod in tension. Centerline lug mounts absorb forces on the centerline; they are the least popular fixed mounting style. When used at higher pressures or under shock conditions, the lugs should be dowel pinned to the machine. Straight line, force not absorbed — Side mounted cylinders do no absorb force along their centerlines. These mounting styles have lugs on the end closures and one style has side-tapped holes for flush mounting, Figure 17. The plane of their mounting surface is not through the centerline of the cylinder; for this reason, side mounted cylinders produce a turning moment as the cylinder moves the load. This turning moment tends to rotate the cylinder about its mounting bolts. If the cylinder is not well secured to the machine, or the load is not well guided, side loads will be applied to the rod gland and piston bearings. To avoid this problem, side mounted cylinders should have a stroke length at least as long as the bore size. Shorter stroke, large bore cylinders tend to sway on their mounts with heavy loading especially with side lugs, end lugs, and end angle mounts. Side mount cylinders depend wholly on the friction of their mounting surfaces in contact with the machine to absorb the forces the cylinder produces. The torque applied to the mounting bolts should equal that of the tie rod torque as recommended by the manufacturer. For heavy loads or shock conditions, side mounted cylinders should
be held in place with a key or pins to prevent shifting. A shear key, consisting of a plate extending from the side of the cylinder can be supplied with most cylinders and should be placed at the proper end to absorb the major loading, that is at the head with the load in tension and at the cap with a thrust load. This method may be used where a keyway can be milled into a machine member. The key takes shear loads and provides accurate alignment of the cylinder. Side lug mounts are designed to allow dowel pins to pin the cylinder to the machine. When used, pins are installed on both sides of the cylinder but not at both ends. Pivot force transfer — Cylinders with pivot mounts that absorb force along the centerline should be used when the actuated load travels through an arc. There are two ways to mount a cylinder so it will pivot during the work cycle: clevis or trunnion mounts, Figure 18. Pivot mount cylinders are available with cap fixed clevis; cap detachable clevis; cap spherical bearing; and head, cap, and intermediate fixed trunnion. Special trunnion assemblies that provide gimballing action are available. Pivot mount cylinders can be used in tension or thrust applications at full rated pressure, except that long stroke cylinders in thrust applications are limited by piston rod column strength. Clevis or single-ear mounts usually are an integral part of the cylinder cap although detachable styles are available and provide a single pivot for mounting the cylinder. A pivot pin of appropriate length and diameter to withstand the maximum shear load at rated cylinder operating pressure is included as part of the clevis mount. The fixed clevis mount is the most popular and is used where the piston rod travels a fixed arc in one plane. It can be used vertically or horizontally. On long-stroke thrust applications, it may be necessary to use a larger diameter piston rod to prevent buckling or use a stop tube to minimize cylinder side loading in its extended position. Fixed clevis mounted cylinders do not function well if the path of rod travel is in more than one plane. Such an application results in misalignment
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CYLINDERS
and causes unnecessary side loading on the bearing and piston. For applications where the piston rod will travel a path not more than 3° either side of the true plane of motion, a cap spherical bearing mount should be used as well as a spherical bearing rod eye. Cap detachable clevis mounts are most often used for air or mediumduty hydraulic service. Trunnion pivot mounts also are used when the piston rod travels an arc in one plane. Trunnion pins are designed for shear loads only and should not be used with bending stresses. The support bearings should be mounted as close as possible to the trunnion shoulder faces. Head trunnion mounted cylinders usually can be specified with smaller diameter piston rods than cylinders with the pivot point at the cap or at an intermediate position. On head trunnion mounted long stroke cylinders, the designer should consider the overhanging weight at the cap end of the cylinder. To keep trunnion bearing loads within limits, stroke lengths should be not more than five times bore size. An intermediate fixed trunnion mount is the best trunnion mount. It can be located to balance the weight of the cylinder or anywhere between the head and cap to suit the application. Its location must be specified at time of order because its location cannot be easily changed once manufactured. Cylinder life — An industrial cylinder should have a design factor of about 4:1 based on yield at rated system pressure. Various mobile equipment manufacturers of heavy-duty cylinders specify a 3:1 design factor. A 15,000-psi stress at rated system pressure with smooth system operation and no pressure pulses is considered conservative. System pressure spikes that cause 30,000-psi stress often are not alarming, but at 30,000-psi unit stress, steel’s dimensional change is 0.001 in./in. of length. For a 30-in. cylinder, a pressure spike of that intensity causes a length change of almost 1/32 in. Dimensional changes in stressed cylinders, or those subjected to temperature change may further limit allowable working pressures. Large dimensional changes can seriously affect performance and life ex-
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Recommendations for leak-free systems 1.
Think zero leakage right from the start. Design for it.
2.
Design should include provisions for tube length variation and component movement due to hydraulic loads or temperature changes.
3.
Use proper clamping to dampen vibration.
4.
Use SAE straight-thread ports for sizes 1 in and under; use 4-bolt flanges for larger sizes.
5.
Specify the most cost-effective fitting that best meets your re quirements.
6.
Specify SAE standard tubes, tube fittings, hose adapters, and hose ends. Add performance requirements where missing.
7.
Insist on ANSI or ISO standard connections on all equipment regard less of the place of manufacture.
8.
Insist on quality components. Deal with reputable suppliers that can provide technical support.
9.
Train assembly personnel regularly.
10.
A well-designed system assembled with quality components by trained personnel using proper tools and procedures is least costly in the long run.
pectancy of nonmetallic cylinder seals. For example, extrusive failures of 80 Shore A durometer synthetic Nitrile seals can occur when clearance exceeds 0.004 in. at fluid pressures over 3000 psi, or a 0.001—in clearance with system pressure of 6000 psi. Such pressures can easily be reached in systems using differential cylinders or those with meter-out flow controls. The system designer must consider system shock pressures. If the hydraulic system contains speed control or energy-absorbing devices, system pressure can increase two to three times above normal system pressure. The designer must determine the loading the cylinder will experience and then mount accordingly to maintain port seal integrity. Operation conditions — The designer also must acknowledge that cylinders undergoing pressure and temperature changes elongate and contract. In addition, flexing and rocking makes the mounting head, Figure
19, sway under load. The type of mount to specify depends on the application, but the effect of pressure and temperature changes must be considered and provided for, or the cylinder will leak. Consider these factors: • cylinders with non-centerline-type mountings tend to change length and sway under load and temperature change. Any rigid tubing connected to the cylinder cap or head will be subject to the resulting force and motion. If a cylinder is rigidly plumbed, the question is not whether it will or will not leak, because it will, Figure 20. The only question is the time at which it will leak too much. Leakage at the port threads is inevitable, regardless of thread type. If cylinders are mounted this way, they will leak and probably have shorter lives than if the designer had considered the elasticity of materials and assembly ● cylinders with non-centerline mountings often require stronger ma-
CYLINDERS
Fig. 19. Cylinders with non-centerline mountings tend to sway under load.
Fig. 20. Rigid tubing connected to cylinder heads will be subject to forces and motion. Cylinders plumbed this way will leak.
Fig. 21. An overhung cylinder needs additional support to prevent cylinder movement at non-flange end. Cylinder length changes with pressure and temperature.
chine members to resist bending, so consider the rigidity of the machine frame. For example, where one end of a cylinder must be overhung, an additional supporting member should be provided, Figure 21, ● in most cases, a layout of the rodend path will determine the best type of pivot mounting ● fixed, non-centerline mounted cylinders with short strokes add another strength problem with mounting bolts subjected to increased tension that can combine with shear forces to overstress the bolts ● do the major applied forces result
in cylinder rod thrust or tension? Cap-end flange mounts are best for thrust loading; head-end flange mounts are best where the rod is in tension. Remember: pressure and temperature changes mean length changes, and ● if misalignment occurs between the cylinder and its load, the mounting style may have to be altered to accommodate the skewing movement. A simple, pivoted centerline mounting compensates for single-plane misalignment, with typical mountings using clevis and trunnion arrangements. If multiple-plane misalignment is encountered, the cylinder should have self-aligning ball joints on the cap and head ends of a clevis-mounted cylinder — and fluid-line connections should be able to accept the movement. Installation Proper installation begins with machine layout; here are some rules: ● if high shock loads are anticipated, mount the cylinder to take full advantage of its elasticity, and don’t forget: the fluid lines are along for the ride, hold fixed-mounted cylinders in place by keying or pinning at one end only ● use separate keys to take shear loads: at the head end if major shock loads are in thrust, at the cap end if they are in tension ● locating pins may be used instead of shear keys to help take shear loads and insure cylinder alignment. Avoid pinning across corners — this can cause severe warpage when a cylinder is subjected to operating temperature and pressure. Such warpage also is imposed on fluid connectors at cylinder ports, and ● pivoted mounts should have the same type of pivot as the cylinder body and the head end. Pivot axes should be parallel, never crossed. Many fluid power cylinders incorporate cushions which absorb the energy of moving masses at the end of a stroke, including the masses of the piston and rod, the load being moved, and the fluid medium operating the cylinder. When the cushion operates, the additional thrust is imposed on the cylinder assembly and it will change length. What about the fluid conductors?
For examples of cylinders used in a wide variety of actual circuits, refer to the Basic Circuits section of this handbook. Individual circuits suggest valve arrangements for specific functions. Consider protecting exposed rods from abrasion and corrosion that could destroy the rod surface and, in turn, the rod seal. Protect the rod with a suitable cover or plating process. General system design Make components accessible to ease installation and the maintenance that will be necessary later. If a fitting cannot be checked for tightness without first removing adjacent lines, for example, there is little incentive to bother with minor leaks. Many pump and motor leaks are found at the shaft seal or at the fluid ports, especially if foot-mounting arrangements are used. Don’t use foot mounts; flange or bell mounts are better. Design fluid lines so they can cope with pump and motor deflections. This practice will insure fewer leaks and fewer problems. If leakage presents a special problem, ask the manufacturer for guidance. If the working environment of the pump or actuator results in abrasion or corrosion, specify special seals and special material and finish for the shaft. The system designer should consider that all components and fluid conductors of the system are elastic: they will flex and change length because of changes in fluid pressure and temperature, changes in mechanical stress, and similar phenomena. The motion is not minor. A pressure pulse to 6000 psi will elongate a steel cylinder with a 24-in stroke by 0.024 in. If made of aluminum or cast iron, the cylinder will elongate about 2 to 2.5 times as much. If this elongation has not been accounted for in the design of the machine, the system will leak even if the latest fitting technology is used. If previous installations have continually leaked, take this as clear evidence that a new design approach would be beneficial.
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Directional control valves B
ecause of the wide variety of DCVs available in the marketplace and because of their importance in fluid power systems, an elaborate system has evolved to describe the general attributes and characteristics of DCVs. Besides the international symbology, Figure 1, included in ISO Standard
1219-1976, “Fluid power systems and components — Graphic symbols,” terms used to describe DCVs include: ● position ● ports, and ● operators. Position — The position of a DCV describes the number of operating positions
Fig. 1. DCV symbols from ISO 1219: (a) 2-port, 2-position, leveroperated; (b) 2/2, piloted and spring returned; (c) 3/2, piloted in each direction; (d) 3/2, solenoid-controlled, spring returned with significant (OFF) condition represented by dashed lines at ends of center square; (e)4/2, shifted by solenoid-controlled pilot, spring returned; (f) 5/2, piloted; (g) 3/3, pilot shifted each direction,
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Directional control valves (DCVs) are the fluid flow controllers of fluid power systems; they allow or stop pressurized fluid flow, and physically alter fluid flow direction from one path to another. When these valves permit fluid flow, they are said to be passing or open (the exact opposite of the meaning of open when applied to an electric circuit). DCVs are closed or non-passing when fluid flow is blocked or stopped.
of the valve fluid-directing mechanisms (spool, ball, poppet, slide, plug, etc.) For example, many fluid circuits require a valve in which the fluid-directing mechanism must direct flow in only two different patters. A 2-position valve meets this need as it directs the fluid appropriately depending on which position it is in.
spring returned to center. Thin lines above and below envelope squares indicate throttling or metering ability; (h) 4/3, singlestage electrohydraulic servovalve with direct operation; (j) 4/3, 2stage electrohydraulic with indirect pilot operation and mechanical feedback, and (k) 4/3, 2-stage electrohydraulic servovalve with indirect pilot operation and hydraulic feedback.
D I R E C T I O N A L C O N T R O L VA LV E S
The fluid-directing mechanism of a 3-position valve can be placed in three different positions; in one position, pressure fluid is directed to one valve port which leads to an actuator port. This pressure fluid operates the actuator in one direction — that is, a fluid motor may turn counterclockwise or a cylinder may extend. At the same time, the opposite valve port permits low-pressure or unpressurized fluid to flow from the actuator to reservoir or exhaust. When the fluid-directing mechanism of the valve is shifted to the opposite position, the valve port that received un-
pressurized fluid now transmits pressurized fluid to the actuator and moves the actuator in the opposite direction — clockwise or retract. Similarly, the originally pressurized valve port now transmits low-pressure or unpressurized fluid to reservoir or exhaust. Three-position valves also have a centered or neutral position in which a wide variety of configuration and porting arrangements are available. A few of these configurations include closedcenter, open center, tandem center, both actuator ports connected to pump, etc. These center position configurations are quite specialized depending on the use of the valve, and yet are often standard, offthe-shelf items in a valve manufacturer’s line. While the large majority of industrial fluid power DCVs are 2- and 3position, many valves used in the mobile equipment industry come in 4Fig. 2. Mobile-type DCV has intricate core patterns.
Fig. 3. Cutaway of multiple-spool stack valve shows interior coring, spooling, and supplementary valving. Spool notching is visible on each spool. As many as 10 sections of this particular valve may be placed between inlet and outlet sections.
position configurations to accommodate special needs such as power beyond and float. Ports — While most U.S. DCV literature refers to the way of a valve, such as 2-way, 3-way, or 4-way, the international standard uses the word ports. Thus, what is known as 2-way, 2-position DCV in this country is called a 2port, 2-position valve internationally and can be abbreviated 2/2. The number before the slash identified the number of ports while the second number refers to the number of positions. Other common mobile and pneumatic DCV port configurations include 3-, 4-, 5-, and 6-port; 3-ported valves are mostly 2-position, 3/2, while 4-, 5-, and 6-ported valves are nearly always 3- or 4-position, 4/3, 5/3, 6/3, and 4/4, 5/4, or 6/4. Operators — All DCVs have a device that changes or shifts the fluid-directing mechanism of the valve from one operating configuration to another. These devices are called valve operator(s), and include various collections of devices — levers, pedals, cams, followers, springs, pilots, ballscrew arrangements, solenoids, proportional solenoids, torque motors, and many more. Many valves have combinations of these operators so that the valve can be shifted in response to more than one type of signal. For example, a 4/3 valve can solenoid shift in one direction with a spring return to neutral when the electric signal ends. The shift in the opposite direction can be powered by a cam, again with a spring return when the cam no longer physically forces the spool in the shifted position. Presently, a great deal of work and development is underway to make industrial fluid power DCVs responsive to solid state and digital control signals. Basic DCV designs The fluid directing mechanism of DCVs often performs its function using a sliding or rising motion within the valve envelope. Sliding-action valves have long been the more popular type of DCVs used in fluid power systems, but recent advances in manifolding and leakage control have increased use of rising-action or poppet valves as well as newer designs and
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D I R E C T I O N A L C O N T R O L VA LV E S
Fig. 5. Lapped-spool and floating sleeve 5/2 pneumatic valve is directly actuated by the single, right-hand solenoid, and is spring-returned. Base has automatic electrical plug-in capability, pilot light to indicate solenoid operation, and manual operator.
Fig. 4. Cutaway view of 4-port hydraulic DCV. Solenoid connector housing sits atop the valve.
other special valving mechanisms. The most common sliding-action valve is the spool-type valve, Figure 2. Fluid flow is controlled at the spool slides between annular ports to open and close flow paths depending on spool position. Spool valves readily adapt to various spool-actuating schemes which broaden their use over a large variety of application areas. As the spool slides under control of a spool actuator, the widths of the spool lands relative to annular ports have a great deal to do with the way the valve reacts to its load. If the land is wider than the port opening, the spool is said to be overlapped, in that it must move a greater distance from neutral before the land begins to uncover the port so fluid can start to flow. When the land is shorter than the port opening, the spool is said to be underlapped and flow is never completely stopped because the land never blocks a port completely. Additionally, many mobile applications require metering or throttling in which the operator closes a manual control loop to slow or gently break away a load. In those instance, the spool may be modified with V notches, for example, Figure 3, so that a small displacement of the spool gradually permits increasing or decreasing fluid flow to gradually speed or slow actuator and load movement. A variation of the single spool valve or multiple spool case is the stack valve, Figure 3, in which a number of spool and envelope sections are bolted together between an inlet and outlet section to provide multiple DCV capability. In addition to providing a central valve location for the machine op-
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erator, the valve grouping reduces the number of fluid connections involved and increases ease of sealing. The number of valves that can be stacked in this manner vary from manufacturer to manufacturer. Many electrically actuated hydraulic DCVs now use the wet armature design. In this arrangement, the solenoid armature is contained in a pressure vessel cavity exposed to hydraulic fluid at return line pressure. The solenoid coil mounts outside the cavity, Figure 4. There are no working dynamic seals between the cavity and the valve’s exterior, so leakage is minimized. In addition, the fluid bath
Fig. 6. This cylindrical DCV, resembling an old radio tube, plugs into a keyed subplate which has eight ports tapped for fittings. Various arrangements of pilots, outputs, and exhaust passages extend longitudinally through valve body that surrounds spool cavity. Of brass, stainless steel, and acetal copolymer, this valve has a much larger, all metal cousin for hydraulic service.
cushions the solenoid and reduces noise, and the moving fluid dissipates armature heat. A pneumatic valve, Figure 5, has a lapped spool which floats on a thin film of air inside a sleeve that has been matched to that particular spool. These heavy-duty valves have been available for many years and have improved in quality with advances in manufactur-
Fig. 7. Five-port, 2-position pneumatic poppet valve has solenoid-controlled internal pneumatic pilot with spring return. Valve may be manifolded with other company valves.
D I R E C T I O N A L C O N T R O L VA LV E S
Fig. 8. Directional poppet valve with polyacetal body has built-in speed controls, manual override, and integral fittings for polyethylene or Nylon tubing. Valve is held to manifold using spring-latch assembly for easy removal, replacement. Fig. 10. Detented rotary plate valve has relatively unobstructed
longitudinal pas- flow passages. Sealing rings have lower seats on wave springs sages to the work- to continually force seals against rotary plate. Valve operating, ing ports. This eased by roller thrust bearings, continues to lap plate, seals. valving system permits close mounting density and valve and a metal envelope and poppet. The valve shown in Figure 8 is made of virremoval without disturbing tubing. Another pneumatic valve, Figure 7, tually all engineered plastic. The weight is of the poppet variety. This 5/2 valve of this valve is considerably less than a is solenoid-controlled and pilot-oper- comparable all-metal valve which ated with manual-override. The materi- makes the plastic valve attractive in apals of this valve include engineered plications where weight is a problem; plastics for some of the internal parts for example, a mount on some cantilevered machine member such as a robot arm. Additionally, plastic valves are less susceptible to damage from unfiltered, high-moisture-content air. A variation of the poppet valve is shown in Figure 9. This valve has a pneumatically-piloted spindle with two floating exhaust poppets. A center, double-faced poppet integral to the spindle routes pressure fluid to the appropriate working port and opens the other work port to exhaust. Rotary-plate valves, Figure 10, are capable of high-pressure operation and can provide superior sealing capabilities because the sealing rings and roFig. 9. Pneumatic 5/2 DCV has floating poppets that open, close as central spindle is pi- tary plate continually lap one another loted to open working ports. In DCV, detents hold spindle in last position when pneumatic as the valve is operated. With relapower is removed. This section, which mounts on subbase, is available with electrical tively unobstructed flow passages, connectors that plug in automatically as valve is placed on mount. these valves have a long service life. ing methods. Clearances between the spool and sleeve are in the millionths of an inch range. O-ring seals around the sleeve are static and not subject to heavy wear cycles. A different type of pneumatic spool valve, made almost entirely of plastic, Figure 6, is configured to mount on an 8port subplate. Supply, exhaust, and working air are routed longitudinally through passages into the valve spool cavity, and then back out through other
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D I R E C T I O N A L C O N T R O L VA LV E S
Hand lever may incorporate auxiliary switching
Elastomeric boot seals out contamination Spring-loaded returnto-center mechanism
Electronic circuitry provides output for controller/valve
Fig. 11. Joystick sends electric signal to hydraulic valves, allowing operator to easily control machinery.
Joysticks, Figure 11, are increasingly popular controls in mobile equipment applications. This valve contains four pilot valves that route pressure fluid to remotely-controlled
Fig. 12. Double-solenoid DCV mounts on rectangular subbase. Cutaway section reveals packed spool construction.
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two 3-position DCVs. The handle may have an electric thumb switch to introduce a second control option. Use of this type control has made multiple hydraulic valve control and coordination much easier for machine operators. DCV mounting Some valves are built in a configuration for inline mounting but, because this arrangement requires extra unions in the piping system for to remove the valve for maintenance or replacement, most designers choose subbase or manifold mounting. In the subbase mounting system, all conducting lines are connected to the subbase and the valve is gasketed onto the base face with matching port pattern, Figure 12. The static subbase does not wear out, so it does not have to be removed; the DCV can be lifted off the subbase
after loosening a couple of bolts without disturbing the plumbing. The manifold system expands the subbase idea by mounting several valves on a common base, Figure 13, with an integral master supply conductor and interconnected tank or exhaust channels. Plumbing is reduced further with the same convenience of simple valve removal. A manifold can be a single drilled plate, but stacking manifolds are becoming increasingly popular. Individual valve manifold stations are stacked to build a large manifold. For versatility, stations can be added or removed as necessary, and parts stocking is minimized. Another way to combine DCVs (and other valves) with manifolds is to use cartridge-type valves, Figure 14. These components may screw into a cavity drilled in the manifold or they may slide into a cavity with a cap or another component to hold the cartridge in place. Advantages are simplified plumbing, reduced leakage potential, and cleaner appearance. Actuators for DCVs Valve actuators or operators are the part or parts through which force is applied to move or position flow-directing elements. The sequence, timing, and frequency of actuator operating is a key factor in fluid power system performance. The majority of directional control valves already installed on equipment and in the marketplace today are digital. Their internal flow-directing elements move quickly to fully open or fully close a flow path when the valve shifts. The parts of these digital valves are designed so that there is essentially no cross-port flow during shifting. The high shifting speed and the fact that the valves are fully opened or fully closed has led to the descriptive term, bang-bang, that frequently is applied to these valves. As long as the actuator force can perform its function, the system designer can select any appropriate DCV actuator for the conditions and type of control under which the system will operate. Actuators for conventional DCVs can be divided into three general types: ● manual or mechanical ● pilot, and ● electrical and electronic.
D I R E C T I O N A L C O N T R O L VA LV E S
spring-offset. In three-position valves, two springs hold the non-actuated valve in its center position until an actuating force shifts it. When the actuating force is removed, the springs re-center the valve, leading to the common identification: springcentered valve. Detents are locks that hold a valve in its last position after the actuating force is removed, until another Fig. 13. Manifold-mounted valves provide benefits such as com- stronger force shifts pact design, centralized functions, light weight, less potential the valve to another leakage, and neater, simpler, more economical plumbing. HICs position. The detents need not incorporate only cartridge valves; standard subplate- then hold that new pomounted valves are also widely used for greatest versatility. sition after the actuating signal is removed. See ISO Standard 1219 (excerpts of Mechanical actuation is probably which appear elsewhere in this hand- the most positive way to control indusbook) for the standard actuator sym- trial fluid power equipment. If an acbols that operate conventional DCVs. tion is to occur only when a machine Note that different actuators usually do element is in a certain position, the not require different designs of the ba- equipment can be designed so that the sic valve. The DCV remains the same, machine element physically shifts the with provision for mounting a variety valve through a mechanical actuator of actuators on the body. when the element reaches the correct With manual or mechanical actua- position. This arrangement practically tors, the operator or some machine ele- eliminates the possibility of false or ment applies force on the DCV’s flow- phantom signals actuating DCVs at the directing element to move or shift it to wrong time. another position. Manual actuators inHowever, mounting mechanically clude levers, palm buttons, pushbut- actuated DCVs on a machine requires tons, and treadles. Mechanical actua- some special cautions. The valve and tors might be cams, rollers, levers, actuator may be exposed to a wet springs, stems, or screws. and/or dirty environment that requires Note that different actuators do not special sealing. The actuator will probrequire different designs of the basic ably be subjected to impact loads, valve. The DCV remains the same, which must be limited to avoid physiwith provision for mounting a variety cal damage. Valve alignment with the of actuators on the body. actuating element is important so the Springs and detents are mechanical valve must be mounted accurately and elements that are part of many DCV securely for long service life. actuation arrangements. In 2-position Pilot-actuated valves are shifted by valves, springs hold the non-actuated air or hydraulic pressure fluid applying valve in one position until and actuat- forces on the DCV’s flow-directing eleing force great enough to compress ments. An important advantage of pilotthe spring shifts the valve. When the actuation is that large shifting forces actuating force is removed, the spring can be developed without the impact returns the valve to its original posi- and wear that affects mechanically action. This type valve often is called tuated DCVs. Pilot-actuated valves can
be mounted in any convenient or remote location to which pressure fluid can be piped. The absence of sparks or heat makes pilot-actuated valves attractive for applications in flammable or explosive environments. Electrical or electronic DCV actuation involves firing or energizing a solenoid. The force generated at the solenoid plunger then shifts the valve’s flow-directing element. Solenoid-actuated valves are particularly popular for industrial machines because of the ready availability of electric power in industrial plants, but most mobile equipment has electric power also. The selection of AC or DC solenoids depends on the form avail-
Fig. 14. Sealed-body cartridge valves slip into manifold cavities.
able. At one time DC solenoids offered longer service life, but improvements in AC solenoid designs has eliminated that advantage. There is a practical limit to the force that solenoids can generate. This means they cannot directly shift valves that require large shifting forces. The solution to this situation is to use small solenoid pilot valves to control the flow of pressure fluid to pilots that actuate the main valve. Two more-sophisticated methods of DCV control are radio signals and fiber optics. Radio signals can reach remote locations, and fiber optics often are used when electrical insulation is desirable for safety reasons. Some valves use more than one type of actuator signal in combination or as alternatives. The relationship between the signals usually is distinguished by the words and or or in the DCV description: solenoid and pilot actuated; solenoid or pilot actuated.
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Electrohydraulic proportional and servovalves
As introduction to proportional electrohydraulic valves, consider the symbol, Figure 1. Observe that horizontal lines just outside the schematic of the valve envelope and parallel to its longitudinal axis indicate that the valve is continuously variable. That is, the spool can take and maintain a position anywhere within its limits of travel, precisely the valve of interest — electrohydraulic valves. The two types in use today are servo and proportional valves.
Ps
Torque
R1
R2 Pivot
Flapper Nozzle A
B
ridges are used extensively in hydraulic components and circuits. All 4-way directional control valves can be drawn as a bridge circuit, a useful fact if one intends to mathematically model the valve for simulation and other analytical studies. Viewing the principal parts of a flapper-nozzle servovalve, Figure 2(a), understand that a torque applied from a torque motor to the
Ps
Fig. 1. Valve-envelope schematic with outer parallel lines indicates spool’s ability to assume any intermediate position over full range of spool travel.
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flapper arm, say in the clockwise direction, moves the flapper closer to nozzle A and tends to close it. Concurrently, the flapper moves away from nozzle B to allow more flow therefrom, so the net result is a rise in pressure Pa and a drop in pressure P b. The pressure difference, P a — Pb is felt across the two ends of the main valve spool, driving it to the right and creating communication from P port to B port and from A port to T port. Drawn schematically, Figure 2(b), a hydraulic bridge circuit clearly is formed. A 4-way directional control valve also is a bridge circuit, Figure 3(a). When the valve spool moves to the right, Rp to a and Rb to t open while Rp to b and Ra to t close. Flow issues from the valve’s A port to the load and returns via B port to tank. Left spool movement opens R p to b and R a to t so that flow issues from the valve’s B port to the load, returning to tank via A port. The schematic, Fig 3(b), clearly
Ps
Nozzle B
Ra Rb T
Ps
A B Main spool (a) Ps
PA
R1
R2
Ra
PB
Rb
(b) Fig. 2. When the flapper nozzle pilot section (a) is drawn in schematic form, (b), it is obvious that a bridge circuit exists. By moving the flapper, restrictions Ra and Rb change in opposite directions. This unbalances the bridge and causes the spool to move against its centering springs.
ELECTROHYDRAULIC INTERFACING
Ra to t R p to a R p to b Rb to t Ps
PEHID
PEMID
PEHID A
Pilot operated
B
A
(a)
Rp to a
R p to b Load
B port
R b to t
Ra to t
B
Move the load
Ps
A port
Direct acting
(b)
Fig. 3. The 4-way spool valve has four individual lands that vary in unison as the spool shifts — two lands open while the other two close. When drawn in schematic form, it is clear that the four lands constitute a bridge circuit, and spool movement unbalances the bridge one way or the other to cause a reversal in load flow.
shows the bridge circuit. It is balanced to stop the load and unbalanced to move it. Electromechanical actuators It is possible to construct proportional electrohydraulic interface devices (PEHIDs) only because of the invention of certain proportional electromechanical interface devices (PEMIDs). The PEMIDs commonly used in the fluid power industry include: ● torque motors ● linear force motors, and ● proportional solenoids. The PEMIDs receive an electrical current input, convert it to mechanical force and motion, and then transform the energy into some sort of hydromechanical action. The direct mechanical action is always within a valve, although that valve may stroke a pump or directly power a load. A circuit designer would select one path or the other of the family tree, Figure 4, for a given application. Valve control is called the energy-loss control method
Stroke a variabledisplacement pump
Move the load
Fig. 4. Proportional electromechanical interface device (PEMID) causes an electrohydraulic valve (PEHID) to shift with two results: valve-control method (A), where valve output moves the load, and (B), where the valve changes the displacement of a pump with resulting pump output then moving the load, an example of the pump control method.
in path A, because the valve, being a restrictive device, consumes excess power as a necessary part of its control function. Path B, on the other hand, is called the volume-control or load-demand method that supplies only as much power as the load needs, wants, or can use. The only losses encountered using this method are those caused by the modest inefficiencies of the pump and actuator, and in total are nearly always less than those for the energyloss method, all other things being equal. This leads to these truths reDC command voltage
+
Error voltage Electric comparator
Fig. 5. Family tree of PEHIDs has two branches: direct-acting branch where the PEMID directly moves the main spool, and pilot- operated branch, where the PEMID shifts a pilot stage whose output moves the main spool.
garding proportioning hydraulic systems: ● path B is always more efficient than path A, and ● path A always has less initial cost because valves are less expensive than variable-displacement pumps, and one fixed - displacement pump can supply pressure fluid to more than one valve and functioning circuit branch. Electrohydraulic valves Continuously variable electrohydraulic valves illustrate the fact that a continuously varying control current always results in a continuously varying, controlled-output variable. That output variable could be flow, pressure, or simply the position of a spool that affects final flow and/or pressure. Broad categories of these continuously variable valves are: ● direct-acting valves where the force
I Power amplifier pre-amp
Main spool force Main spool Proportional solenoid
DC feedback voltage
Transducer signal conditioner
Spool position transducer
Spool position
Fig. 6. A small current must cause a small spool shift and a large current must cause a large spool shift in continuously variable electrohydraulic valves. To assure such proportional, stepless spool positioning, some valves use a spool-position transducer to measure actual spool position. The spool is made to stop in a position commensurate with the command voltage through feedback- loop closure. This closed feedback loop often is called the inner loop.
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ELECTROHYDRAULIC INTERFACING
potential (pressure) en- tion not totally controlled by the conergy is converted into trol current (sole-noid force). This kinetic (velocity) en- does not mean the valve does not N N ergy in the constricting function; it does mean that the region of the valve. In spool’s exact position at any moment spool valves, the flow is harder to predict. It is true, howforce always acts to ever, that in some direct-acting close the valve regard- valves, flow forces can be so high less of the direction of that they cause an automatic nearS S flow. The consequence closure at high pressures and flows. Flapper is that when using the Load-dependent spool position can Air gaps 4 places valve, say at a low be detected by mathematical analysis Lower frame pressure drop (flow), of valve pressure drop at a steady Nozzles the spool is in a posi- control current but varying flows. Fig. 7. The four nominally equal air gaps of an electromag- tion where the solenoid The curve, if there is no load-induced netic torque motor each carry equal magnetic flux from the force is balanced by the spool shift, will relate pressure drop permanent magnets producing zero net torque on the arma- restoring spring. As the to the square of the passing flow. If ture. When current enters the coil, coil-induced magnetic flux valve’s pressure drop the data does not fit the square relaadds to or subtracts from the four air-gap fluxes creating a increases, either be- tionship, flow forces are probably torque on the armature. Armature movement typically causes cause of a reduction in causing a spool shift. a flapper to move, changing resistivity of the two nozzles. load restriction or an Stiction forces also act upon the increase in supply pres- spool and the solenoid’s armature, sure, the flow force in- s o t h a t s p o o l p o s i t i o n d o e s n o t of the proportional solenoid acts di- creases so as to close the valve. smoothly vary as control current rectly upon the main spool to provide As a result, the spool takes a posi- continuously and smoothly changes. the desired degree of hydraulic control, and PEMID ● pilot-operated valves in which the PEMID acts first upon a primary hyA B dromechanical device whose output acts on a main spool. These are also Flapper nozzle Steam deflection sometimes called multi-stage valves C D (momentum change) Armature connected and are either 2-or 3-stage but never to flapper sitting between two opposed nozzles more. Each of the two paths of the Jet pipe with motor rotation simple family tree of all continuously changing flapper-to-nozzle Swinging wand Armature connected to steerable, distances variable electrohydraulic valves, Figflexibly mounted jet pipe that issues high-velocity fluid that ure 5, can be further subdivided, but is then collected in either of two that is not elaborated here. receiving ports. Velocity converted Permanent magnet
Armature
Pivot center
Coil
Direct-acting valves Further subdivisions include those valves that use some means of spoolposition feedback and those that do not. The non-feedback types simply take the force of the proportional solenoid and put it against a restoring spring. Thus, the main spool would take a position commensurate with the force generated by the solenoid if those were the only two forces acting on the main spool. Unfortunately, there are two other significant forces that act on the spool: flow forces and stiction forces. Flow forces are a natural phenomenon in all control valves that result from the momentum change that takes place as the result of the valve’s throttling effect. This occurs when
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Upper frame
to pressure through momentum change
Stiff system
Soft system
Dual nozzle
Single nozzle
Has stiff flexure mount where fluid momentum forces of nozzle effluent are small compared to magnetic and spring forces
Has weak flexure spring and magnetic force. Nozzle effluent forces are significant when compared to magnetic forces
Produces two flow streams that deflect off edges of wand
Produces single flow stream deflected by single hole in wand
Deflected stream(s) collected in either of two receiving ports, converted to pressure through momentum change
Hydromechanical unbalances result in hydrostatic forces that act on ends of main valve spool, motivating it to shift. Final position of main spool caused by feedback or restoring spring
Fig. 8. A family tree of torque-motor electromechanical interfaces indicates all of the common piloting methods in present industrial use.
ELECTROHYDRAULIC INTERFACING
Instead, the flow (spool position) has a staircase effect. Additionally, the curve trace for increasing control current is not the same as the curve trace for decreasing control current, producing stiction-induced hysteresis. If the valve is to be used in manually operated control systems, this hysteresis is not a major problem because a human operator can compensate easily for such performance aberrations. But when automatic controls using feedback are contemplated, hysteresis can cause a continual hunting or oscillation rather than smooth and stable operation. Incorporating electronic dithering — that is causing the spool to be in a continuous but acceptably small state of agitation — can help significantly to reduce the detrimental effects of stiction-induced hysteresis. When properly implemented, a closed-loop feedback control system around the spool, called the inner loop in the hydraulics industry, Figure 6, can all but eliminate stiction and flow-force effects. This loop is closed by measuring actual spool position, usually with an LVDT position transducer, and comparing it to the commanded position. If the position is incorrect, the electric current to the valve’s proportional solenoid is adjusted until spool position becomes correct. Thus, the spool is always in the exact spot commanded, dynamically induced lags notwithstanding. While not exactly true, the foregoing statement is acceptable for all practical purposes. The spool position feedback transducer of choice is nearly always an LVDT. Because LVDTs must be operated with AC voltage, they must always be accompanied by a special electronic signal conditioner that has: ● an oscillator section that generates AC voltage to excite the transformer. This AC voltage is not derived from the 60 Hz power- company line; it is generated by a solid-state electronic oscillator usually outputting a few volts at a fixed frequency, generally between 3 kHz and 10 kHz, and ● a phase-sensitive demodulator section that converts a transduced AC signal into an equivalent DC sig-
,,,,,,,, ,,,,,,,, N,,,,,,,, ,,,,,,,,
,,,,,,,, ,,,,,,,, ,,,,,,,,N ,,,,,,,,
S,,,,,,,, ,,,,,,,, ,,,,,,,, ,,,,,,,,
,,,,,,,,S ,,,,,,,, ,,,,,,,, ,,,,,,,,
Valve responding to change in electrical input
S
N
Ps
Ps Ps
R
PB
R
Feedback spring P s
PA
V High
Low (a) C1
C2
,,,,,,,, ,,,,,,,, N ,,,,,,,, ,,,,,,,,
,,,,,,,, ,,,,,,,, ,,,,,,,,N ,,,,,,,,
,,,,,,,,
,,,,,,,, ,,,,,,,,S ,,,,,,,, ,,,,,,,,
N
S ,,,,,,,, ,,,,,,,,
,,,,,,,,
Valve condition following change
S
Ps
Ps Ps
PB
R
Ps
R
PA
V
(b) C1
C2
Fig. 9. Current entering the torque-motor coil, (a), causes the armature to rotate against a stiff feedback spring. The flapper, attached to the armature, blocks nozzle A and relieves nozzle B, causing pressure PA to rise and PB to fall. This unbalance moves the spool to the left. As the spool moves, (b), the feedback spring, anchored to the spool and the flapper, forces the flapper toward center. Eventually, the flapper and spool reach a position where the flapper is nearly centered, the pressures are nearly equal, and the spool comes to rest at a position commensurate with the amount of torque (coil current).
nal with the full sense of the algebraic sign of the measured position. Pilot-operated valves Valves with larger flow capacities need pilot stages to boost the power necessary to shift the larger spools. The electromechanical methods used for this staging are torque motors, force motors, and proportional solenoids. Torque motors, Figure 7, are electromechanical rotary machines whose
rotational travel is restricted, often less than 1 or 2 and are nearly always used for a piloting function. They are fitted with permanent magnets as the major flux source with the flux paths arranged to form a force bridge. Their limited rotation allows the armature to be mounted on a stiff flexure spring rather than bearings, although there is one known proprietary exception that uses a soft spring. The stiff spring and lack of bearings virtually eliminate hysteresis caused by bearing restriction.
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ELECTROHYDRAULIC INTERFACING
Armature pivot
Armature First-stage progector
Receivers Filter screen Feedback spring
C1 P
C2 R
P
Fig. 10. Current in the torque motor of a jet-pipe servovalve steers a jet nozzle, causing a pressure difference between two collector ports. If A -port pressure is high, for example, the main spool moves to the right. Concurrently, the feedback spring drags the jet nozzle toward center and approximately equalizes collector pressures. Thus, the main spool has been positioned as directed by the coil current.
Second-stage spool
Supply pressure
Inlet orifices
Return port
Blade
CP 1, 1 Diverter plate
CP 2,2 Receiving (a) P = P orifices 1 2
C1,P1
CP 2,2 (b) P > P 1 2
Incoming current creates a second set of magnetic fluxes that unbalance the force bridge and results in net torque. The torque causes angular rotation until the flux-induced torque equals the counter-torque of the flexing spring plus any external load. An important characteristic of the torque motor is that the direction of rotation is affected by the direction of current through the coil. The electromagnetic field caused by the current is compared to the field of the permanent magnet in the magnetic bridge circuit and rotation ensues in a commensurate direction. In the final valve assembly, the torque-motor armature is connected to a flapper sitting between two opposed nozzles, a jet pipe, or a swinging wand or blade. These last two steer a fluid stream, Figure 8, branch B. Basic operating principles and conceptual construction of flapper-nozzle and jet pipe servovalves are indicated in Figures 9 and 10, respectively. Torque motors almost exclusively pilot servovalves, and usually requires less than one watt of power to fully operate although that is not a hardand-fast rule. Torque from the torque motor of a jet pipe servovalve steers the jet to one receiver or the other, unbalancing spool end pressures. Movement of the main spool continues until the feedback spring between the main spool and jet forces the jet pipe back to near null. Main spool position then is commensurate with coil current.
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The flapper-nozzle has two different implementations: the one already mentioned is the stiff design, wherein the force due to the impinging nozzle flow is small in comparison to spring and torque-motor forces. In soft designs, the torque motor and nozzles are deliberately sized so that nozzle effluent causes a significant force on the flapper. One argument concludes that this design is more tolerant of certain contamination problems. The argument goes like this: when the two fixed orifices are open fully and unclogged, the unpowered flapper will center due to a combination of fluidmomentum force acting on the flapper, restoring force on the light spring, and magnetic force in the torque motor. Should one of the nozzles or fixed orifices become partially blocked, reduced effluent flow produces less force on the blocked side of the flapper. The flapper then moves toward the clogged bridge leg until reduced force from the opposite receding nozzle equals the diminished flow force from the partially blocked nozzle. Current input to the torque motor then causes the flapper to move about a shifted neutral, but the pressure does not go to a hard-over level. The main spool might not shift fully in one direction, however. The swinging wand, Figure 8 path D, has a mechanical-to-hydraulic interface that is proprietary. The versions of this interface include: ● dual nozzle where the two fluid
C1,P1
CP 2,2 (c) P < P 1 2
Fig. 11. The swinging-wand pilot stage generates a differential pressure in receiver ports C 1 and C 2 by deflecting two fluid streams off each edge of the wand. An unseen torque motor moves the wand in proportion to the amount of current. Thus the pressure difference between C1 and C2 is a reflection of coil current. Port pressure are equal, (a), C1 pressure is higher, (b), and lower, (c).
streams are deflected off the outside edges of the wand, and ● single nozzle, where a single fluid stream passes through a central hole in the wand. Consider the dual-nozzle version, Figure 11. The two fluid streams issuing from the source side of the pilot head are collected in opposing receiving ports. When a current into the torque motor causes the wand to swing, one receiving port experiences a rise in pressure while the other experiences a pressure reduction. As in the case of the jet pipe and flapper-nozzle pilots, the resulting difference in pressure shifts the valve’s main spool. The single-nozzle version has a hole laterally bored and centrally located in the wand such that the single
ELECTROHYDRAULIC INTERFACING
Electromechanical section Air gap B
Permanet magnet
,, ,
Coarse filter
Hydraulic section Supply pressure Inlet orifice
Air gap A
Controlled pressure output
,,
I
Electrical interface
Nozzle
Poppet
Air gap C
Air gap D
Armature
Coil
Fig. 12. A permanent magnet creates equal fluxes in the four air gaps of electromagnetic force motor that results in net zero force on the armature. Current into the coil in the direction shown, for example, strengthens flux in gaps B and D and weakens flux in gaps A and C. Now there is a net force to the left, pushing the poppet against the nozzle. Through control of force, the current controls output pressure.
ent pressures being Trapezoidal air gap Armature End cap Non-magnetic collected in the two reforce pin ceivers. The resulting Override push pin differential pressure between the two receiving ports shifts the main spool. This swinging-wand design has a supply pressure limitation in that the pilot head must be sized for a particular supply pressure range. Ferromagnetic Inner Coil Outer tube Non-magnetic If flow issuing from the nose flux tube end piece nozzle(s) is excessive, fluid momentum force acting on the wand can Fig. 13. The trapezoidal air gap of a proportional solenoid pin it against the reis shaped to create a relatively constant force regardless of ceiver side to lock the armature position when the current is constant. Because wand. Installing an orithere are no permanent magnets, the force is always in fice, matched with the }one direction (to the left here), regardless of current direc- supply pressure and the tion. Thus, bidirectional valves always require two pro- needs of the pilot stage in series with the nozzle portional solenoids. side, remedies this fluid stream issuing from the single problem. Approximately a 2:1 change nozzle must pass through the hole. in supply pressure is possible with a When the wand is centered, equal single orifice. pressures are collected in the two reForce motors are the linear equivceiving ports. A current into the coil alent of torque motors in that they causes the wand to shift and the fluid also have permanent magnets inside. stream is deflected off the inside edge Therefore the direction of motion deof the central hole resulting in differ- pends upon the direction of input cur-
rent, Figure 12. There is only one manufacturer of force motors in the U. S.: FEMA Corp., Portage, Mich. The two permanent magnets each create attractive forces, each urging the armature toward it, but nominally offsetting one another when centered. Additionally, a stiff centering spring prevents either of the natural regenerative attractive forces from pulling the armature either way. When a current is applied in the direction shown in Figure 12, the resulting electromagnetic fields act to strengthen the magnetic fields by increased flux density in air-gaps B and D while at the same time weakening the fields in air-gaps A and C. The resulting force moves the armature and poppet to the left. State-of-the-art force-motor design produces a maximum of about five pounds of stall force, about 0.02-in. of travel (no load) at about 5 watts of power. Proportional solenoids Proportional solenoids are about 25 to 30 years old and are manufactured by a number of companies worldwide. Some market their solenoids to U. S. industry, while others supply themselves. All competing products have similar performance specifications. State-of-the-art proportional solenoid design yield these approximate typical specifications: ● maximum force, 20 lb ● proportioning travel, 0.10 in. current, 12-V coil, 1.5 to 2.5 A, and ● power, 15 to 25 watts. Figure 13 indicates approximate construction detail of a proportional solenoid. The secrets to success lie in forming the proper trapezoidal air gap dimensions and in keeping stiction low. Figure 14 shows a representative force-displacement curve for a proportional solenoid. Its unique characteristics are a region of relatively constant force as the armature changes position, plus relatively linear changes in force for changes in solenoid current, both performance goals sought by the solenoid’s designers. It is probably true that neither the constancy of force nor the linearity with current are as important as the manufacturers would claim. Things incorporated external to the solenoid
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ELECTROHYDRAULIC INTERFACING
Device comparisons Contrasts must be made between proportional solenoids and force/torque motors, as some differences are apparent in the specifications while others are not. Torque/force motors require lower current levels. Proportional solenoids: ● require much higher electrical output power than their motor counterparts ● produce substantially greater mechanical travel than motors produce higher force levels ● produce higher stiction levels ● operate with greater hysteresis, and ● generate forces is a direction independent of current direction. Therefore, to make a 4-way directional valve operate requires two proportional solenoids but only one force/torque motor. All of these factors make proportional-solenoid drive electronics more complex than that necessary for force/torque motors. Proportionalsolenoid power requirements have caused manufacturers of proportionalvalve drive electronics to adopt pulse width modulation (PWM) as the
Proportioning region - 0.10 in.
Force
lmax l3 l2 l1 Armature position-enlarging air gap
Fig. 14. Typical force vs. armature position curves show region of proportional solenoid armature travel where there is relatively constant force at constant current. Valve designers must use the solenoid so the armature operates in this proportional region. With current technology, the region is about 0.10-in wide.
substantially affect performance of the total valve: use of pressure feedback or armature-position feedback, for example. Furthermore, the enormous versatility of the hardware and software of modern computers makes control and linearization a fairly straightforward task.
Pilot-operated PEHIDs A
B
Momentum change (kinetic)
Sliding spool (hydrostatic) W
D
C
Linear electromechanical transducer
Rotary mechanical transducer E
Y
F
Flapper nozzle Stream steering
Force motor (single coil) AA
Z
Current-to-pressure transduction G
X
Solenoid (dual coil)
Variable orifice
Current-to-position transduction
H Stream steering
Flapper nozzel
CC
BB Main spool feedback (electrical or mechanical)
K
J
I
N
M Feedback
O
Open loop Feedback
U Single spool
P Open loop
Q Feedback
V Double spool
Fig. 15. Family tree of pilot-operated PHEIDs.
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Swinging wand
Main spool
Main spool
Main spool
Main spool
L
Jet pipe
Stiff
Soft
Main spool open loop
R Open loop
S Feedback
T Open loop
power-output method of choice. The major reason for the use of PWM is to handle the high-output power required of the solenoid without overburdening the power- output transistors. There is a second benefit of using PWM: if the PWM frequency is sufficiently low, it automatically provides a mechanical dither that helps minimize stiction-induced hysteresis. In some valves the effects of dither can only be described as dramatic when looking at the reduction in hysteresis. The correct dither frequency must be determined after the valve is designed. Furthermore, the frequency selected must be a compromise between propagating the dither pulsations imperceptibly into the hydraulic circuit and yet achieving sufficient reduction in stiction. A low frequency helps the stiction problem but if too low, the user of the hydraulic system can feel the pulsations. U. S. industry uses PWM frequencies from about 33 Hz to about 400 Hz. At least one European manufacturer uses 40 kHz and receives no dither effects whatsoever. Their amplifier supplies dither with a separate on-board dither generator. There is an advantage to this method: dither power remains constant throughout the modulation range, whereas when relying on the PWM frequency for dithering, the dither power varies with the amount of modulation. There is none at the 0% and 100% modulation points, but maximum at 50% modulation. Summary of pilot-operated valves Figure 15 shows a family tree of all electrically modulated, continuously variable pilot-operated valves. It is a peculiarity of the U. S. hydraulic valve-manufacturing industry that each terminus of the tree also tends to define a specific manufacturer’s product. For example, the A/C/E/I/N/V path rather accurately describes the servovalve of Sauer-Sundstrand’s Controls Div. in Minneapolis; its mate, U, has no supplier. In contrast, though, the A/C/E/J/O path is well populated. That is where the products of Moog, Vickers, Rexroth, Dynamic Valve, and others have congregated.
Among the basic elements of virtually every hydraulic system is a series of fittings for connecting tube, pipe, and hose to pumps, valves, actuators, and other components.
Fittings I
f components within the hydraulic system never had to be removed, connections could be brazed or welded to maximize reliability. However, it is inevitable that connections must be broken to allow servicing or replacing components, so removable fittings are a necessity for all but the most specialized hydraulic systems. To this end, fitting designs have advanced considerably over the years to improve performance and installation convenience, but the overall function of these components remains relatively unchanged. Fittings seal fluid within the hydraulic system by one of two techniques: all-metal fittings rely on metal-to-metal contact, while O-ring type fittings contain pressurized fluid by compressing an elastomeric seal. In either case, tightening threads between mating halves of the fitting (or fitting and component port) forces two mating surfaces together to form a high-pressure seal. All-metal fittings Threads on pipe fittings are tapered and rely on the stress generated by forcing the tapered threads of the male half of the fitting into the female half or component port, Figure 1. Pipe threads are prone to leakage because they are torque-sensitive — over-tightening distorts the threads and creates a path for leakage around the threads. Moreover, pipe threads are prone to loosening when exposed to vibration and wide temperature variations — certainly no strangers to hydraulic systems. Seepage around threads should be expected when pipe fittings are used in high-pressure hydraulic systems. Because pipe threads are tapered, repeated assembly and disassembly only aggravates the leakage problem by distorting threads, especially if a forged fitting is used in a cast-iron port. Thread sealant
Body
Nut
Sleeve
Seal area
Fig. 1. Pipe fittings have given way to newer fitting designs that simplify assembly, maintenance, and reduce or eliminate leakage. Shown is a 90° adapter elbow with pipe threads at one end that mount permanently into the component port. Other end of fitting uses straight-thread flare fitting for tubing connection.
compound is recommended for pipe fittings, which is still another reason why most designers consider them to be obsolete for use in hydraulic systems. Flare-type fittings, Figure 2, were developed as an improvement over pipe fittings many years ago and probably remain the design used most often in hydraulic systems. Tightening the assembly’s nut draws the fitting into the flared end of the tubing, resulting in a positive seal between the flared tube face and the fitting body. The 37° flare fittings are designed for use with thinwall to medium-thickness tubing in systems with operating pressures to 3000 psi. Because thick-wall tubing is difficult to deform to produce the flare, it is not recommended for use with flare fittings. The 37° flare fitting is suitable for hydraulic systems operating at temperatures from 165° to 400°F. It is more compact than most other fittings and can easily be adapted to metric tubing. It is readily available and one of the most economical. The flareless fitting, Figure 3, gradually is gaining wider acceptance in the U.S. because it requires minimal tube
Fig. 2. Flare-type fittings offer several design and performance improvements over pipe fittings and are used with thin-walled and medium-thickness tubing.
preparation. It handles average fluid working pressures to 3000 psi and is more tolerant of vibration than other types of all-metal fittings. Tightening the fitting’s nut onto the body draws a ferrule into the body. This compresses the ferrule around the tube, causing the ferrule to contact, then penetrate the outer circumference of the tube, creating a positive seal. Because of this, flareless fittings must be used with medium- or thick-walled tubing. O-ring type fittings Surprising as it may seem, leakage in hydraulic systems was licked more than a generation ago — or should have been. Although leak-free hydraulic operation has always been desirable, the need became more acute with higher operating pressures that became necessary during World War II, primarily in the hydraulic systems of military aircraft. Until then, common operating pressures had hovered around 800 to 1000 psi. The post-war era ushered in systems designed to operate at pressures to 1500 psi and higher on applications where rapid cycling and high shock pressures were common. It was not long until pressures climbed to 2500 and 3000 psi — which certainly are not uncommon today.
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FITTINGS
Body
Nut
Ferrule
O-ring seal Back-up washer
Lock nut
Seal area
Fig. 3. Flareless fittings offer advantages similar to those of flare fittings and are used with medium- to thick-walled tubing.
Faced with increased hydraulic fluid leakage brought on by higher pressures, a consortium of fittings manufacturers — working under the umbrella of SAE’s Committee on Tubing, Piping, Hoses, Lubrication, and Fittings — undertook solving the problem of hydraulic leakage. Their joint effort in the early 1950s culminated in the straight-thread design, which ultimately became known as the SAE straight-thread O-ring boss. Fittings that use O-rings for leaktight connections continue to gain acceptance by equipment designers around the world. Three basic types now are available: SAE straight-thread O-ring boss (also known as straightthread port) fittings, face seal or flatface O-ring (FFOR) fittings, and Oring flange fittings. The choice between O-ring boss and FFOR fittings usually depends on such factors as fitting location, wrench clearance, or individual preference. Flange connections generally are used for applications requiring tubing with an OD greater than 7/8-in. or those involving extremely high pressures. O-ring boss fittings seat an O-ring between threads and wrench flats around the OD of the male half of the connector, Figure 4. A leak-tight seal is formed against a machined seat on the female port. O-ring boss fittings fall into two general groups: adjustable and non-adjustable. Non-adjustable (or non-orientable) fittings include plugs and connectors. These are simply screwed into a port, and no alignment is needed. Adjustable fittings, such as elbows and tees, need to be oriented in a specific direction. The basic design difference between the two types is that plugs and connectors have no locknuts and require no
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Branch T
Straight connector
Fig. 4. Non-adjustable, left, and adjustable SAE straight-thread O-ring fittings offer ease of assembly and high potential for leak-tight connections.
back-up washer to effectively seal a joint. They depend on their flanged annular area to push the O-ring into the port’s tapered seal cavity and squeeze the O-ring to seal the connection. Adjustable fittings are screwed into the mating member, oriented in the required direction, and locked in place and when a locknut is tightened. Tightening the locknut also forces a captive backup washer onto the O-ring, which forms the leak-tight seal. Assembly is always predictable, because technicians need only make sure that the backup washer is firmly seated on the port’s spot face surface when the assembly is completed and that it is tightened properly. The FFOR fitting forms a seal between a flat, finished surface on the female half and an O-ring held in a recessed circular groove in the male half, Figure 5. Turning a captive threaded nut on the female half draws the two halves together and compresses the O-ring. Fittings with O-ring seals offer a number of advantages over metal-tometal fittings. While under- or overtightening any fitting allows leakage, all-metal fittings are more susceptible to leakage because they must be tightened to within a higher, yet narrower torque range. This makes it easier to strip threads or crack or distort fitting components, which prevents proper sealing. The rubber-to-metal seal in Oring fittings does not distort any metal parts and provides a tangible “feel” when the connection is tight. All-metal fittings tighten more gradually, so technicians may have trouble detecting when a connection is tight enough but not too tight.
O-Ring
Body
Nut
Sleeve
Fig. 5. Flat-face O-ring fitting uses O-ring in recessed groove in male half that mates with flat, smooth surface on female half.
On the other hand, O-ring fittings are more expensive than their allmetal counterparts, and care must be exercised during installation to ensure that the O-ring doesn’t fall out or get damaged when the assemblies are connected. In addition, O-rings are not interchangeable with all couplings. Selecting the wrong O-ring or reusing one that has been deformed or damaged can invite leakage. Once an O-ring has been used in a fitting, it is not reusable, even though it may appear free of distortions. Some manufacturers offer specially designed, high-pressure fittings that are equal in leak and weep resistance to FFOR fittings and interchangeable with a number of international fittings. Testing has shown these new designs to surpass all requirements with no evidence of leakage when exposed to vibrations up to 15 times more severe than those experienced on a typical hydrostatic drive. These designs may appear similar to standard fittings, but should not be mated with fittings from different manufacturers.
FITTINGS
Hydraulic flanges Fittings for tubing larger than 1-in. OD have to be tightened with large hexes which, in turn, require larger wrenches to enable workers to apply sufficient torque to tighten the fittings properly. To install such large fittings, system designers must provide the necessary space to give workers enough room to swing these large wrenches. Extensions would likely be needed for some workers to exert an applicable amount of torque. In addition, worker fatigue and strength must be considered. Fittings manufacturers have designed split-flange fittings so that they overcome both of these problems. Split-flange fittings, Figure 6, use an O-ring to seal a joint and contain pressurized fluid. An elastomeric O-ring rests in a groove on a flange and mates with a flat surface on a port — an arrangement similar to the FFOR fitting. The O-ring flange is attached to the port using four mounting bolts that tighten down onto flange clamps, thus eliminating the need for a large wrench when connecting large-diameter components. When installing flange connections, it is important to apply even torque on the four flange bolts to avoid creating a gap through which the O-ring can extrude under high pressure. The basic split-flange fitting consists of four elements: a flanged head connected permanently (generally welded or brazed) to the tube, an O-ring that Tubing
Flange
Bolt
Port surface
Clamp
Deep socket weld O-ring flange
Flange head adapter
90° socket weld elbow flange Split flange
Female thread O-ring flange connected to a hose assembly
Saddle weld flange
90° socket weld elbow flange
Fig. 7. Weld-type fittings used in conjunction with SAE 4-bolt flange clamp halves and O-ring flange head couplings offer convenience, economy for hydraulic connection assemblies.
fits into a groove machined into the end face of the flange, and two mating clamp halves with appropriate bolts to connect the split-flange assembly to a mating surface. All mating surfaces must be clean and smooth. Where perpendicular relationships are critical, all parts must meet appropriate tolerances. While 64-µ in. surface finishes are acceptable, most flange manufacturers prefer and recommend 32-µ in. finishes on mating surfaces to ensure leak-free connections. Joints are more likely to leak if either of the mating surfaces are scratched, scored, or gouged. Additionally, wear tends to accelerate on O-rings which are assembled against rough surfaces. In a properly designed split flange assembly, the flange shoulder protrudes approximately 0.010 to 0.030 in. beyond the clamp face to ensure adequate contact and seal squeeze
with the mating face, Figure 6. However, the clamp halves do not actually contact the mating surface. The most critical phase of assembling a splitflange fitting to its mating surface is to make certain that the four fastening bolts are tightened gradually and evenly in a cross pattern. Air wrenches should not be used because they are difficult to control and can easily over-tighten a bolt. Fully tightening one of the bolts while the others are still loose will tend to cause the flange to tip upward, Figure 8. This action pinches the Oring, and the joint can then be expected to leak. When the bolts are fully tightened, the flanges sometimes bend downward until they bottom on the port face, and the bolts bend outward, Figure 8. Should flanges and bolts bend, they tend to lift the flange off the shoulder; once again, the result will be a leaking joint.
0.010 in. to 0.030 in.
Fig 6. Properly designed and installed split-flange fitting has a uniform clearance of 0.010 to 0.030 in. between the port surface and clamp halves.
Fig. 8. Unevenly tightened split-flange bolts may cause flange to tip up and damage Oring, as shown at left, while over-tightened bolts, right, can bend the flange and bolts.
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Flow-control valves C
ontrolling flow of a fluid-power system does not necessarily mean regulating gpm from a valve. Flow rate can be specified three different ways, so it is important all involved in a project to be aware of how flow is to be specified or measured: Volumetric flow rate, Qv, expressed in units of in3/sec or min — or cc/sec or cc/min in SI metric measure — is used to calculate the linear travel speeds of piston rods or rotation speed of motor shafts. Weight flow rate, Qw, expressed in units of lb/sec or lb/min, is used to calculate power using CU units of measure. Mass flow rate, Q g , expressed in units of slugs/sec or slugs/min for CU measure — or kg/sec or kg/min in SI metric measure — is used to calculate
(a)
Eight basic configurations of valves can be used to control flow rate according to fluid volume, weight, or mass.
inertia forces during periods of acceleration and deceleration. Because they control the quantity of fluid that flows through the valve per unit of time, the same control valves are used for all three types of flow rates. Control of flow rate with valves There are eight types of flow-control valves: Orifices — A simple orifice in the line, Figure 1(a), is the most elementary method for controlling flow. Note that this is also a basic pressure control device. When used for flow control, the valve is placed in a series with the pump. An orifice can be a drilled hole in a fitting, in which case it is a fixed orifice; or it may be a calibrated needle valve, in which case it functions as a
Variable orifice (b)
Fig. 1. Simple fixed orifice (a) and variable orifice (b) flow controls.
variable orifice, Figure 1 (b). Both types are non-compensated flow control devices. Flow regulators — This device, Figure 2, which is slightly more sophisticated than a fixed orifice, consists of an orifice which senses flow rate, as a pressure drop across the orifice; a compensating piston adjusts to variations in inlet and outlet pressures. This compensating ability provides closer control of flow rate under varying pressure conditions. Control accuracy may be 0.5%, possibly less with specially calibrated valves which operate around a given flow-rate point. Bypass flow regulators — In this flow regulator, flow in excess of set flow rate returns to reservoir through a bypass port, Figure 3. Flow rate is controlled by throttling fluid across a variable orifice regulated by the compensator piston. The bypass flow regulator creates lower energy losses in a system. Pressure-compensated, variable flow valves — This flow control is equipped with an adjustable variable orifice placed in series with a compensator. The compensator automatically adjusts to varying inlet and load pressures, maintaining an essentially constant flow rate under these operating Handwheel
Fixed orifice
Inlet
Bypass
Fig. 2. Flow regulator adjusts to variations in inlet and output pressures.
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Regulated flow
Fig. 3. Bypass flow regulator returns excess flow from pump to tank.
F L O W - C O N T R O L VA LV E S Compensator piston
Drain
Adjustable Temperature-sensitive orifice element
Fig. 4. Pressure-compensated, variable flow control valve adjusts Fig. 5. Pressure-, temperature-compensated, variable flow control valve adjusts control orifice settings to offset effects of visto varying inlet and load pressures. cosity changes.
conditions to accuracy’s of 0.5%, Figure 4. Pressure-compensated, variable flow control valves are available with integral free-reverse-flow check valves and integral overload relief valves. Pressure- and temperature-compensated, variable flow valves — Because viscosity varies with temperature, so does the clearance between a valve’s moving parts. For this reason, output of a flow control valve may tend to drift with temperature changes. An attempt has been made to compensate not only for such temperature variations, but pressure variations as well, Figure 5. Temperature compensators adjust the control orifice setting to offset the effects of viscosity changes caused by temperature fluctuations of the fluid. Pressure compensators adjust the control orifice for pressure changes, as described above. Demand-compensated flow controls — Flow controls are available to bypass excess system flow to a secondary circuit, Figure 6. Controlled flow rate is ported to the primary circuit. Bypass fluid can be used for work functions in secondary circuits without affecting the primary one. There must be flow to the primary one. There must be flow to the primary circuit for this type of valve to function: if the primary circuit is blocked, the valve will cut off flow to the secondary circuit. Priority valves — A priority valve, Figure 7, is essentially a flow control valve which supplies fluid at a set flow rate to the primary circuit, thus functioning like a pressure-compensated flow control valve. Flow in excess of that required by the primary circuit bypasses to a secondary circuit at a pressure somewhat below that in the primary circuit. Should inlet or load
Fixed orifice
Secondary circuit
Fig. 6. Demand-compensated flow control bypasses full pump output to tank during idle portion of work cycle.
Primary circuit
Circuit 1
Pilot pressure
Circuit 2
Pilot pressure
Fig. 7. Priority valve supplies fluid at a set rate to a primary circuit.
Solenoid
Adjustable bypass orifice
Fig. 8. Deceleration valve slows load by being gradually closed by action of cam mounted on cylinder load.
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F L O W - C O N T R O L VA LV E S
Vp Load
Fig. 9. In meter-in control circuit, flow control valve is connected in series with directional flow control valve. Vp Load
Fig. 10. In meter-out control circuit, flow control valve is installed in cylinder return line.
Vp Load
Fig. 11. In bleed-off control circuits, flow control valve is mounted in parallel with cylinders.
pressure (or both) vary, the primary circuit has priority over the secondary as far as supplying the design flow rate is concerned. Deceleration valves — A deceleration valve, Figure 8, is a modified 2way, spring-offset, cam actuated valve used for decelerating a load driven by a cylinder. A cam attached to the cylinder rod or load closes the valve gradually. This provides a variable orifice which gradually increases backpressure in the cylinder as the valve closes.
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Some deceleration valves are pressure-compensated. This force-balance concept of control function also applies to flow rate control valves. Flow control methods There are three basic ways to control flow: meter-in, meter-out, and bleedoff. Meter-in control — The circuit in Figure 9 illustrates meter-in control. The flow control valve is placed in se-
ries with the directional control valve in the cylinder’s high pressure line. Thus, flow control valve A meters the amount of fluid entering the cap end of the cylinder. This type of control is best suited for resistive loads, where it is essential to control the speed at which a cylinder extends. Meter-out control — Here, the flow control valve is placed in the cylinder return line, Figure 10. The valve controls the rate of fluid flow from the cylinder to tank and is best used with overrunning loads. The valve controls the rate at which fluid leaves the head end of the cylinder. Thus, it controls the speed of the piston rod and load. Also, because it is placed in the return line, the overrunning load cannot force the piston rod to move at higher speed than that set by the flow control valve. Bleed-off control — This flow control device is placed in parallel with the cylinder, bypassing a part of the pump output flow to tank over the flow control valve, Figure 11. The flow control valve can be sized to handle bleed-off flow only, rather than entire pump output. Because the bleed-off valve is mounted in parallel (not in series) with the active elements, this flow control valve does not introduce a pressure drop into the active part of the circuit. Inlet pressure will be actual load pressure rather than the pressure of the relief valve setting. Figure 11 illustrates how a typical bleed-off circuit might be installed. Other flow controls Flow-dividers — A flow-divider valve is a form of pressure-compensated flow control valve which receives one input flow and splits it into two output flows. The valve can deliver equal flows in each stream or, if necessary, a preset ratio of flows. The circuit in Figure 12 shows how a flow divider could be used to roughly synchronize two cylinders in a meter-in configuration. Note that like all pressure- and flow-control devices, flow dividers operate over a narrow bandwidth rather than at one set point. Thus, there are likely to be flow variations in the secondary branches, and for this reason, precise actuator synchronization cannot be achieve with a flow-divider valve alone.
F L O W - C O N T R O L VA LV E S
Combining ports Cylinder 1
Control orifice
Cylinder 2
Dividing port
Restriction area
Flow divider
Fig. 12. Linear type flow divider splits input flow into two output flows.
Flow dividers can also be used in meter-out circuit configurations. Bleed-off does not affect the performance of a flow divider valve. Flow dividers can also be “cascaded,” that is, connected in series, to control multiple actuator circuits, Figure 13. Rotary flow dividers — Another technique for dividing one input flow into proportional, multiple-branch output flows is with a rotary flow divider, Figure 14. It consists of several hydraulic motors connected together mechanically in parallel by a common shaft. One input fluid stream is split into as many output streams as there are motor sections in the flow divider. Since all motor sections turn at the same speed, output stream flow rates are proportional and equal to the sum of displacements of all the motor sections. Rotary flow dividers can usually handle larger flows than flow divider valves. The pressure drop across each motor section is relatively small because no
Fig. 13. Flow dividers can be cascaded in series to control multiple actuator circuits.
energy is delivered to an external load, and is usually the case with a hydraulic motor. However, designers should be aware of pressure intensification generated by a rotary flow divider. If, for any reason, that load pressure in one or more branches should drop to some lower level or to zero, full differential pressure will be felt across the motor section(s) in the particular branch(es). The sections thus pressurized will act as hydraulic motors and drive the remaining section(s) as pump(s). This results in higher (intensified) pressure in these circuits branches. When specifying rotary flow dividers, system designers must be careful to minimize the potential for pressure intensification. Rotary flow dividers can also integrate multiple branch return flows into a single return flow. Pump control of flow rate — Pump control of flow rate presupposes the use of a variable-displacement pump. Non pressure-compensated pumps require an auxiliary control to stroke the pump-
ing element to vary the pump’s displacement. These auxiliary controls are available in hydraulic, pneumatic, mechanical, and electrical versions to match the needs of most control applications. Though pressure-compensated pumps are usually considered to be pressure control devices, designers must remember that flow control is achieved by reducing the displacement of the pump when a set pressure level is reached. Thus, a change in flow rate is involved. If this change occurs while the actuator is still moving, it will result in a change in actuator speed. The purpose of flow control is speed control. All the devices discussed in this section control the speed of the actuator by controlling the flow rate. Flow rate also determines rate of energy transfer at any given pressure. The two are related in that the actuator force multiplied by the distance through which it moves (stroke) equals the work done on the load. The energy transferred must also equal the work
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F L O W - C O N T R O L VA LV E S
done. Actuator speed determines the rate of energy transfer (i.e., horsepower), and speed is thus a function of flow rate, Figure 15. Directional control — Directional control does not deal primarily with energy control, but rather with directing the energy transfer system to the proper place in the system at the proper time. Directional control valves can be thought of as fluid
switches which make the desired “contacts,” that is, they direct the high-energy input stream to the actuator inlet, and provide a return path for the lower-energy oil. That is an important function can be inferred from the scores of different directional control valve configurations available in the marketplace. Moreover, it is of little consequence to control the energy transfer of the sys-
tem via pressure and flow controls, if the flow stream does not arrive at the right place at the right time. Thus, a secondary function of directional control devices might be defined as the timing of cycle events. Since fluid flow can be throttled in a directional control valve, some measure of flow rate or pressure control can also be achieved with these valves.
Proportional flow-control valves
Proportional flow control valves combine state-of-the-art hydraulic valve actuation with modern, sophisticated electronic control. These valves are helping fluid power designers to simplify hydraulic circuitry by reducing the number of components a system may require while, at the same time, substantially increasing system accuracy and efficiency. An electronically controlled, proportional flow control valve modulates hydraulic fluid flow in proportion to the input current it receives. The valves can easily control cylinders or smaller hydraulic motors in applications which require precise speed control, controlled acceleration and deceleration, or remote electrical programming. Most proportional flow control valves are pressure-compensated to minimize flow variations caused by changes in inlet or outlet pressure. An electronically actuated, proportional valve consists of three main elements: ● a pilot or proportional solenoid ● a metering area (where the valve spool is located), and ● a linear variable differential transformer (LVDT) electronic feedback device. Valve operation begins when it receives a signal from an outside controlling device such as a computer, programmable controller (PC), traditional logic relay, or potentiometer. The control device delivers electrical signals to the valve driver card, which, in turn, signals the valve pilot. Depending on
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As the spool shifts, its motion is detected and monitored very accurately by a LVDT. This data is fed back to the driver card where it is continuously compared with the input signals from the controller. If the two differ, the driver adjusts spool position until the two signals match.
the nature of the signal the valve pilot receives, it generates a magnified force which acts on the appropriate end of the valve spool to start it moving. As hydraulic force acts on the spool end, the spool shifts, gradually opening a flow path that lets fluid supplied by the pump flow to the appropriate actuator port. The important feature of this proportional valve is that all elements are proportional; thus, any change in input current changes force signals proportionately as well as the distance the valve spool will shift, the size of the flow path, the amount of fluid flowing through the valve, and finally the speed at which the actuator moves.
Pressure-compensated proportional flow control valves Pressure-compensated proportional flow control valves are 2-port valves in which the main control orifice is adjusted electronically. Similar to conventional pressure-compensated flow control valves, a pressure-compensated Fig. 14. Rotary flow divider consists of several fluid motors connected with parallel.
Q2
Q1
Q2
Q3
Q4
Vp pAp = F S
Q
Q = Vd t
Fig. 15. Actuator speed determines rate of speed transfer, which is a function of flow rate. For the equations in the drawing, p is fluid pressure, A p is piston area, and F is applied force; Q is flow rate, Vd is cylinder displacement, and t is time per stroke; v p is velocity of the piston, and S is cylinder stroke.
F L O W - C O N T R O L VA LV E S
Amplifier
A2
A
P1 P3
P2
B
Hydrostat LVDT Proportional selenoid A3
Cover C Pilot control X
Fig. 16. Operating circuit diagram for pressure-compensated flow-control valve.
Spring chamber
Sleeve
B
Main piston
proportional flow control valve maintains constant flow output by keeping the pressure drop constant across the main control orifice. The proportional valve, however, is different in that the control orifice has been modified to work in conjunction with a stroke controlled solenoid. In a 2-port pressure-compensated proportional flow control valve, an electrically adjustable control orifice is connected in series with a pressure reducing valve spool, known as a compensator or hydrostat, Figure 16. The compensator is located upstream of the main control orifice and is held open by a light spring. When the input signal to the solenoid is zero, the light spring force holds the main control orifice closed. When the solenoid is energized, the solenoid pin acts directly on the control orifice, moving it downward against the spring, to open the valve and allow oil to flow from port A to port B. At the same time, the LVDT provides the necessary feedback to hold position. In this case, the LVDT provides feedback to maintain a very accurate orifice setting. Pressure compensation is achieved by supplying a pilot passage from the front of the control orifice to one end of the hydrostat, A 2, and feeding a pilot passage beyond the control orifice to
the opposite end of the hyA Window opening drostat, A 3, assisted by the force exerted by the spring. Load-induced pressure at the outlet port or pressure devia- Fig. 17. Cross-sectional view of proportional flow tions at the inlet port are thus logic valve. compensated by the hydrostat, providing constant output flow. controller, Figure 17. The amplifier provides time conWhen an electrical signal is fed into trolled opening and closing of the ori- an electronic amplifier, the solenoid fice. For reverse free-flow, check valve and controller adjust the pilot pressure C, built into the valve, provides a flow supplied from port A to change spool path from port B to A. Proportional flow position. An LVDT then feeds back the control valves are also available with ei- position to the amplifier to maintain the ther linear or progressive flow charac- desired orifice condition for flow from teristics. The input signal range is the port A to port B. The proportional logic same for both. However, the progressive valve is available with either linear or flow characteristic gives finer control at progressive flow characteristics which the beginning of orifice adjustment. are adjusted by a 0 to 6-V, 0 to 9-V, or In case electrical power or feedback differential ±10-V command signal. is lost, solenoid force drops to zero and Because the valve remains relatively the force exerted by the spring closes unaffected by changes in system presthe orifice. When feedback wiring is sure, it can open and close the orifice in connected incorrectly or damaged a the same length of time. This maximum LED indicates the malfunction on the time can be changed on the amplifier card amplifier card. by adjusting a built-in ramp generator. The amplifier can be used in several Proportional flow logic valves ways. An external potentiometer can Proportional flow control logic make the orifice remotely adjustable valves are basically electrically ad- while maximum spool acceleration is justable flow controls that fit into a still limited by this internal ramp; or a standard logic valve cavity. The cover limit switch can be added to turn the and cartridge are assembled as a single ramp on and off. In case of power failunit, with the cover consisting of a pro- ure, the element will return to its norportional force solenoid and a pilot mally closed position. 1998/1999 Fluid Power Handbook & Directory
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Flushing procedures T
Fluid velocity is critical to successful removal of manufacturing and installation debris from hydraulic systems by flushing
olerances that exist in toExperience has shown that neiday’s high-pressure hyther of these flushing velocities is Dye Dye injector draulic systems demand sufficient to assure the cleanliness tight control of system contaminaof the ID of the system’s conducValve tion. Contamination that is built tors. A short review of basic fluid H2O into systems during manufacture dynamics explains why. Laminar flow and assembly must be removed beDimensionless N R s are used fore start-up to insure proper and (along with other factors) to classify predictable system performance fluid flow as either laminar, turbuTransition flow (or critical zone) throughout its service life. lent, or somewhere in between, FigA new or rebuilt hydraulic sysure 1. NR depends on the fluid’s vistem should be flushed before it becosity and velocity and the ID of the Turbulent flow comes operational. This procedure pipe. The flow condition that exists often is a contractual matter with Fig. 1. Simplified sketch of experiment that when NR is less than 2000 is termed the system manufacturer or re- Reynolds used to study and define the three laminar, signifying orderly flow builder. The concept of flushing is regimes of fluid flow. with parallel streamlines.1 When NR to loosen and remove contaminais greater than 3000, the flow betion particles inside the system by nore the interior cleanliness of the comes turbulent, defined as the condiforcing flushing fluid through it at system. Even if the tubing and con- tion when fluid streamlines are no high-velocity. In theory, this leaves ductors have been installed with the longer orderly. Between Reynolds the inside walls of the fluid conduc- greatest of visual care, the human eye Numbers of 2000 and 3000, flow exists tors at the same cleanliness level as can only see particles that are larger in transition. This sometimes is called the new fluid that will be installed. than 40 mm. That degree of cleanli- the critical zone. The hydraulic fluid velocity reThen, during normal operation, the ness is well below the needs of even system will experience only exter- the most crude and elementary hy- quired to achieve the textbook definition of turbulent flow is well within nally and internally generated con- draulic system. the recommended fluid velocity tamination that can be controlled guidelines for hydraulic fluid conwith appropriate filtration. The sys- How high a velocity? The critical variable in flushing to ductors. 2 This equation reinforces tem will not have to contend with achieve acceptable fluid and conduc- that statement: built-in contamination. NR 5 VD/n, Instructions for flushing usually tor cleanliness is fluid velocity. Traspecify a level of system cleanliness ditional flushing methods usually esthat must be achieved, and sometimes tablish this velocity in one of two where: V is the fluid velocity in ft/sec, D is the ID of the fluid conducspecify a fluid velocity that must be ways: tor in ft, and maintained during the flushing proce- ● the velocity must be high enough to n is the fluid kinematic viscosdure. Typical instructions state that achieve a Reynolds Number (N R) of ity in ft2/sec. flushing must be accomplished at 3000 or more, or normal system fluid velocities for a ● the velocity must meet or exceed certain period of time with a certain the system fluid’s normal operating First example Suppose that the NR is 3000, the conlevel of filtration. More stringent velocity. ductor is a 1-in. tube with a wall thickspecifications may call for a particuness of 0.049 in., and n is 1.28831024 lar fluid contamination level and re- 1T. B. Hardison, "Fluid Mechanics ft2/sec. Calculated fluid velocity then is quire documentation by fluid-con- for Technicians," Reston Publishing, 1977. 5.14 ft/sec, which corresponds to a tamination analysis flow rate of 10.24 gpm. One shortcoming of all these flush- 2Vickers Inc., "Vickers Industrial Note that the viscosity (and therefore ing methods is that they are based on Hydraulic Manual," 1989, page the Reynolds Number) of a typical hyprocedures to clean the fluid, but ig- C-16.
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FLUSHING PROCEDURES
draulic fluid is influenced by temperature and pressure. That is, the hotter the oil, the higher the NR for the same fluid velocity and pressure. The higher the pressure, the lower the NR for the same fluid velocity and temperature. Thus, specifying that NR should be 3000 is not a stringent requirement, but is well within the normal operating fluid velocities of a system. By definition, turbulent flow has been created because the fluid streamlines are no longer parallel, but sufficient fluid motion to clean the inside walls of the conductors has not been generated. Even at the recommended maximum fluid velocities and NRs for hydraulicsystem working conductors, fluid flow still is not turbulent enough to greatly affect contamination on conductor walls. Boundary-layer fluid at the interior surfaces of the fluid conductor remains undisturbed.
cosity as in the first example, but with the velocity increased to 20 ft/sec. This higher velocity results in a Reynolds Number of 11,671, which corresponds to a flow rate of 39.8 gpm. As NR increases, flow conditions go from laminar, through the critical zone, to turbulent. It has been proven empirically that once NR exceeds 3000, resistance to fluid flow is a combination of the effects of turbulence and of viscous drag at the conductor wall. (This region of viscous drag at the conductor wall is known as the viscous sub-layer.) There is a transition zone within the turbulent flow range where flow resistance goes from being governed by turbulence effects to being governed by the roughness of the inside wall of the conductor. This is shown clearly when inspecting the Moody diagram,3 Figure 2, which graphically demonstrates the relationship between Reynolds
Second example The NR for flow at normal system velocities next can be calculated using the same conductor size and kinematic vis-
3
Critical zone
Vennard & Street, "Elementary Fluid Mechanics," John Wiley & Sons, Inc., 1982, page 381. 4 Also from reference 3.
Number N R , friction factor f, and e, the roughness of the conductor’s inside surface. Resistance to flow through a fluid conductor — as represented by friction factor f — is only affected by the surface roughness of the fluid conductor4 when NR is above 4000. This means that the majority of the resistance to flow is created by turbulence effects. Only when N R is high enough so that surface projections of the conductor walls extend beyond the viscous sublayer does the surface come in contact with the turbulent flow and affect the pressure drop in the conductor. Surface roughness For drawn tubing, average surface roughness e is 0.000005 ft. If the conductor is the same 1-in. tubing with 0.049-in. wall thickness, ratio e/D will be 0.000067. The Moody diagram indicates that for this conductor, NR must be at least 25,000 before the inside surface exposes its resistance to fluid flow. Therefore, to ensure the inside wall of the conductor will be cleaned, NR must
Transition zone
0.1 0.09 0.08
Laminar flow
Rough 0.05 0.04
0.07 0.06
0.03 0.02
0.05
0.015 0.04
f
0.01 0.008 0.006 0.03 0.004
e/D
0.025 0.002 0.02
0.001 0.0008 0.0006 0.0004
Smooth 0.015
0.0002 0.0001 0.0003
0.01 0.009 0.008 103
104
105
106
0.00001
107 0.00005
0.0001 108
NR
Fig. 2. Modified Moody diagram represents relationship of friction factor f, Reynolds Number NR, and conductor surface roughness e.
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be greater than 25,000. For flow to be fully in the rough zone of turbulent flow, NR must be greater than 3.253107. Using 1.28831024 ft2/sec, the same fluid kinematic viscosity as in the first example, a NR of 25,000 corresponds to a fluid velocity of 42.8 ft/sec, or a flow rate of 85 gpm — still easily attainable with conventional hydraulic pumps. Real-world systems It can be argued that if the walls of a conductor are not greatly affected by normal system fluid velocities, contaminants lodged there will have little chance of entering the fluid stream. This may be partially true but the argument applies only to smooth, straight conductors at steady flows and pressures. It is not representative of normal installations that combine straight runs, bends, and numerous fittings where flow patterns are only predictable empirically, and where pressure fluctuations and spikes are commonplace. Depending on the severity of service that the system will experience, pressure spikes will dislodge contaminants held in the walls of the conductors and between fitting interfaces. Remember that in critical systems, 3to 25-µm particles can impact system performance significantly. The only way to guarantee that conductor contamination (that can be released at any time during operation) does not affect system performance is to protect each component with a filter, an option so costly that it would not be used in most systems. Although flushing hydraulic system conductors at the normal system operating-fluid velocities can provide fluid velocities higher than flushing at a NR of 3000, the inside wall of the conductors still will not be cleaned. High-velocity/high-pressure flushing Flows that produce N R s .25,000 are needed to ensure that conductor walls are exposed to turbulent flow. Because system conductors may consist of pipe, tube, and/or hose and associated fittings, the specification of a contractual number for N R is diffi-
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Calculated flushing velocities and NRs for 200 ft of Schedule 80 pipe Pipe size - in. /2 /4 1 11/2 2 1 3
D
Test pressure - psi
Relative roughness - e/D
Maximum velocity - fps
0.546 0.742 0.957 1.500 1.939
3500 2900 2700 2100 1830
0.0033 0.0024 0.0019 0.0012 0.0009
65 74 84 96 117
cult and still does not guarantee that conductors will be cleaned. The best one can do is establish conditions that will maximize N R. These conditions are: the highest possible velocity at the lowest possible fluid viscosity. Limiting factors are the conductor’s pressure rating and the fluid’s maximum operating temperature. When flushing a system, the valving and actuators must be “jumpered” for safety reasons so the only resistance to fluid flow is the pressure drop in the conductors and fittings. When flow becomes turbulent, the pressure drop is proportional to the square of the velocity. Extrapolating this relationship to its maximum, the highest possible velocity occurs when the pressure drop in the conductor generated by fluid flow is equal to the maximum test pressure of the conductor. Flushing a system at these high flows and pressures has the added advantage of expanding and contracting the conductors and fittings as the pressure fluctuates while inducing highly turbulent flow. This optimizes the flushing action. By equating the pressure drop in a conductor to the maximum pressure rating of that conductor, the maximum fluid velocity possible, along with the corresponding N R, can be calculated. The temperature of the fluid directly affects its viscosity and is the other variable that can control N R . Flushing pressure also affects viscosity, but this is hard to quantify because pressure in the pipe being flushed will vary from maximum at the pumping source to atmospheric at the conductor outlet. The equation used to calculate head loss in the turbulent zone is:
Flushing flow rate - gpm 48 100 188 528 1077
NR 23,000 35,000 52,000 93,131 146,722
hl 5 fLV2/2D, where: hl stands for head loss, f is the friction factor found in the Moody diagram, L is the conductor length in ft, V is the fluid velocity, and D is the conductor’s ID in in. This equation will calculate the maximum velocities and N R s that can be achieved for a given maximum pressure. Determining f for pipe flow requires iterative calculations using the Moody diagram. Given the pressure rating, ID, length, and relative roughness of the conductor, assume an f and then calculate the fluid velocity. next calculate N R and determine a new f from the Moody diagram. Repeat the calculation until f converges. The table above contains velocities and NRs that have been calculated for 200 feet of Schedule-80 pipe using the maximum test pressure for the pipe and a surface roughness of 0.00015 ft for wrought iron pipe. These calculations did not take into account the pressure drop produced by the various fittings normally used, so the values for the attainable fluid velocities and NRs are optimistically high. Also, special fluids with lower viscosities or flushing at higher temperatures to reduce the fluid viscosity can increase NR. The values determined for maximum flushing velocity and flow rate indicate that some of these conditions, mainly for lines smaller than 3/4 in., can be satisfied using conventional high-pressure pumps of appropriate size, although it may be difficult to induce the pressure fluctuations needed to dislodge contaminants. For systems with larger conductors, special methods must be used to achieve the necessary pressures, fluid velocities, and NRs to properly flush the lines.
Heat exchangers H
ydraulic systems can use either of two methods of load control: the energy-loss method, in which flow to the actuator is limited by valving, or the volume-control method, in which the stroke of a variable-displacement pump controls the rate at which pressure fluid is sent to the actuator. While the latter method is inherently more efficient, a system using that design principle is more costly and reacts less rapidly. Most industrial hydraulic systems are designed using the energy loss method. Such systems cost less to build and are more responsive because system energy is immediately available. But because of the inherent poor efficiency of these systems, energy losses in the form of heat can sometimes approach the prime mover’s nameplate horsepower. As an example of this heat build-up, even well-designed electrohydraulic servovalve or proportional valve systems may convert 60 to 80% of input horsepower to heat. Well-designed non-servo systems may produce heat losses of 20 to 30%. Some hydraulic system heat is desirable to bring fluid up to operating temperature. Cold hydraulic oil has a higher viscosity than warm oil. So maintaining an operating temperature of 1008 F would cause sluggish operation and excessive pressure drop in a system designed to operate at 1408 F. When a system begins operation on a cold winter morning, for example, the oil should be allowed to warm until it reaches a temperature where heat is generated at the same rate as system heat radiating into the atmosphere or other cooling medium. If heat generation exceeds the radiation rate, the excess heat can cook the oil, start oil decomposition, form varnish on system component surfaces, and begin deteriorating system seals. Excess heat sooner or later spells trou-
ble for any hydraulic system. Too much heat breaks down oil, damages seals and bearings, and increases wear on pumps and other components. The solution to these problems is the inclusion of a properly sized heat exchanger as a component of the system. Thermodynamics Heat is a form of energy that transfers from one region to another because of a temperature difference (gradient) between the regions. Just as liquids naturally flow downhill, a temperature gradient is a condition in which heat energy naturally flows from the hotter region to the cooler region. The rate of heat transfer is an important consideration in determining how quickly a heat exchanger can remove heat from a system. A small heat exchanger with a high heat transfer rate can remove as much heat from a system as a larger heat exchanger with a lower heat transfer rate. The defining equation for any heat exchanger is q5UADT, where q represents the heat load transferred in BTU/hr, U stands for the overall heat transfer coefficient in BTU/hr-ft2-8 F, A denotes the heat transfer surface in ft2, and DT signifies the fluid temperature difference in 8 F. These three factors take varying forms depending on the heat exchanger and the application involved. Inspecting the equation shows that as each term on the right increases, more heat will transfer. If a larger surface area comes in contact with the heated fluid, more heat is removed. A temperature gradient between the heated oil and the cooler region toward which the heat flows also increases q.
Most hydraulic systems generate substantial heat, so most can benefit from using a heat exchanger. Here’s why and how The same is true for the overall heat transfer coefficient, U. Heat dissipates from a fluid system through natural and forced convection. Natural convection occurs as heat moves from system components into the surrounding atmosphere because of the temperature gradient. In smaller hydraulic systems, temperatures generally are lower than in larger systems, and heat transfer from the oil to tubing and other component surfaces often provides sufficient cooling. But if this natural convection cannot remove enough generated heat, a heat exchanger must be installed to control system temperature. The heat exchanger uses forced convection to remove heat. Another mode of heat transmission, radiation, may often occur, but its effect is small and usually can be ignored. Generally, a heat exchanger is necessary for a hydraulic system if: ● a specific oil temperature limitation is necessary to stabilize oil viscosity ● cycle dwell time is a major portion of the total duty cycle, especially in systems with fixed-displacement pumps, and ● problems with hot oil or shortened oil or seal life have been encountered with similar systems. Heat-transfer mechanisms Considering shell-and-tube heat exchangers, U is composed of several heat-transfer mechanisms. The first is the convective heat transfer from the hot fluid in the shell to the tube wall. This can be called the hot fluid thermal resistance, which depends primarily on physical and thermal fluid properties. Geometric tube bundle patterns (square or triangular centerline spacing when viewed from the tube ends) help the oil flow turbulently over the tubes. The baffles inside the shell increase the distance the oil must travel through the exchanger. This increases
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oil velocity, which helps destroy the laminar-flow film layer to promote effective flow of heat. The turbulence and increased velocity enhance heat transfer values. The second heat transfer mechanism is thermal conductance through the tube wall. Most modern heat exchanger tubes are built from materials that have high thermal conductivity. As a result, the thermal conductance factor is consistently quite high. The third mechanism is the convection of heat from the tube wall to the cooler fluid in the tube side. This mechanism acts in much the same manner as the hot fluid thermal resistance. Use of multi-pass flow patterns takes advantage of the fluid velocity and turbulence for increased U values. The last conditions to influence the U values are fouling factors that may occur on both sides of the tubes over a period of time. When a flowing fluid system deposits material or scale on boundary walls, they become fouled. This layer: ● acts as an insulator ● increases effective fluid-film thickness, and ● affects fluid velocity distribution near the tube wall. As the scale deposit builds, U decreases accordingly. Mechanical and/or chemical cleaning can remove the scale when excessive deposit degrades heat exchanger efficiency. Water cooling Shell and tube heat exchangers port cooling water inside the tubes (tube side) of the exchanger; the heated oil flows around the tubes in the shell side. These heat exchangers are made of red brass, copper, cast iron, admiralty brass, stainless steel, aluminum, or other special metals. They have an outer flanged shell with end bonnets appropriately sealed to the shell ends. Within the shell is a precise pattern of tubing that runs the length of the shell and terminates in tube sheets or end plates. Tube ends are mechanically metallurgically fastened to the tube sheets, which seal each end of the shell. The tubes that make up the tube bundle run through varying numbers of baffle plates that support them and cause the oil to flow at right angles to
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the tubes as oil travels from one end of the shell to the other. Many tube bundles for hydraulic applications are permanently sealed in the shell. Models with removable tube bundles are available but are more expensive and have different sealing conditions at the shell ends.
structions to flow, called turbulators, Figure 2, help destroy laminar flow and remove the insulating fluid layer. Although these obstructions to flow increase pressure drop through the heat exchanger, the improvement in heat transfer more than compensates for the higher pressure drop.
Tube-pass configurations New designs Heat exchangers are available in 1-, Shell-and-tube heat exchangers have 2-, and 4-pass configurations, Figure 1. been the mainstay of the industry for These multiple passes result from di- over 50 years. Recently, however, new viders in the bonnets that mate with designs have been developed to insegmenting bars on the tube sheets to crease effectiveness, provide equivainitially force the cooling medium only lent heat transfer surface in a smaller through a fraction of the tubes. This envelope, yet reduce cost. causes the water to flow one, two, or A new extended-surface design, Figfour times the length of the heat ex- ure 3, adds a great number of fins to the changer before it leaves the outlet. external sides of tubes. These fins proThe equation introduced earlier ap- vide more surface area, improve the plies to shell-and-tube heat exchangers, heat transfer coefficient, and reduce the but several factors expand and compli- amount of heat transfer surface recate the basic equation; this quired, thereby reducing the size of the expansion is of no interest here. Some heat exchanger when compared to basic rules still do apply: the larger the area, A, the greater Baffles Shell side flow the flow, q; this is logical because Tubeside flow larger heat exchangers transfer more heat. The flow pattern through the heat exchanger can vary Single pass the heat removed from the system. Shell side flow Fluid can flow in laminar, transitional, and turbulent modes. Tubeside When flow is lamiflow nar and at low velocities in a tube, there is little or no fluid Two pass movement immediately next to the tube Shell side flow wall. This layer of stagnant fluid hinders heat transfer and can be thought Tubeside flow of as insulation. Faster, turbulent flow has no smooth Four pass velocity gradient. The jumbled, tumbled flow pattern Fig. 1. One-, two-, and four-pass heat exchangers increase can disrupt much of cooling rate, promote turbulence, and destroy the insulating the stationary fluid film layer that exists with laminar fluid flow. Baffle plates speed film. Built-in ob- oil flow through the shell side for the same reason.
H E AT E X C H A N G E R S
older shell-and-tube versions. However, because of the greater internal heat-transfer surface area, pressure drop is correspondingly greater than in the older versions. The extended-surface design shelland-tube heat exchanger is also available with a spring-loaded pressuresurge protector. If system pressure surges, a bypass valve opens to protect the tubes from excess pressure damage. Another newer heat exchanger design on the market is the brazed-plate type, Figure 4. The heat-transfer surface consists of a series of stainlesssteel plates, each stamped with a corrugated pattern designed for a combination of high strength, efficiency, and fouling resistance. The number and design of the plates vary according to the desired heat-transfer capacity. The plates are stacked together with thin sheets of copper or nickel between each plate. The plate pack, end plates, and connections are
brazed together in a vacuum furnace to secure the plates together at the edges and at all contact points. Inlet and outlet connections are available in a number of different styles. The brazed-plate heat exchanger is compact, rugged, and provides high heat transfer capability. Its heat-transfer surface area is concentrated in a very small volume. The corrugations in the plates induce turbulent flow to increase heat transfer and reduce fouling. A brazed-plate heat exchanger holds about one-eighth the liquid volume of a thermally comparable shelland-tube counterpart. Normally the hot and the cold fluids make only one pass through this exchanger. The incoming fluids are directed through alternate parallel channels created by the stacked channel plates. The single pass through the unit means the flow passage is as long as the heat exchanger is high. This short flow path offsets any pressure drop which is caused by the turbulation. These same heattransfer principles (q5UADT) govern Fig. 2. Turbulators help tumble fluid in tubes to destroy lami- p e r f o r m a n c e o f nar flow, but increase pressure drop of fluid that flows brazed-plate and shellthrough them. The increased heat transfer efficiency with tur- and-tube heat exchangbulent flow generally justifies the higher pressure drop. ers. Stainless steel construction of the brazed-plate design allows design flow velocities as high as 20 ft/sec. These higher velocities, coupled with turbulent flow, provide heat-transfer rates of three to five times those of shell-andtube heat exchangers. The higher heat-transfer rates (U) mean less area (A) is required for a given capacity (q). Tests show the brazed-plate design can handle particles Fig. 3. Extended-surface is an outgrowth of the original shell- up to about 0.040 in. and-tube design. A large number of fins, mechanically fixed without plugging. A to the exterior of the tubes, increase surface area for greater strainer should be heat transfer capability. The finned tube bundle can be as used if larger particles much as 40% smaller than the conventional shell-and-tube may be encountered. variety with the same cooling capacity, yet uses about half as Because of their conmuch water. struction, these heat
Fig. 4. Brazed-plate heat exchangers provide large surface area for improved heat transfer in a small volume.
exchangers must be cleaned chemically rather than mechanically. Air-cooled heat exchangers When the air sink is the choice to receive waste heat, a familiar heat exchanger like that shown in Figure 5 is used. Even though it radiates heat only to a small degree, it is traditionally called a radiator instead of a convector. It convects heat and uses the same equation examined earlier to describe performance. Hot oil passes through the tubes of these heat exchangers, and turbulators help break up laminar flow to promote efficient heat transfer from the fluid to the tube wall. The tube metals also have high thermal conductivity. Stagnant air around tube exteriors presents a problem in the effort to increase heat transfer. Still air is a poor conductor and has a high thermal resistance that limits heat transfer. As might be supposed, increasing air flow over the tubes helps decrease this thermal resistance. The amount of decrease again depends on whether the air flow is laminar or turbulent. In either case, the air still inhibits heat transfer because it is not as good a conductor of heat as water or oil.
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As with shell-and-tube heat exchangers, increasing area A of the heat transfer surface increases heat transfer capability. Fins, physically fastened to the tubes, increase surface area and, as an added benefit, help break up laminar air flow. Here are some considerations that help determine heat exchanger core configurations: ● oval tubes promote more turbulent flow at lower liquid flow rates than round tubes do ● round tubes give higher flow rates and lower pressure drops than with oval tubes ● materials generally are admiralty brass, brass, aluminum, or steel. Choice can depend upon structural needs or service environment, and ● fins vary in heat transfer capability and cleanability. Fin types are flat plate, humped, and louvered, ranked in order of increasing ability to generate turbulent flow and in decreasing order of cleanability. Application considerations When determining the heat exchanger requirements of a hydraulic system, consider these factors: ● what temperatures can the oil and system components tolerate? ● how much heat does the system generate, and
Fig. 5. Often found on mobile applications, air-to-oil heat exchangers (commonly known as radiators) use surrounding air to receive excess heat. Specially designed fins promote turbulent air flow to improve heat transfer capability.
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● should a water- or air-cooled heat exchanger be chosen? Determining hydraulic cooling needs can be confusing, because heat generation may vary as a machine progresses through different cycles, ambient temperatures shift, and other unexpected changes that produce additional heat occur. The first step in determining heat transfer requirements of a system is to know the temperature limits of the system fluid and components. This information is available from the respective suppliers. Next, estimate total system heat output before building the system. To arrive at this figure, many designers use the nominal percentage of input horsepower method. Using this method, heat is estimated to be a percentage of total system inefficiency, based on individual component inefficiencies, plumbing surface area, and past experience. The total of these percentages is multiplied by the input horsepower and converted to BTU/min, BTU/hr, or kW. For example, a 300-hp hydraulic system is assigned an efficiency rating of 70% or 210 hp. The remaining 30%, 90 hp, is lost as heat. The 90 hp converts to 3820 BTU/min or 229,200 BTU/hr, or 67.14 kW. After the system is built, heat rejection is determined by measuring the fluid temperature rise during system operation over a period of time. Temperature rise/unit of time, along with system capacity, determines heat input. Heat exchangers should be installed in the tank return line to reduce their need to withstand high pressures. A bypass valve protects them from high pressure surges, such as those commonly experienced during cold starts. Bypass valves also can be specified based on fluid temperature. In this manner, fluid bypasses the heat exchanger until warming to a predetermined temperature, which closes the bypass valve and routes fluid through the heat exchanger. An alternate piping method uses a separate circuit from the reservoir with a small pump to drive the oil through the heat exchanger. Filters often are incorporated into these side circuits — or kidney loops, as they are sometime called.
Oil-to-air heat exchangers are most commonly used on mobile applications. Engine coolant is available for oil cooling, but the heat must eventually be removed into the air sink through the radiator. Some installations use a separate section of the radiator in front of the engine-driven fan for oil cooling. When considering application and sizing of heat exchangers, the steadystate temperature of the hydraulic fluid and the time it takes to arrive at that temperature must be used. A heat exchanger manufacturer can assist with application and selection. When contacting the representative, be prepared to provide the following: ● oil heat load in BTU/min ● oil flow in gpm ● maximum oil temperature ● ambient air temperature during system operation ● maximum allowable pressure drop ● system elevation, and ● environmental conditions such as salt, airborne particles, or chemicals that can affect the system. If the heat exchanger is liquid cooled, the supplier will also need to know the following: ● cooling water inlet temperature ● water flow rate, and ● maximum allowable pressure drop. Most manufacturers’ literature includes examples, steps, and simplified equations to help properly size heat exchangers. Now the heat exchanger choice can be made: will it be an air- or watercooled version? Generally, air-cooled heat exchangers are more expensive than water-cooled ones on a per-unit basis, because on other than mobile applications, the air-cooled version requires a fan/motor package. But there are more considerations: ● electric costs to run the fan ● costs to purchase water, pump it, and perhaps treat it ● fan noise and hot air exhaust ● clogged cooling surfaces if debris becomes airborne, and ● special vibration mounts and flexible piping. A properly sized and applied heat exchanger can save time, money, and repair costs. Many fluid power systems should not operate without one.
Hydraulic filtration H
ydraulic fluids perform four basic functions. Their primary function is to create force and motion as flow is converted to pressure near the point of use. Second, by occupying the space between metal surfaces, the fluid forms a seal, which provides a pressure barrier and helps exclude contaminants. A third function — often misunderstood — is lubrication of metal surfaces. The fourth and final function provided by hydraulic fluid is cooling of system components. If any one of these functions is impaired, the hydraulic system will not perform as designed. Worse yet, sudden and catastrophic failure is possible. The resulting downtime can easily cost a large manufacturing plant thousands of dollars an hour. Hydraulic-fluid maintenance helps prevent or reduce unplanned downtime. It is accomplished through a continuous program to minimize and remove contaminants. Aside from human interference, the most common source of system impairment is fluid contamination. Contamination can exist as solid particles, water, air, or reactive chemicals. All impair fluid functions in one way or another. Sources of contaminants Contaminants enter a hydraulic system in a variety of ways. They may be: ● built-in during manufacturing and assembly processes ● internally generated during normal operation, and ● ingested from outside the system during normal operation. If not properly flushed out, contaminants from manufacturing and assembly will be left in the system. These contaminants include dust, welding slag, rubber particles from hoses and seals, sand from castings, and metal debris from machined components. Also,
when fluid is initially added to the system, a certain amount of contamination probably comes with it. Typically, this contamination includes various kinds of dust particles and water. During system operation, dust also enters through breather caps, imperfect seals, and any other openings. System operation also generates internal contamination. This occurs as component wear debris and chemical byproducts from fluid and additive breakdown due to heat or chemical reactions. Such materials then react with component surfaces to create even more contaminants. Contaminant interference In broad terms, contaminant interference manifests itself as either mechanical or chemical interaction with components, fluid, or fluid additives. Mechanical interactions, Figure 1, include blockage of passages by hard or
soft solid particles, and wear between particles and component surfaces. Chemical reactions include: formation of rust or other oxidation, conversion of the fluid into unwanted compounds, depletion of additives — sometimes involving harmful byproducts — and production of biochemicals by microbes in the fluid. Any of these interactions will be harmful. Without preventive measures and fluid conditioning, their negative effects can escalate to the point of component failure. One of the most common failure modes is excessive wear due to loss of lubrication. Lubrication and wear The pressures required in modern hydraulic systems demand sturdy, precisely matched components. And precision machining leaves very small clearances between moving parts. For
(a)
(b)
(c)
(d)
Stress raisers caused by particle collisions
Stress cracks
Fig. 1(a). Three-body mechanical interactions can result in interference. (b) Two-body wear is common in hydraulic components. (c) Hard particles can create three-body wear to generate more particles. (d) Particle effects can begin surface wear.
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Particle-generated wear The symptoms of wear are diminished system performance and shorter component life. In pumps, wear may first be detected as reduced flow rate. This is because abrasive wear has increased internal clearance dimensions. Sometimes called increased slippage,
Water contamination of liquids Water in oil-base fluids can be just as destructive as particle contaminants. And water exclusion is very difficult to
100 80 60
SAE 20 + rust and oxidation inhibitors
40 30
2
6083
28
Synthetic hydrocarbon
83
0
17
46
ste r
example, it is not uncommon for control valves to have pistons and bores matched and fitted within a mechanical tolerance of 60.0002 in. (two ten-thousandths of an inch). In metric units, this is about 5 mm (five millionths of a meter). In modern electrohydraulic devices, tolerances may be even tighter and clearances can be less than 1 mm. The surface finishes on high-pressure bearings and gears can result in rolling clearances as small as 0.1 min. The hydraulic fluid is expected to create a lubricating film to keep these precision parts separated. Ideally, the film is thick enough to completely fill the clearance between moving parts, Figure 2. This condition is known as hydrodynamic or full-film lubrication and results in low wear rates. When the wear rate is kept low enough, a component is likely to reach its intended service-life expectancy — which may be millions of pressurization cycles. The actual thickness of a lubricating film depends on fluid viscosity, applied load, and the relative speed of the two dynamic surfaces. In many applications, mechanical loads are so high that they squeeze the lubricant into a very thin film, less than 1-min. thick. This is elastohydrodynamic (EHD) or thinfilm lubrication. If loads become high enough, the film will be punctured by the asperities of the two moving parts. The result is boundary lubrication.
Die
Fig. 2(a). In highly magnified representation, ideal lubrication forms full film between two surfaces. (b) During boundary lubrication, metal asperities of surfaces make physical contact and are torn off.
this condition means that the pump is less efficient than it was when new. When pump flow rate decreases, the fluid system may become sluggish, as evidenced by hydraulic actuators moving slower. Pressure at some locations in the system also may decrease. Eventually, there can be a sudden catastrophic failure of the pump. In extreme cases, this can occur within a few minutes after initial start-up of the system. In valves, wear increases internal leakage. The effect this leakage has on the system depends on the type of valve. For example, in valves used to control flow, increased leakage usually means increased flow. In valves designed to control pressure, increased leakage may reduce the circuit pressure set by the valve. Silting interference, Figure 3, causes valves and variableflow pump parts to become sticky and operate erratically. Erratic operation shows up as flow and pressure surges, causing jerky motion in actuators.
6
Metal-to-metal contact area
Fig. 3. Even accumulations of soft particles can cause silting interference if they are the same size as the clearance dimension.
560
(b)
SA E5
Lubricant film
Component and system designers try to avoid boundary lubrication by making sure that fluid has the proper viscosity. However, viscosity changes as the fluid temperature changes. Also, loads and speed may vary widely during normal operating cycles. Therefore, most hydraulic components operate at least part of the time with only boundary lubrication. When that happens, parts of moving surfaces contact each other and are torn away from the parent material. The resulting particles then enter the fluid stream and travel throughout the system. If not removed by filtration, they react with other metal parts to create even more wear. Fluid chemists continually try to minimize potential lubrication problems by improving fluids with additives. Viscosity-index (VI) improvers are added to help keep viscosity stable as temperature changes. Antiwear additives increase film strength. If very heavy loads will be applied, the fluid should contain an extreme pressure (EP) additive that reacts with metal surfaces to form a hard protective film. For fluids in circulating systems, defoamant, demulsifier, detergent, or dispersant may be added. Rust and oxidation (RO) inhibitors are used in most hydraulic fluids because air and water are always present to some extent.
Relative humidity -- percent
(a)
20
SAE 20 Polyphenyl + phosphorus ether 10 10
20
40
60
80 100 200 400 Water concentration -- ppm
600
1000
Fig. 4. Water solubility in various hydrocarbon-based hydraulic fluids.
Neopentyl polyol ester 2000
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Descriptive factors for particles Particulate contamination may be characterized by the following identification factors. ● agglomeration — the tendency of particles to bond together. This action is generally detrimental in fluid-contamination control. ● compaction — degree of packing from sedimentation process. As void spaces decrease and bulk density increases, silt condition intensifies. ● concentration — weight per unit volume of fluid or number of particles greater than given size per unit volume of fluid. ● density — mass of particle per unit of volume. Density affects the rate at which particles settle out from the fluid.
● dispersion — the tendency of particles to remain separated. This is a factor in particle separation and analysis. ● hardness — resistance to abrasion and the particle’s potential to abrade exposed surfaces. ● settling — terminal velocity of particles controls the degree of particle suspension provided by the flowing fluid. ● shape — degree of irregularity of particle structure or topography. A factor important to the cutting or abrading ability of the particle. ● size — structural extent of particle as defined by geometric, derived, and hydrodynamic diameters. Such diameters have significance on a statistical basis. ● size distribution — frequency of occurrence of each particle size in the
population. Cumulative particle size distribution curves are the most popular type in fluid-contamination control. ● size limits — the size range in which only fractureless deformation occurs and the lower filtration limit of interest. ● state — condition ;where size and shape cannot be altered without forceful shearing of crystalline or molecular bonds. A concept important to the understanding of particle generation and growth. ● transport — the life force needed to overcome the buoyant weight of the particle. When this is achieved, the flow conduit does not retain particles on its surface.
accomplish. Because of its affinity for other liquids, water is present in some concentration in most hydraulic systems, Figure 4. The hygroscopic nature of liquids causes them to pick up a certain amount of water simply from contact with humid air. When condensation occurs in a reservoir, with subsequent mixing into the base liquid, more water can be added to the system. Water can even enter with new oil. An oil barrel stored outdoors in a vertical position is likely to have rainwater collect around its bung. With changes in ambient temperature, some of this moisture can be sucked into the barrel. Eventually, this water enters the system when the reservoir is filled from that barrel. Besides these natural phenomena, there are several system- and maintenance-related sources of water. In machine tool applications, a good deal of water-base coolant can enter hydraulic systems through breather caps and imperfect seals. Worn and damaged heat exchangers can allow cooling water to leak through seals and ruptured lines into the oil system — and vice versa. Each fluid has its own saturation level for water. Below the saturation level, water will be completely dissolved in the other fluid. For oil-base hydraulic fluids, the saturation level is likely to be in the range of 100 to 1000 parts per million (0.01% to 0.1%) at
room temperature. At higher temperatures, the saturation level is greater. Above the saturation level, water becomes entrained, meaning that it takes the form of relatively large droplets. This also is called free water. Sometimes these droplets combine and precipitate to the bottom of the reservoir. At other times, due to churning or other mixing action, undissolved water is emulsified so that it exists as very fine droplets suspended in the oil.
characteristic sounds of cavitation may be noticeable when this happens. The result is cavitation damage on the interior surfaces of hydraulic components because the metal has fatigued.
Water’s mechanical effects When water concentration in hydraulic oil reaches 1 or 2%, the response of a hydraulic system may be affected. If water alters the hydraulic fluid’s viscosity, the operating characteristics of the hydraulic system change. When the rate of water influx is swift, poor system response could be the first indication that water is present. Cavitation is another symptom of water in the fluid. Because the vapor pressure of water is higher than hydrocarbon liquids, even small amounts of water in solution can cause cavitation in pumps and other components. This occurs when water vaporizes in the low pressure areas of components, such as the suction side of a pump. Vaporization is followed by the subsequent violent collapse of vapor bubbles against metal surfaces in these areas. The loud
Emulsified water in oil Tiny water droplets may be emulsified or suspended in oil-base fluids. Evidence of this is when the fluid appears cloudy or milky. Sometimes an oil/water emulsion is so tight that it is very difficult to separate the two fluids — even with the addition of a coalescing agent formulated to do this. While this is desirable in emulsion-type hydraulic fluids, it is highly undesirable for ordinary oil-base fluids. Some fluid additives encourage emulsification, while others (demulsifiers) discourage it. The viscosity of a water emulsion can be much different from its original base liquid. As noted earlier, lower viscosity reduces the thickness of lubricating films, leading to increased wear of surfaces in dynamic contact. If free water is present in hydraulic fluid and the system operates at temperatures below 328 F, ice crystals may form. These crystals can plug component orifices and clearance spaces. In hydraulic systems, this will cause slow or erratic response. Without fluid analysis to warn of water’s presence and appropriate con-
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trol measures, water content probably will increase to the point where these and other symptoms appear. Other symptoms include evidence of chemical-reaction products and eventually, component failures. Chemical reactions due to water Water reacts with almost everything in a hydraulic system. Water promotes corrosion through galvanic action by acting as an electrolyte to conduct electricity between dissimilar materials. The most obvious sign is rust and other oxidation that appears on metal surfaces. The inside top surface of the reservoir often is the first place rust becomes visible, but such rust still may go undetected unless the reservoir is drained and opened for servicing. Also, the time it takes for rust to form depends partly on the original surface treatment used to protect the metal used to build the reservoir. Unfortunately, before rust is noticed in the reservoir, water probably has damaged other system components. An inspection of failed bearings and other components may point to corrosion damage. Corroded aluminum and zinc alloys could have a whitish oxide film. Steel bearing and gear surfaces may show signs of rust and pitting. Water’s reaction with oxidation inhibitors produces acids and precipitates. These water-reaction products also increase wear and interference. At high operating temperature (above 1408 F), water reacts with and actually can destroy zinc-type antiwear additives. For example, zinc dithiophosphate (ZDTP) is a popular boundary lubricant added to hydraulic fluid to reduce wear in highpressure pumps, gears, and bearings. When this type of additive is depleted by its reaction with water, abrasive wear will accelerate rapidly. The depletion shows up as premature component failure, resulting from metal fatigue and other wear mechanisms. Water frequently can act as an adhesive that causes smaller contaminant particles to clump together in a larger mass. These gluey masses may slow down a valve spool, or cause it to stick in one position. Or these clumps could plug component orifices. In any case, the result is erratic operation or a complete system failure.
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Microbial growth Over time, water contamination can lead to the growth of microbes — minute life forms such bacteria, algae, yeasts, and fungi — in the hydraulic system. And the presence of air exacerbates the problem. Microbes range in size from approximately 0.2 to 2.0 mm for single cells, and up to 200 mm for cell colonies. Left unchecked, microbes can destroy hydraulic systems just as they destroy living organisms. Under favorable conditions, bacteria can reproduce (doubling themselves) as rapidly as every 20 minutes. Such exponential growth can form an interwoven mat-like structure that requires significant shear force to break up. This resistance quickly renders a fluid system inoperable. Besides their mass volume, bacteria produce acids and other waste products which attack most metals. When this happens, fluid system performance is degraded, and components fail more rapidly. Evidence of microbial contamination The first indication of microbial contamination may be the foul odor that comes from waste and decomposition products of the microbes. Fluid viscosity may increase due to the mass of material produced by these organisms. At the same time, the fluid may have a brown mayonnaise-like appearance, or slimy green look. Unfortunately, by the time these symptoms appear, system components and the fluid itself may be severely damaged. This could require a major overhaul or replacement of the system. Properly selected filters will remove microbes. But without adding biocides (substances capable of destroying these living organisms) to the fluid, fastgrowing microbes can place a heavy load on system filters. Combined with wear debris and chemical reaction products, microbial contamination can result in rapid plugging of filter elements, requiring frequent replacement. Water and air are essential for microbe growth. Eliminating water and air from a fluid minimizes microbial problems. But some systems use water as the base fluid, and air is very difficult to exclude from fluids in operating hydraulic systems. With water and air present, microbes can usually find
Return line
Suction line
Diffuser
Strainer
Fig. 5. Baffle in reservoir slows fluid velocity so large contaminant particles can settle to bottom. Diffuser prevents churning action which might entrain air in fluid.
some fluid component to feed their growth. When water can not be controlled by exclusion or removal, biocides should be added to the fluid. A biocide combined with properly selected water-absorption filters can help minimize chemical-reaction byproducts and microbial contamination. Exclusion practices The first defense against fluid contaminants is preventing their entry into a hydraulic system. After that, removing contaminants before system startup prevents much damage that can occur early in a system’s life. Thereafter, well-planned routine maintenance will maintain the fluid in peak condition. Here are some of the initial positive steps that can be taken: ● fit the reservoir with baffles and return-line diffusers, Figure 5, to prevent churning that whips air into the fluid ● equip the reservoir with a breather having an air-filter element with a rating of at least 99% efficiency at 2 µm ● make sure all fittings are properly tightened (besides causing leakage, loose fittings can allow airborne dust to be sucked into the system) ● flush the system thoroughly before it goes into service ● prefilter fluid before filling the reservoir (it should be as clean as your specification for the system fluid) ● inspect filter indicators to make sure they are working ● use boots and bellows to protect cylinder rods and seals ● replace filter elements before the filter bypass valve opens; otherwise, the system will operate with no filtration ● replace any worn seals and hoses
H Y D R A U L I C F I LT R A T I O N
promptly; if not done, the negative effects are the same as loose fittings ● practice good housekeeping whenever a system is opened for maintenance; protect replacement components from contamination, and ● analyze fluid regularly to detect problems such as overheating, leaking water, clogged heat exchangers, additive breakdown, etc.
so that smaller particles remain in suspension. This is as true for the reservoir as elsewhere in the system. A tapered reservoir bottom will help prevent the collection of smaller contaminant particles due to its reduced bottom surface area and tendency to extend the turbulence effect. As in many design projects, reservoir construction and piping configuration involves compromises. Outgassing can be thought of as the inverse of settling. If fluid turbulence is low enough to prevent mixing action, dissolved air can come out of suspension and rise to the surface of a liquid. Whether the air actually leaves the liquid or not depends on the relative surface tensions and partial pressures of the air and the liquid. The lower the turbulence in the reservoir, the more likely
it is that a contaminant will leave the fluid by way of outgassing or settling. Natural mechanisms, such as settling and outgassing, cannot by themselves reduce contamination to an acceptable level. In the absence of filtration and separation devices, the only alternative is to replace the fluid at periodic intervals. Even with adequate filtration, fluid replacement cannot be postponed forever. This certainly is true for automotive lubricants, and points out a fundamental fact of fluid life. There is an economic trade-off between the cost of buying, installing, and servicing filters and separators, and the cost of replacing the hydraulic fluid more often.
Removal mechanisms Once contamination is in the fluid, it may be reduced and controlled by settling, outgassing (e.g. in aerated liquids), filtration/separation, and fluid replacement. The first two mechanisms — settling and outgassing — occur naturally, but their effect can be enhanced Fluid conditioning objectives by controlling the fluid environment The objective of hydraulic fluid conthrough system design. The latditioning is to lower total operter two also require human inating costs. If the system can volvement, again during system meet or exceed minimum standesign or in maintenance actividards for fluid cleanliness, one Specify NO Manufacturer's Optimum ties after installation. or more of these intermediate required stated sensitivity tradeoff? cleanliness For settling to occur, a congoals can be achieved: for component in circuit taminant must have a density ● reduce maintenance requireYES branches greater than the fluid transportments for the fluid system and ing it. The lower the density of a components Decide on Decide contaminant particle, the more filter housing ● improve the performance of on filter options buoyant it will be in the fluid. the system and its fluid locations The flow rate of the fluid also ● assure the quality of the final helps determine how quickly a product by improving machine Write filter Determine initial contaminant will settle. A conoperation, and manufacturer's contamination model code taminant transported by a fluid ● enhance safety and/or reduce level in reservoir will stay in suspension if the risk of injury to personnel (for flow velocity supplies enough example, by eliminating the Estimate contaminant lifting force to overcome gravneed for maintenance on or ingression rate, bypass ity. If flow is turbulent, it is around operating equipment). flow, and upstream more likely that contaminants Appropriate fluid conditioncontamination levels will stay in suspension. ing increases the mean time beAs mentioned earlier, the tween hydraulic component Calculate the minimum reservoir can be designed with failures. Still, this benefit has to beta ratio required for stated cleanliness levels baffles and return-line diffusers be properly balanced against the to reduce fluid velocity enough cost of purchasing the filters, reso that larger particles will setplacing elements, and maintainDecide on tle. On the other hand, contamiing filtration equipment. Carefilter media to be used nants must remain in suspenful filtration system design and sion if they are to be transported component selection will help to a filter for removal. This is minimize these costs. The best Decide on Environmental particularly important in fluid way to optimize the benefit/cost filter housing variables style and size lines and components, where trade-off is to follow sound particle settling can cause unpractices for the selection of filpredictable contaminant reters, elements, and filter media. Evaluate moval rates, or silting interferOne general process is illuscost/benefit tradeoffs ence between moving parts. trated in the filter-specification Therefore, system designers flow chart, Figure 6. want a reasonable degree of turMany questions should be bulence in the hydraulic system Fig. 6. Suggested steps in hydraulic-filter selection process. answered regarding contami1998/1999 Fluid Power Handbook & Directory
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nant removal: ● how clean does the fluid be have to? ● what size particles must be removed? ● how many particles within a given size range need to be removed? ● how efficient must the filter media be in terms of the percentage removal of a given size range — and in terms of dirtholding capacity? ● will the fluid contamination stabilize at an acceptable level for a given combination of filters and media? Component sensitivity As the flow chart implies, specifiers need to have a feel for the sensitivity of hydraulic components to contaminants of various sizes and concentrations. Designers and users have observed that some components are more sensitive to contaminants than others. For example, they may have seen a certain pump quickly fail, while another type lasts for months in the same system. They also probably have noticed that higher pressures and flow rates tend to make all components wear out more quickly.
Component sensitivity
Severity of duty cycle
Component service life
Contaminant removal
Those who are particularly observant may have noticed that the higher the concentration of airborne contaminants around systems, the sooner they fail. These factors combine to influence the service life of components, Figure 7. Another point is that filter media with small pore sizes frequently are more costly, and must be replaced more often than coarser media. For practical economic reasons, designers must find a compromise between costly ultrafine filtration and the cost of early component failures. This compromise is to have fluid only as clean as it needs to be, not as clean as possible. Designers tend to rely on their own experience as well as information from component manufacturers to determine how clean hydraulic fluid needs to be. Some conservative manufacturers assume that worst-case conditions exist and specify a very low acceptable level of contamination for their components. Others take a middle-of-the-road approach, and specify cleanliness for more or less average conditions.
Better lubricity
Contaminant ingression
Thicker lube film
Additional information sources Manufacturers’ recommendations can be augmented by information that is available from other sources. For example, OEMs and research laboratories have carried out projects to analyze the sensitivity of pumps, valves, and other components to contaminants. As a result, guidelines and standards for hydraulic-fluid cleanliness have been published. These guidelines attempt to interrelate diverse factors such as: ● fluid lubricity (e.g., water-base fluids have lower lubricity than oil) ● abrasiveness of the contaminants commonly found in hydraulic systems ● system duty cycle and cycle rate (high pressure and high cycle rates, combined with contaminants, lead to earlier fatigue failures) ● component replacement cost ● design life objective in terms of mean time before failure (MTBF); a common goal today is 10,000 hours or more, and ● degree of risk associated with contaminant-related failures (high risk of personal injury or high cost of lost production dictates a need for cleaner fluid.) Fluid variables and system variables both have an effect on a component’s sensitivity to contamination. This sensitivity eventually is reflected in system performance, Figure 8. The International Standards Organization (ISO) recommends cleanliness levels for various types of components, see the table below. The levels are stated in terms of industry standards that have been recognized for the past 20 years. Many fluid power designers apply these Fluid cleanliness required for typical hydraulic components
Fig. 7. Contaminant sensitivity is a major factor in component service life. Component type Fluid variables Lubricity Temperature Concentration of contaminants
Component Contaminant sensitivity
System variables Pressure Duty cycle Ingression rate Failure tolerance
Flow rate Force Torque Speed MTBF Performance
Fig. 8. Complex relationship among operating variables, component’s sensitivity to contaminants, and hydraulic system performance involves many factors.
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Fluid classification ISO code
Servovalves
14/11
Vane and piston pumps/motors
16/13
Directional and pressure control valves
16/13
Gear pumps/motors
17/14
Flow control valves and cylinders
18/15
Aircraft test stands
13/10
Injection molding
16/13
Metal working
17/14 - 16/13
Mobile equipment
18/15 - 16/13
New unused oil
18/15
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recommendations as rules of thumb. Many specifiers now accept and use ISO 4406 (see table on page A/99) as a means of designating the fluid cleanliness required for their systems. Importance of records Still, component manufacturers’ and industry guidelines should be modified by experience. That requires gathering enough operating and maintenance data over sufficient time and from enough systems to provide confidence to make decisions. The data gathered should include the results of regular fluid analysis on systems. The categories of data collected might include: Fluid variables — flow, pressure, temperature, and viscosity for circuit branches with the most sensitive or expensive components. Fluid analysis — particle counts in various size ranges (e.g., >2, >5, >15, >25, >50, and >100 mm), spectrochemical analysis (e.g., most likely metals and other contaminants), and water content (% by volume). Filtration information — model number and manufacturer for the filter(s) and element(s) protecting the circuit for which other data was gathered; element performance ratings in terms of beta ratios and dirt-holding capacity. Maintenance data — date system placed in service; dates and descriptions of routine maintenance performed (including element replacements); reading of the filter element condition indicator (e.g., “needs replacement” or “in bypass”); dates and descriptions of component failures, including manufacturers’ names and model numbers; failure mode analysis (e.g., fracture, corrosion, wear, etc.) also would be very helpful in determining if contamination was a factor in any failures. PC data base and statistical analysis programs also can be used to correlate failures with fluid contamination levels. This will create a picture of the contamination tolerance of the most sensitive components. It also allows for the calculation of MTBFs for specific components, certain circuit branches, or the system as a whole. Obviously, this is data the user must collect. Still, manufacturers can monitor warranty claims as an opportunity to capture some of this data, and create
a clearer picture of component sensitivity. That may cover only the first year or two of service. A close relationship with customers and distributors can provide an opportunity to gather similar data over longer periods of time as replacement parts are ordered. Contamination — dynamic, not static Another reason for regular fluid analysis is that the contamination level changes with time, and varies by location in the system. At any point, the amount of contamination present in the fluid depends on three factors: 1. How contamination much was in the fluid when the system was started 2. How much was added to the fluid from all sources during operation (Ingression rate is the term used to describe the amount of contaminant entering the fluid per unit of time.) 3. How much contamination left the fluid due to all removal mechanisms (settling, and filtration or separation) These three factors account for the total mass of contaminant in a system at any time. That mass can be calculated using a material-balance equation: CT = Ct + Ca 1 Cs, where: C is contaminant T is any point in time t is time since start of process Ca is amount added since t Cs is amount removed since t The term material balance is used because the equation calculates the net difference between the amount of material or contaminant entering and leaving the fluid, and adds this difference to what was already there. The calculation applies to a specific location in the system. In a circulating system, contaminants not removed will appear at the filter inlet again, along with new contaminant added to the fluid. This is called a multipass system because the fluid and contaminant make multiple passes through the filter. As a result, the contaminant concentration in the system fluctuates continuously. If we consider the initial start-up of a system, the contaminants already present are there as a result of manufacturing processes or have entered with new fluid. (Each milliliter of fluid out of the original barrel typically contains at least
2500 particles that are 5 mm and larger in diameter.) A few minutes after startup, the particulate level will be considerably higher due to flushing action of the fluid as it flows through new components and piping to pick up debris. Eventually, more particles enter the system through the reservoir breather and imperfect seals. Still more will be added over time due to internal wear. Estimating ingression rate Assume that a new hydraulic system has been flushed properly before being put into service. If the system has a multipass filtration system with a given flow rate, the eventual stabilized level of contaminants will depend on the system’s ingression rate and filter media removal efficiency. If filter efficiency is too low, the contaminant level will continue to increase due to the wear particles generated within the system and new particles entering from outside the system. (This is the scenario in most automobile lubrication systems, and why motor oil should be changed periodically.) If filter efficiency is high enough, the contaminant level will decrease and become stabilized, extending the service life of the hydraulic fluid. Because operating conditions vary, this is a kind of dynamic stability. The contaminant level varies within a range determined by these conditions. Therefore, to select the appropriate filter media, it is necessary to have some idea of the ingression rate. Of course the ingression rate probably varies at different locations in the system, and depends on these factors: ● concentration of ambient airborne contaminants (which enter through worn filler/breathers, loose fittings, leaking seals, etc.) ● use or absence of an air-filter element in the reservoir breather ● number of components in the system or circuit branch ● types of components that make up the system, particularly if there are rotating components such as pumps and motors (some types wear faster than others) ● fluid velocity (because higher velocity often may accelerate wear — after flushing is completed) ● system pressure (because higher pressure also tends to increase wear rates) ● fluid temperature (excessive heat can
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cause fluid and additives to break down, creating contamination), and ● the filter media used (more-efficient media results in lower contaminant levels and reduced wear rates). This parade of factors makes accurate estimation of ingression rates difficult. Still, tables and nomographs, such as Figure 9, that provide a range of values for this variable are available from a number of sources. To be more spe-
cific, an estimate can be made by conducting particle counts on fluid samples taken from a system with known operating conditions and filtration efficiency. (In multibranch circulating systems, the reservoir frequently is picked as a convenient location from which to take samples.) Then, by using a simple filtration model based on the materialbalance equation, an ingression rate can be inferred. Filtration specifiers can
10
10
9
10
8
Number of particles per minute > indicated size
10
Mobile equipment
7
10
Metalworking plants 6
10
5
10
Assembly operations 4
10
3
10
2
10
1
5
10
20
30
40
50
60 70
80 90 100
Particle size — µm
Fig. 9. Nomograph suggests average ingression rates of various contaminant particle sizes under three general sets of operating conditions for hydraulic equipment.
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construct their own estimates using the same technique. Such an estimate may be more accurate than one of the averages published in a table. Computerized filtration models also are available that allow a large number of variables to be quickly manipulated in a kind of ingression-rate What if? analysis. Cleanliness reference In order to detect or correct problems, a contamination reference scale is used. Particle counting is the most common method to derive cleanliness level standards. Very sensitive optical instruments count the number of particles in various size ranges in a fluid sample. These counts are reported as the number of particles greater than a certain size found in a specified volume. The ISO 4406 cleanliness level standard has gained wide acceptance in most industries today. A modified version of this standard references the number of particles greater than 2, 5, and 15 micrometers in a known volume — usually 1 milliliter or 100 milliliters. (The number of smaller-size particles helps predict silting problems. A high number of larger particles might indicate catastrophic component failure.) Filter media The filter media is that part of the element which actually contacts contaminant and captures it for subsequent removal. The nature of the particular filter media and the contaminant-loading process designed into the element explains why some elements last longer in service than others. During manufacture, media usually starts out in sheet form, then is pleated to expose more surface area to the fluid flow. This reduces pressure differential across the element while increasing dirt-holding capacity. In some designs, the filter media may have multiple layers and mesh backing to achieve certain performance criteria. After being pleated and cut to the proper length, the two ends are fastened together using a special clip, adhesive, or other seaming arrangement to form a cylinder. The most common media include wire mesh, cellulose, and fiberglass composites, or other synthetic materials. Filter media is generally classified as either surface- or depth-type.
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Surface media For surface-type filter media, the fluid stream basically flows in a straight path through the element. Contaminant is captured on the surface of the element which faces the fluid flow. Surface-type elements are generally made from woven-wire cloth. Because the process used to manufacture the wire cloth can be controlled very accurately, and the wire is relatively stiff, surface-type media have a consistent pore size. This consistent pore size is the diameter of the largest hard spherical particle that will pass through the media under specified test conditions. However, during use, the build-up of contaminant on the element surface will reduce the pore size and allow the media to capture particles smaller than the original pore-size rating. Conversely, particles (such as fiber strands) that have smaller diameters but greater length than the pore size may pass downstream through surface media. Depth media For depth-type filter media, fluid is forced to take convoluted indirect paths through the element. Because of its
construction, depth-type media has many pores of various sizes formed by the media fibers. This maze of multisized openings throughout the material traps contaminant particles. Depending on the distribution of pore sizes, the media can have a very high capture rate for very small particle sizes. The two basic media that are used for depth-type filter elements are cellulose (or paper) and fiberglass. The pores in cellulose media tend to have a broad range of sizes and are very irregular in shape due to the irregular size and shape of the fibers. In contrast, fiberglass media consist of man-made fibers that are very uniform in size and shape. These fibers are generally thinner than cellulose fibers, with a consistently circular cross-section. The differences between these typical fibers account for the performance advantage of fiberglass media. Thinner fibers can provide more pores in a given area. Furthermore, thinner fibers can be arranged closer together to produce smaller pores for finer filtration. Dirt-holding capacity, as well as filtration efficiency, are improved as a result.
ISO 4406 range numbers Range number 24 23 22 21 20 19 18 17 16 15 14 13 12 11 10 9 8 7 6
Number of particles per ml More than 80,000.00 40,000.00 20,000.00 10,000.00 5,000.00 2,500.00 1,300.00 640.00 320.00 160.00 80.00 40.00 20.00 10.00 5.00 2.50 1.30 0.64 0.32
Up to and including 160,000.00 80,000.00 40,000.00 20,000.00 10,000.00 5,000.00 2,500.00 1,300.00 640.00 320.00 160.00 80.00 40.00 20.00 10.00 5.00 2.50 1.30 0.64
15-µm size range indicates the quantity of larger particles present, those which contribute greatly to possible catastrophic component failure. To identify a cleanliness level, the number of particles in the sample for each of the three measured sizes is referred to the ISO 4406 chart, and given an appropriate range number. If a fluid sample contained between 1300 and 2500 2-µm and larger particles (range 18); between 320 and 640 5-µ m and larger particles (range 16); and between 40 and 80 15-µ m and larger particles (range 13); the sample would be classified as 18/16/13. Note that the numbers that make up the ISO cleanliness-code classification will almost never increase as the particle size increases. Most manufacturers of hydraulic (and load-bearing) equipment conduct tests and then specify an optimum or target cleanliness level for their components. Exposing components to hydraulic fluid with higher than optimum contamination levels may shorten the component’s service life. It always is best to consult with component manufacturers and obtain their written fluidcleanliness-level recommendations. This information is needed in order to select the proper level of filtration. It also may prove useful for any subsequent warranty claims, as it may draw the line between normal operation and excessive or abusive operation.
Particle counting Knowing the cleanliness level of the hydraulic fluid in a system is the basis for selecting contamination-control measures. Particle counting is the most common method of deriving cleanliness-level standards. The Multipass Test Very sensitive optical instruThe filtration industry uses the ISO ments count the number of par- 4572 Multipass Test Procedure (also ticles in various size ranges in a recognized by ANSI and NFPA) to measured fluid sample. These evaluate filter element performance. counts are reported as the num- During the Multipass Test, Figure 10, ber of particles greater than a certain size found in a speciFresh contaminant fied volume of fluid. Multipass contaminant The ISO 4406 CleanlinessN down Level Standard is accepted in Fluif most industries today. A sample widely used, modified version Test of this standard references the filter number of particles greater N up than 2, 5, and 15 µ m in a known volume — usually 1 or 100 milli-liters. The number Fluif sample of particles greater than 2 and 5 µm is a reference point for silt particles, those which can Fig. 10. Simplified representation of compocause clogging problems. The nent arrangement for Multipass Test. 1998/1999 Fluid Power Handbook & Directory
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fluid circulates through the test circuit under precisely controlled and monitored conditions. The differential pressure across the element being tested is continuously recorded, while a constant amount of contaminant is injected upstream of the element. On-line laser particle sensors measure the contaminant levels upstream and downstream from the test element. (Note that the results of this test are very dependent on flow rate, type of contaminant, and terminal pressure differential.) The Multipass Test determines three important element-performance characteristics: 1. Dirt-holding capacity. 2. Pressure differential of the test filter element. 3. Separation or filtration efficiency, expressed as a Beta Ratio. Beta Ratio The Beta Ratio (also known as the filtration ratio) is a measure of a filter element’s particle-capture efficiency. Therefore, it is a performance rating. Here is how a Beta Ratio is derived from Multipass Test results. Assume that 50,000 particles, l0 mm and larger in size, were counted upstream from the test filter, and 10,000 particles in that same size range were counted downstream from the filter. Substituting in the equation: bx = (NU)/(ND) where: x is a specific particle size, NU is the number of particles upstream, and N D is the number of particles downstream. Cleanliness required for typical hydraulic components Component Servovalves Proportional valves Vane and piston pumps and motors Directional and pressure control valves Flow control valves and cylinders New unused fluid
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ISO code 16/14/11 17/15/12 18/16/13 18/16/13 20/18/15 20/18/15
General comparison of filter media
Media material
Capture efficiency
Dirtholding capacity
Differential pressure
Service life
Initial cost
Fiberglass
High
High
Moderate
High
Moderate to high
Cellulose (paper)
Moderate
Moderate
High
Moderate
Low
Wire mesh
Low
Low
Low
Moderate
Moderate to high
Therefore, b10 = 50,000/10,000 = 5 This result would be read as “Beta ten equal to five.” Now, a Beta Ratio number alone means very little. It is a preliminary step to finding a filter’s particle-capture efficiency. This efficiency, expressed as a percent, can be found by a simple equation: Efficiencyx = 100 (1-1⁄b) Efficiency10= 100 (1-1⁄5) = 80% In this example, the particular filter element tested was 80% efficient at removing 10-mm and larger particles. For every five particles in this size range introduced to the filter, four were trapped in the filter media. Contaminant loading The term contaminant loading in a filter element refers to the process of filling and blocking the pores throughout the element. As contaminant particles load the element’s pores, fewer open paths remain for fluid flow, and Beta Ratios and capture efficiencies Beta Ratio (at a given particle size) 1.01 1.10 1.50 2.00 5.00 10.00 20.00 75.00 100.00 200.00 1000.00
Capture efficiency (at the same particle size) 1.0% 9.0% 33.3% 50.0% 80.0% 90.0% 95.0% 98.7% 99.0% 99.5% 99.9%
the pressure required to maintain flow through the media increases. Initially, the differential pressure across the element increases slowly because plenty of open pores remain for fluid to pass through. The gradual pore-blocking process has little effect on overall pressure loss. Eventually, however, successive blocking of media pores significantly reduces the number of pores open for flow. The differential pressure across the element then rises exponentially as the element nears its maximum life. As the element continues to load with contaminant, the pressure differential across the filter will continue to increase. This goes on until the bypass valve (if installed) opens, the element (if without bypass protection) fails structurally, or the clogged element is replaced. The quantity, size, shape, and arrangement of the pores throughout the element accounts for why some elements last longer than others. Consider two of the most common filter media: cellulose and fiberglass, For a given media thickness and filtration rating, there are fewer pores in cellulose media than fiberglass. Accordingly, the contaminant-loading process would block the pores of the cellulose media element more quickly than an identical fiberglass media element. Multi-layer fiberglass media elements are relatively unaffected by contaminant loading for a longer time. The upstream media has relatively larger pores to capture larger particles; the downstream media layer with very small pores captures the greater quantity of small particles present in the fluid. Filter-element life profile Every filter element has a characteristic relationship between pressure differential and contaminant loading. This relationship can be defined as the filter
H Y D R A U L I C F I LT R A T I O N
element life profile. The actual life profile obviously is affected by the system operating conditions. Variations in the system flow rate and fluid viscosity affect the clean pressure differential across the filter element and have a well-defined effect upon the actual element life profile. The filter element life profile is very difficult to evaluate in actual operating systems. The system’s ratio of operating time to idle time, the duty cycle, and the changing ambient contaminant conditions all affect the life profile of the filter element. In addition, precise instrumentation for recording the change in the pressure loss across the filter element seldom is available. Most machinery users and designers simply specify filter housings with differential-pressure indicators to signal when the filter element should be changed. Multipass Test data can be used to develop the pressure-differential-versus-contaminant-loading relationship. As mentioned, operating conditions such as flow rate and fluid viscosity affect the life profile of a filter element. Life profile comparisons can be made only when all these operating conditions are identical, and the filter elements are the same size. Under those conditions, the quantity, size, shape, and arrangement of the pores in the filter element determine the characteristic life profile. Filter elements manufactured from cellulose media, single-layer fiberglass media, and multi-layer fiberglass media have very different life profiles. Filter housings The filter housing is the pressure vessel which contains the filter element. It usually consists of two or more subassemblies, such as a cover (or head) and a removable bowl that allows access to the element. The housing has inlet and outlet ports that enable fluid to enter and leave. Housing options may include bypass valves and elementcondition indicators. Primary concerns in the housing-selection process include mounting methods, porting options, indicator options, and pressure rating. Except for the pressure rating, all depend on the physical system arrangement and the preferences of the system designer. Pressure
rating of the housing is far less arbitrary; it is determined by system needs before the housing style is selected. Pressure ratings Location of the filter in the circuit is the primary determinant of pressure rating of the component. Filter housings are generically designed for one of three locations in a circuit: suction, pressure, or return lines. One characteristic of these locations is their maximum operating pressures. Suction and return line filters are generally designed for lower pressures — 500 psi or less. Pressure filter locations may require ratings from 1500 to 6000 psi. Note that it is essential to analyze the circuit for pressure-spike potential as well as steady-state conditions. Some housings have restrictive or lower fatigue pressure ratings. In circuits with frequent high-pressure spikes, another type housing may be necessary to prevent fatigue-related failures. Bypass valves Bypass valves open flow paths around filter elements to prevent their collapse or bursting when they become heavily loaded with contaminant. As contaminant builds up in the element, the differential pressure across the element increases. At a pressure well below the failure point of the filter element, the bypass valve opens, allowing flow to go around the element. Some bypass valve designs have a bypass-totank option. This directs the unfiltered bypass flow back to the tank through a third port, preventing unfiltered bypass fluid from entering the system. Bypass valves also prevent pump cavitation when used with suction line filters. When specifying a bypass-type filter, it generally can be assumed that the manufacturer has designed the element to withstand the bypass valve differential pressure when the bypass valve opens. Note that some of the upstream contaminant particles also bypass the filter element with the fluid and enter the downstream system. When this happens, the effectiveness of the filter element is compromised and the attainable system fluid cleanliness degrades. Other filters are designed specifically with no bypass valve (sometimes called a blocked bypass). They prevent
any unfiltered flow from going downstream, thus protecting servovalves and other contaminant-sensitive components. In filters without bypass valves, higher collapse-strength elements may be required, especially if installed in high-pressure locations. When specifying a non-bypass filter design, make sure that the element has a differentialpressure rating close to the maximum operating pressure of the system, and that the filter has a condition indicator. After a housing style and pressure rating are selected, the bypass valve setting needs to be chosen. This setting must be established before sizing the filter housing. Everything else being equal, the highest bypass cracking pressure available from the manufacturer should be specified. This will provide the longest element life for a given filter size. Occasionally, a lower setting may be selected to help minimize energy loss in a system, or to reduce backpressure on another component. In suction filters, either a 2- or 3-psi bypass valve is used to minimize the chance of potential pump cavitation. Element-condition indicators The element-condition indicator signals when the element is loaded to the point that it should be cleaned or replaced. The indicator usually has calibration marks which also indicate if the bypass valve has opened. The indicator may be linked mechanically to the bypass valve, or it may be an entirely independent differential-pressure sensing device. Indicators may give visual or electrical signals or both. Generally, indicators are set to trip at a differential pressure anywhere from 5 to 25% below that which opens the bypass valve. Sizing housing and element The filter housing size should be large enough to achieve at least a 2:1 ratio between the bypass valve setting and the pressure differential of the filter with a clean element installed. For longer element life, this ratio should be 3:1 or even higher. Referring to typical flow/differentialpressure curves from a manufacturer’s catalog, the filter specifier needs to know the operating viscosity of the fluid and the maximum (not the average) flow rate to assure that the filter
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does not spend a high portion of time in bypass due to flow surges. This is particularly important in return-line filters, where flow multiplication from large cylinders may increase the return flow compared to the pump’s flow rate. Always consider ambient temperature conditions when sizing filters. Low ambient temperatures may increase fluid viscosity to the point where pressure differential across the filter assembly also may increase considerably. If a filter was fitted with a 50-psi bypass valve, the initial (clean) pressure differential should be no greater than 25 psi and preferably 162⁄3 psi or less. These pressures are calculated from the 3:1 and 2:1 ratios of the 50-psi bypass setting and initial pressure differential. Standard filter assemblies normally are manufactured with a bypass-valve cracking pressure between 25 and 100 psi. The bypass valve in most of these assemblies actually limits the maximum pressure drop across the filter element. As the element becomes blocked with contaminant, the pressure differential increases until it reaches the bypass valve cracking pressure. At this point, part of the flow through the filter assembly begins to bypass the element through the valve. This action limits the maximum pressure differential across the filter element. The relationship between the starting clean pressure differential across the filter element and the bypass valve pressure setting must be considered. A cellulose element has a narrow region of exponential pressure rise. For this reason, the relationship between the starting clean pressure differential and the bypass valve pressure setting is very important. This relationship in effect determines the element’s useful life. In contrast, the useful element life of single-layer and multi-layer fiberglass elements is established by the nearly horizontal, linear region of relatively low pressure drop increase, not the region of exponential pressure rise. Accordingly, the filter assembly’s bypass valve cracking pressure, whether 25 or 75 psi, has relatively little impact on the useful life of the element. Thus, the initial pressure differential and bypass valve setting is less a sizing factor for fiberglass media.
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Filter types and locations The type of filter — suction, return, pressure, or off-line — and its physical location in the circuit are almost inseparable by definition. Suction filters serve to protect the pump from fluid contamination. They are located upstream from the pump’s inlet port. Some may be simple inlet strainers, submersed in fluid in the tank. Others may be mounted externally. In either case, suction filters have relatively coarse elements, due to cavitation limitations of pumps. (Some pump manufactures do not recommend the use of a suction filter. Always consult the pump manufacturer for inlet restrictions.) For this reason, suction filters are not used as a system’s primary protection against contamination, and in fact, the use of suction strainers and filters has greatly decreased in modern hydraulic equipment. Return filters may be the best choice if the pump is particularly sensitive to contamination. In most systems, the return filter is the last component through which fluid passes before entering the reservoir. Therefore, it captures wear debris from all of the system’s working components and any particles that enter through worn cylinder rod seals before such contaminant can enter the reservoir and be pumped back into the system. Because this filter is located immediately upstream from the reservoir, its pressure rating and cost can be relatively low. Note that retracting some cylinders with large diameter rods may result in flow multiplication. This high returnline flow rate may open the filter bypass valve, allowing unfiltered fluid to pass downstream. This probably is an undesirable condition and should be considered when specifying the filter. Pressure filters are located downstream from the system pump. They are designed to handle the system pressure and are sized for the specific flow rate in the pressure line where they are located. Pressure filters are especially suited for protecting sensitive components, such as servovalves, directly downstream from the filter. Because pressure filters are located just downstream from the pump, they also help protect the entire system from any
pump-generated contamination. Duplex filters, a common special configuration, may include both pressure and return filters. Duplex filters provide continuous filtration. They have two or more filter chambers and include the necessary valving to allow for uninterrupted operation. When one filter element needs to be serviced, the duplex valve is shifted, diverting flow to the opposite filter chamber. The dirty element can then be changed, while flow continues to pass through the cleaner element. The duplex valve typically is an open cross-over type, which prevents any flow blockage. Off-line filtration This increasingly popular filtration arrangement — also referred to as recirculating, kidney loop, or auxiliary filtration — is totally independent of a machine’s main hydraulic system. This makes it attractive as a retrofit project for problem systems. An off-line filtration circuit includes its own pump and electric motor, a filter, and the appropriate connecting hardware. These components are installed off-line as a small subsystem separate from the working lines, or they may be included in a fluidcooling loop. Fluid is pumped continuously out of the reservoir, through the off-line filter, and back to the reservoir. A rule of thumb: the off-line pump should be sized to flow a minimum of 10% of the main reservoir volume. With its polishing effect, off-line filtration is able to maintain the system’s fluid at a constant contamination level. As with a return line filter, the off-line loop is best suited to maintain overall system cleanliness; it does not provide protection for specific components. An off-line filtration loop has the added advantage that it is relatively easy to retrofit on an existing system that has inadequate filtration. Also, the off-line filter can be serviced without shutting down the main system. Conclusion Almost every hydraulic system would benefit greatly from having a combination of suction, pressure, return, and off-line filters — enjoying the comprehensive performance advantages of each type.
Hydraulic fluids T
he demands placed on hydraulic systems constantly change as industry requires greater efficiency and speed at higher operating temperatures and pressures. Selecting the best hydraulic fluid requires a basic understanding of each particular fluid’s characteristics in comparison with an ideal fluid. An ideal fluid would have these characteristics: ● thermal stability ● hydrolytic stability ● low chemical corrosiveness ● high anti-wear characteristics ● low tendency to cavitate ● long life ● total water rejection ● constant viscosity, regardless of temperature, and ● low cost. Although no single fluid has all of these ideal characteristics, it is possible to select one that is the best compromise for a particular hydraulic system. This selection requires knowledge of the system in which a hydraulic fluid will be used. The designer should know such basic characteristics of the system as: ● maximum and minimum operating and ambient temperatures ● type of pump or pumps used ● operating pressures ● operating cycle ● loads encountered by various components, and ● type of control and power valves Influential factors Each of the following factors influences hydraulic fluid performance: Wear — Of all hydraulic system problems, wear is most frequently misunderstood because wear and friction usually are considered together. Friction should be considered apart from wear. Wear is the unavoidable result of
metal-to-metal contact. The designer’s goal is to minimize metal breakdown through an additive that protects the metal. By comparison, friction is reduced by preventing metal-to-metal contact through the use of fluids that create a thin protective oil or additive film between moving metal parts. Note that excessive wear may not be the fault of the fluid. It may be caused by poor system design, such as excessive pressure or inadequate cooling. Anti-wear — The compound most frequently added to hydraulic fluid to reduce wear is zinc dithiophosphate (ZDP), but today, ashless anti-wear hydraulic fluids have become popular with some companies and in certain states to reduce loads on waste treatment plants. No ZDP or other type heavy metals have been used in the formulation of ashless anti-wear fluids. The pump is the critical dynamic element in any hydraulic system, and each pump type (vane, gear, piston) has different requirements for wear protection. Vane and gear pumps need antiwear protection. With piston pumps, rust and oxidation (R & O) protection is more important. This is because gear and vane pumps operate with inherent metal-to-metal contact, while pistons ride on an oil film. When two or more types of pumps are used in the same system, it is impractical to have a separate fluid for each, even though their operating requirements differ. The common fluid selected, therefore, must bridge the operating requirements of all pump types. Foaming — When foam is carried by a fluid, it degrades system performance and therefore should be eliminated. Foam usually can be prevented by eliminating air leaks within the system. However, two general types of foam still occur frequently:
● surface foam, which usually collects on the fluid surface in a reservoir, and ● entrained air. Surface foam is the easiest to eliminate, with defoaming additives or by proper sump design so that foam enters the sump and has time to dissipate. Entrained air can cause more serious problems because this foam is drawn into the system. In worst cases, it causes cavitation, a hammering action that can destroy parts. Entrained air is usually prevented by properly selecting the additive and base oils. Caution: certain anti-foam agents, when used at a high concentration to reduce surface foam, will increase entrained air. Also linked to the foam problem, is fluid viscosity, which determines how easily air bubbles can migrate through the fluid and escape. R & O — Most fluids need rust and oxidation inhibitors. These additives both protect the metal and contain antioxidation chemicals that help prolong fluid life. Corrosion — Two potential corrosion problems must be considered: system rusting and acidic chemical corrosion. System rusting occurs when water carried by the fluid attacks ferrous metal parts. Most hydraulic fluids contain rust inhibitors to protect against system rusting. The tests used to measure this capability are ASTM D 665 A and B. To protect against chemical corrosion, other additives must be considered. The additives must also exhibit good stability in the presence of water (hydrolytic stability) to prevent break down and acidic attack on system metals. Oxidation and thermal stability — Over time, fluids oxidize and form acids, sludge, and varnish. Acids can attack system parts, particularly soft metals. Extended high-temperature operation and thermal cycling also en-
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courage the formation of fluid decomposition products. The system should be designed to minimize these thermal problems, and the fluid should have additives that exhibit good thermal stability, inhibit oxidation, and neutralize acids as they form. Although not always practical or easy to attain, constant moderate temperature and steady-state operation are best for system and fluid life. Water-retention — Large quantities of water in a hydraulic fluid system can be removed by draining the sump periodically. However, small amounts of water can become entrained, particularly if the sump is small. Usually, demulsifiers are added to the fluid to speed the separation of water. Filters can then physically remove any remaining water from the hydraulic fluid. The water should leave the fluid without taking fluid or additives with it. Viscosity — Maximum and minimum operating temperatures, along with the system’s load, determine the fluid’s viscosity requirements. The fluid must maintain a minimum viscosity at the highest operating temperature. However, the hydraulic fluid must not be so viscous at low temperature
that it cannot be pumped. Temperature — System operating temperature varies with job requirements. Here are a few general rules: the maximum recommended operating temperature usually is 150° F. Operating temperatures of 180° to 200°F are practical, but the fluid will have to be changed two to three times as often. Systems can operate at temperatures as high as 250° F, but the penalty is fairly rapid decomposition of the fluid and especially rapid decomposition of the additives — sometimes within 24 hours! Fluid makeup Most fluids are evaluated based on their ratings for rust and oxidation (R & O), thermal stability, and wear protection, plus other characteristics that must be considered for efficient operation: Seal compatibility — In most systems, seals are selected so that when they encounter the fluid they will not change size or they will expand only slightly, thus ensuring tight fits. The fluid selected should be checked to be sure that the fluid and seal materials are compatible, so the fluid will not interfere with proper seal operation. Fluid life, disposability — There
are two other important considerations that do not directly relate to fluid performance in the hydraulic system but have a great influence on total cost. They are fluid life and disposability. Fluids that have long operating lives bring added savings through reduced maintenance and replacement-fluid costs. The cost of changing a fluid can be substantial in a large system. Part life should also be longer with the higher-quality, longer-lived fluid. Longer fluid life also reduces disposal problems. With greater demands to keep the environment clean, and ever-changing definitions of what is toxic, the problem of fluid disposability increases. Fluids and local anti-pollution laws should both be evaluated to determine any potential problems. Synthesized hydrocarbon (synthetic) hydraulic fluids have been introduced recently. They contain no waxes that congeal at low temperatures nor compounds that readily oxidize at high temperatures which are inevitable in natural mineral oils. The synthesized hydrocarbon hydraulic fluids are being used for applications with very low, very high, or a very wide range of temperatures.
Fire-resistant fluids
T
he overwhelming majority of hydraulic components and systems are designed to use oil-based hydraulic fluids. No wonder; these fluids rarely present significant operating, safety, or maintenance problems. Unfortunately, there are circumstances where using oil-based fluid should be avoided. One common fluid power application is in an environment with potential ignition sources — an open flame, sparks, or hot metal. In these environments, a leak spraying from a high-pressure hydraulic system could cause a serious fire and result in major property damage, personnel injury, or even death. Even though most oil-based hydraulic fluids have relatively high flash/fire points (.3008 F), small leaks in a high-pressure system can produce a finely atomized spray that can travel significant distances. If an ignition source is encountered, complete ignition of the spray envelope can occur. The alternative is to use a hydraulic
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fluid that eliminates or significantly reduces this hazard: any of several fireresistant hydraulic fluids (FRHFs). How far we’ve come Apart from isolated segments of basic research, little progress was made in developing suitable FRHFs until the end of World War II. During the war, tragic incidents related to hydraulic fluid fires and major property losses at steel mills and foundries graphically illustrated the urgent need for something to be done. Similar incidents in captive environments such as coal mines during the rapid post-war industrial expansion helped motivate a major joint research effort between government and industry. This work was directed at developing fluids that could replace oilbased hydraulic fluids at a reasonable cost and with no significant reduction in hydraulic system performance. Two basic approaches were undertaken. One involved the introduction of water
into the fluid to act as a “snuffer” if the fluid ignited. The other involved synthetic, non-aqueous products whose chemistry resisted burning or generated products of combustion that helped extinguish any flame. Commercial products in both categories evolved during the 1950s and ‘60s and are still in use today. In the early 1970s, an additional synthetic type of fluid was introduced to address many of the drawbacks inherent in the earlier types. Since the introduction of each type, many improvements have been made in fire resistance, anti-wear properties, and overall quality. Where we are Water glycol and invert emulsion constitute the major fluid types of water-containing products. Water glycol is a true solution of a glycol (such as ethylene glycol) in water, along with a variety of additives to impart viscosity, corrosion protection, and anti-wear
HYDRAULIC FLUIDS
properties. A shear-stable thickener, which has improved over the years, represents the novel technology aspect of the fluid. Water glycol contains approximately 40% water. Despite a number of drawbacks, water glycol is the dominant FRHF on the market today and is used in a wide variety of applications. An invert emulsion also contains approximately 40% water but is a stable emulsion of water dispersed in oil. The outer phase, oil, represents the wetting surface; the inner phase, water, provides the fire retardant-element. Oilsoluble additives provide anti-wear properties, corrosion protection, and emulsion stability. Inverts, at one time, were commonly used but are losing favor in industry today. Synthetic fluids initially were represented by a class of chemical compounds known as phosphate esters, which are reaction products between phosphoric acid and aromatic ringstructure alcohols. These fluids are extremely fire resistant and have widespread industrial use, as well as military and aircraft service. However, their popularity has declined because of environmental, cost, and compatibility factors. The other type of synthetic fluids in use are synthetic hydrocarbons, more specifically, polyol esters. These fluids are the reaction products between longchain fatty acids (derived from animal and vegetable fats) and synthesized organic alcohols. These products contain additives to impart anti-wear properties, corrosion protection, and viscosity modification. Fire resistance results from a combination of high thermal properties and physical characteristics. This is the most recent category of FRHFs and has gained widespread and growing use. What is fire resistance? The term “fire resistant” often is misunderstood or interpreted to be overly inclusive; it seems appropriate to standardize the terminology and review the accepted test methods for judging the fire resistance of a given fluid. First, there is no single property or test of a fluid, such as flash/fire point, auto ignition temperature (AIT), etc. that will quantitatively rate its relative fire resistance. This has led to a “simulated incident” approach in
which tests are designed to replicate a worst-case scenario in typical applications where fluid power is used near a potential fire hazard. Fluids generally pass or fail these tests, and those that pass are incorporated into an Approval Guide or List of Qualified Fluids. In the United States, two test protocols have evolved and are generally regarded as benchmarks in the industry. One was developed by Factory Mutual Research Corporation (FMRC). Their original intent was to use the test results in the risk-assessment programs of those insurance companies under the Factory Mutual System umbrella. The test has since become the chief qualification for commercial companies using FRHFs; all fluid suppliers submit products seeking “FMRC Approval.” The 1992 FMRC Approval Guide lists over 300 FRHFs from approximately 50 suppliers. Factory Mutual‘s program is now global in scope. FMRC addresses the definition of FRHF in the following excerpt in their introduction to the hydraulic fluids sections of their Approval Guide: Less flammable hydraulic fluids approved and listed here have been tested to evaluate fire hazard only. All presently available fluids will burn under certain conditions. In each case the fire hazard has been reduced to an acceptable degree, meeting the Approval Standards of FMRC; other fluid properties are not investigated. This paragraph accurately puts the intent of FRHFs into the proper perspective. They are not fireproof but, rather, they significantly reduce the potential hazard associated with oil-based products. In the FMRC tests, the fluid is conditioned to 1408F, pressurized to 1000 psi in a steel cylinder, and discharged through an oil burner-type nozzle. The spray generated is intended to simulate a high-pressure hydraulic system leak. A gas flame is passed through (not retained in) the spray envelope at two distances downstream of the nozzle. There may be local burning at the point of flame entry, and the pass criteria dictate that any flame must self-extinguish when the ignition source is removed; no flame may propagate back to the nozzle. This process is repeated 20 times, and the burn duration timed. Any burn duration over 5 sec is considered a fail.
A second test uses the same spray directed at an inclined metal channel heated to 13008F. In this test, the spray is continuous for 60 sec. The criteria are: 1. The spray in contact with the channel may not burn, or 2 If spray ignition takes place, fluid rolling off the channel cannot continue to burn, and the flame cannot follow the spray if directed away from the channel. If these conditions are satisfied, the fluid is approved. Statistics are not available, but many products in all of the fluid categories described do not pass this test. The Mine Safety & Health Administration (MSHA) has had in place for many years an evaluation program for qualifying fluids that are used underground, primarily in coal mines. MSHA testing is similar to FMRC’s in the sense that a spray mist of the candidate fluid is generated. However, the ignition mechanism is somewhat different in the MSHA test. Under this procedure, a spray mist is directed continuously at a variety of ignition sources that include an open gas flame, a welding arc, and burning rags. The pass criteria are that localized burning in the spray mist extinguish within 5 sec, and there can be no sustained propagation along the spray axis. They also have an AIT criterion and a wick test to assess the rate of evaporation of water from a candidate product. MSHA tests also have a relatively high rate of product rejections. Since both of these tests involve fluids submitted by the supplier to the testing agency, both FMRC and MSHA have comprehensive manufacturer auditing programs in which quality-assurance programs are carefully evaluated and monitored by periodic, on-site inspections. This may include retests of approved fluids. Other tests In addition to these “third party” ratings of FRHFs, many companies have developed their own fire-resistance tests that must be considered in addition to a product having FMRC approval. Again, these tests generally follow the simulated incident philosophy and are specific to the type of industry involved. Examples of these in-
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clude exposing the candidate fluid — in spray or non-spray form — to a hot manifold, molten metal, heated blocks of a representative metal, burning rags, hot sand, etc. The evaluation criteria may be no burning, limited burning, no smoke, non-propagation, etc. Minimum AIT and flash/fire point temperatures also are used either independently or in combination with a test described above. In all of these tests, a product is either approved or rejected; there is no ranking or rating of approved products. This aspect, the occasional lack of reproducibility, and the absence of service history of a fluid has led FMRC to develop a new test that will quantify the relative fire resistance of various fluids. The test procedure involves measuring the heat release of a fluid under a fixedburn condition and combining this value with a separately determined measurement of the energy required to initiate burning. These values are used to establish a Spray Flammability Pa-
rameter for each product evaluated. This test and a new approval standard currently are under review by FMRC and have not been formally adopted.
Many suppliers offer products for each type of FRHF, which may vary considerably in price, quality, and after-sale service.
Other concerns The major problem facing a designer converting a hydraulic system from an oil-based fluid to FRHF is selecting the particular type that will minimize the cost of conversion and maximize the operating and safety benefits. The choice becomes a trade-off of characteristics associated with each type. Each product group offers advantages and disadvantages for any given application. It is beyond the scope of this article to attempt to make recommendations for certain end-users, but the major attributes and shortfalls of the various fluid types can be addressed. The table on pages 38 and 39 summarizes these characteristics, price ranges, and some of the considerations associated with converting a system containing oil-based fluid to a FRHF.
Where we're going With regard to future developments, there are no new products on the horizon which will make this process any easier. Significant improvements have been made in recent years with both water glycol and polyol ester fluids, and this trend should continue. Moreover, the impact of more-stringent environmental regulations will be more strongly felt in the next few years and may even restrict the choice. The motivation for converting from an oil-based fluid will also strengthen as waste control regulations expand for any product containing oil. In some areas, “hydraulic oil” already is considered a hazardous material. Fluids having the capability of being non-toxic and readily biodegradable will further expand the need to replace oil-based hydraulic fluids.
Environmentally safe and fire-proof
A
drawback of most hydraulic fluids, including some fire-resistant fluids, is their toxicity — either to personnel, the environment, or both. Furthermore, they are only fire resistant, and most will burn under certain conditions. Recently introduced synthetic water additives, on the other hand, mix with water (usually in a concentration of 5%) to become fire proof; the solution actually could extinguish a fire. These water-based fluids, in general, also offer a cost advantage over most other fluids because one gallon of concentrate produces 20 gallons of hydraulic fluid. When disposal expenses enter calculations, the cost differential becomes even greater — especially with a solution containing non-toxic, readily biodegradable synthetic water additives that require no treatment. The accompanying table summarizes characteristics of common fire-resistant and fire-proof fluids.
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There are, however, important performance and operating characteristics of water-based fluids that cannot be ignored. First, water-based fluids in general have much lower viscosity, film strength, and lubricating qualities than oil-based fluids. This means that system components — especially pumps, valves, and actuators — must be designed specifically for operation with water-based fluid. You can’t just drain fluid from a system containing oilbased fluid and expect it to run on water-based fluid. A perception remains today that components for water-based fluid are much more expensive and larger — especially valves — than their conventional counterparts. While this may have been true 20 years ago, the cost premium for valves and other components designed for waterbased fluid has narrowed to about 30%. This investment can easily be recovered in the cost of fluid alone, not to mention disposal and treatment costs. Moreover, valve size has been
reduced dramatically: many are available with standard NFPA footprints. Next, any potential for freezing must be considered. Traditionally, ethylene glycol is added to water to lower the solution’s freezing point. However, using highly toxic ethylene glycol in a solution containing the synthetic additive would completely negate the purpose of using an environmentally safe additive. Using propylene glycol instead as anti-freeze maintains the environmental integrity of the solution because propylene glycol is so non-toxic that it is approved for use in food by the U. S. Food & Drug Administration. Finally, because the fluid is nontoxic, it naturally tends to support microbial growth. To minimize or prevent consequences associated with this problem, judicious use of bacteriostatic additives and effective sealing and reservoir design should be practiced.
HYDRAULIC FLUIDS
Environmental fluids
I
n some cases, environmental considerations necessitate the selection of a zinc-free ashless petroleum or a biodegradable hydraulic fluid. The base fluids of biodegradable hydraulic fluids are usually vegetable oils, selected synthetic esters, or a blend of the two. Biodegradable hydraulic fluids typically contain low toxic, ashless inhibitors and additives to enhance performance. Properly formulated biodegradable hydraulic fluids can provide effective wear resistance similar to petroleum anti-wear hydraulic fluids. However, the biodegradable fluids may be susceptible to water contamination and may exhibit poor oxidative stability, especially when the base fluids are vegetable oils. The use of a synthetic-ester base usually improves the water tolerance and oxidation resistance of the fluids. The tradeoff between environmental advantages and potential performance deficiencies of biodegradable hydraulic fluids suggests that these fluids are most suitable for applications in environmentally sensitive areas, and that they are not meant as a direct replacement for petroleum hydraulic fluids. Their use should be considered in outdoor equipment, such as in timber harvesting, agricultural equipment, airport service fleets, construction machinery, recreational resorts, or wherever contamination of ground or water by petroleum lubricants could be a problem. Vegetable oil-based fluids may be considered when operating temperatures range from 0° to 180° F. For operations in sub-zero temperatures or temperatures higher than 180° F, synthetic ester-based fluids are preferred. Additives Like petroleum oils, vegetable oils or synthetic esters rely on specially selected additives to improve their performance as lubricants. The additives contained in biodegradable hydraulic fluids typically exhibit very low toxicity. Unlike petroleum oils, vegetable oils contain unsaturated hydrocarbons and are natural occurring esters. The unsaturation leads to rapid oxidation at elevated temperatures and poor low temperature flow properties. This low-temperature
fluidity can be improved by additives, but their oxidation stability remains a performance concern. International guidelines Throughout Europe, the development of guidelines for biodegradable lubricants is typically left to local authorities or non-government organizations. In Germany, Blue Angel labels will be awarded to biodegradable hydraulic fluids. The Blue Angel for biodegradable hydraulic fluids will likely require that the base fluids must be readily biodegradable — greater than 80% biodegradation in 21 days by CEC L-33-A93 Test, or greater than 70% biodegradation in 28 days by the Modified Sturm Test. In addition, all components must be Water Hazard Class 0 or 1, which means the components are not water pollutants. Environmental Choice Program of Canada is currently in the process of reviewing a guideline on biodegradable, non-toxic hydraulic fluids. It will likely include a requirement that base fluids exhibit greater than 90% biodegradation in 21 days by CEC L-33-A93. In the United States, ASTM D2.N.3 on eco-evaluated hydraulic fluids has drafted an information guide that addresses the means of assessing the biodegradability of hydraulic fluids. D-2.N.3 is currently developing environmental classifications for hydraulic fluids. A study group of D2.N.3 is making progress in developing the performance classification and specifications for biodegradable hydraulic fluids. In December 1995, ASTM D-2.12 on Environmental Standards for Lubricants completed a standard test especially designed to determine the aerobic aquatic biodegradability of all lubricants and their components. The test is similar to the Modified Sturm Test, which measures the evolution of carbon dioxide in 28 days. This standard is being published as ASTM D 5864. ASTM D-2.12 is currently developing other environmental standard tests for lubricants, which include an aquatic toxicity test for fish and large invertebrates; a manometric respirometry biodegradation test method; and a Gladhill Shake Flask biodegradation test.
Initially designed to measure the biodegradability of 2-cycle engine oils, CEC L-33-A93 has been the most widely applied biodegradation test for lubricants in Europe since the early 1980s. The test uses infrared spectroscopy to measure the disappearance of certain hydrocarbons over a 21-day period when the lubricant is mixed with an inoculum containing micro-organisms. Thus, the CEC test is a only a measure of primary biodegradation. Unlike the CEC test, the Modified Sturm Test is a measure of ultimate biodegradation. By measuring the production of CO2 over 28 days, the test estimates the extent to which the carbon in a lubricant is converted by micro-organisms to the elements found in nature — namely: CO 2, water, inorganic compounds, and biotic mass. Because this test was designed originally for watersoluble, pure compounds, it is difficult to use for testing lubricants, most of which are water-insoluble, complex mixtures. The new ASTM D 5864 test is similar to the Modified Sturm Test. It is specially designed for testing water insoluble complex lubricants. The readily biodegradable question One question that often comes up is whether a fluid is readily biodegradable or just biodegradable. Most things are biodegradable, given enough time and proper conditions. Readily biodegradable means that a substance exhibits a result equal to or greater than a pre-set requirement in a standard test. For example, XYZ Standard requires 80% or higher biodegradation by CEC L-33-A93 in 21 days. If a lubricant meets this requirement, it is considered readily biodegradable by the XYZ Standard. Ideally, any claim that a lubricant is readily biodegradable also also specify the test and standard. Vegetable oils or synthetic esters? Being natural occurring esters, vegetable oils are susceptible to hydrolysis, which leads to fluid decomposition and degradation, especially in the presence of heat. Because of their polarity, vegetable oils tend to cause elastomers to swell, though in most cases the degree of swell is insufficient to cause any seri-
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ous concern in hydraulic applications. On the other hand, vegetable oils offer excellent lubricity, intrinsic high viscosity index, and good anti-wear and extreme-pressure properties. Well-formulated, biodegradable hydraulic fluids based on vegetable oils can easily pass the demanding Vickers 35VQ25 or Denison T5D-42 vane-pump wear tests. They also can meet the requirements of major OEMs for premium hydraulic fluids, except hydrolytic, thermal, and oxidation stability. Experience has shown that vegetable oil-based biodegradable hydraulic fluids can perform satisfactory for years under mild climate and operation conditions (temperatures below 160° F, and hydraulic systems kept free of water contamination). The use of synthetic esters — typically polyol esters — provides better hydrolytic, thermal, and oxidative stability, and excellent low-temperature fluidity, while preserving the high biodegradability and low toxicity of the fluids. For nearly 30 years, polyol esters have been used to formulate aviation gas turbine lubricants, which demand high thermal and oxidation stability at extreme temperatures. While a vegetable
oil-based hydraulic fluid can perform between 0° to 180° F, a similar fluid based on synthetic esters can be used between –25° and 200° F. Similar to vegetable oils, synthetic esters have the tendency to swell and soften elastomers, although again, the swell should not be a concern for most hydraulic applications. Fluid handling Vegetable oil or synthetic esterbased biodegradable hydraulic fluids are fully miscible with each other and with petroleum hydraulic fluids. However, when a biodegradable hydraulic fluid is mixed with petroleum lubricants, its biodegradability typically decreases, and its toxicity increases. Because of their susceptibility to hydrolysis, vegetable oil- or synthetic ester-based fluids should be kept free of water contamination, both in storage and in everyday use. There is no regulation permitting shortcuts in the disposal of biodegradable hydraulic fluids. Such disposal should be handled in the same manner as the disposal of petroleum fluids, in accordance with applicable federal, state, and local laws and regulations.
The future of biodegradable fluids Government regulations and codes, and the environmental awareness of lubricant users are the driving forces for the growing use of biodegradable hydraulic fluids. However, the lack of definition and standards for biodegradable fluids in the United States impedes the market development for these fluids. Development of new standards and guidelines by ASTM and other industrial and governmental organizations will inevitably influence the growth of biodegradable fluids. Meanwhile, lubricant suppliers continue to develop and evaluate new additive chemistries that provide greater oxidative, thermal, and hydrolytic stability properties for biodegradable fluids. Vegetable oil suppliers are using genetic engineering to produce new vegetable oils with improved stability. Ester manufacturers are considering improving ester performance by incorporating additive-type functional groups into molecular structures. The improvement in the performance quality of biodegradable hydraulic fluids will eventually lead to more applications and increased popularity of these important fluids.
Glossary of hydraulic-fluid terminology Absolute viscosity — the ratio of shear stress to shear rate. It is a fluid’s internal resistance to flow. The common unit of absolute viscosity is the poise. Absolute viscosity divided by fluid density equals kinematic viscosity. Absorption — the assimilation of one material into another. Additive — chemical substance added to a fluid to impart or improve certain properties. Common petroleum product additives are: antifoam agent, anti-icing additive, antiwear additive, corrosion inhibitor, demulsifier, detergent, dispersant, emulsifier, EP additive, oiliness agent, oxidation inhibitor, pour point depressant, rust inhibitor, tackiness agent, viscosity index improver. Adsorption — adhesion of the molecules of gases, liquids, or dissolved substances to a solid surface, resulting in relatively high concentration of the molecules at the place of contact; e.g. the plating out of an anti-wear additive on metal surfaces.
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Anti-foam agent — one of two types of additives used to reduce foaming in petroleum products: silicone oil to break up large surface bubbles, and various kinds of polymers that decrease the amount of small bubbles entrained in the oils. Asperities — microscopic projections on metal surfaces resulting from normal surface-finishing processes. Interference between opposing asperities in sliding or rolling applications is a source of friction, and can lead to metal welding and scoring. Ideally, the lubricating film between two moving surfaces should be thicker than the combined height of the opposing asperities. Bactericide — additive included in the formulations of water-mixed fluids to inhibit the growth of bacteria. Boundary lubrication — form of lubrication between two rubbing surfaces without development of a full-fluid lubricating film. Boundary lubrication can be made more effective by includ-
ing additives in the lubricating oil that provide a stronger oil film, thus preventing excessive friction and possible scoring. There are varying degrees of boundary lubrication, depending on the severity of service. Bulk modulus — the measure of a fluid’s resistance to compressibility; the reciprocal of compressibility. Cavitation — formation of an air or vapor pocket (or bubble) due to lowering of pressure in a liquid, often as a result of a solid body, such as a propeller or piston, moving through the liquid; also, the pitting or wearing away of a solid surface as a result of low fluid levels that draw air into the system, producing tiny bubbles that expand explosively at the pump outlet, causing metal erosion and eventual pump destruction. Corrosion inhibitor — additive for protection of wetted metal surfaces from chemical attack by water or other contaminants. Polar compounds wet the metal surface preferentially,
HYDRAULIC FLUIDS
protecting it with a film of oil. Other compounds may absorb water by incorporating it in a water-in-oil emulsion so that only the oil touches the metal surface. Another type of corrosion inhibitor combines chemically with the metal to present a non-reactive surface. Demulsibility — ability of an oil to separate from water. Dewaxing — removal of paraffin wax from lubricating oils to improve low temperature properties, especially to lower the cloud point and pour point. Emulsifier — additive that promotes the formation of a stable mixture, or emulsion, of oil and water. Common emulsifiers are: metallic soaps, certain animal and vegetable oils, and various polar compounds. Emulsion — intimate mixture of oil and water, generally of a milky or cloudy appearance. Emulsions may be of two types: oil-in water (where water is the continuous phase) or water-in-oil (where water is the discontinuous phase). EP additive — lubricant additive that prevents sliding metal surfaces from seizing under conditions of extreme pressure (EP). At the high local temperatures associated with metal-tometal contact, an EP additive combines chemically with the metal to form a surface film that prevents the welding of opposing asperities, and the consequent scoring that is destructive to sliding surfaces under high loads. Reactive compounds of sulfur, chlorine, or phosphorus are used to form these inorganic films. Fire-resistant fluid — hydraulic oil used especially in high-temperature or hazardous applications. Three common types of fire-resistant fluids are: (1) water-petroleum oil emulsions, in which the water prevents burning of the petroleum constituent; (2) water-glycol fluids; and (3) non-aqueous fluids of low volatility, such as phosphate esters, silicones, polyolesters, and halogenated hydrocarbon-type fluids. Full-fluid-film lubrication — presence of a continuous lubricating film sufficient to completely separate two surfaces, as distinct from boundary lubrication. Full-fluid-film lubrication is normally hydrodynamic lubrication, whereby the oil adheres to the moving
part and is drawn into the area between the sliding surfaces, where it forms a pressure, or hydrodynamic wedge. Hydraulic fluid — fluid serving as the power transmission medium in a hydraulic system. The most commonly used fluids are petroleum and synthetic oils, oil-water emulsions, and water glycol mixtures. The principal requirements of a premium hydraulic fluid are proper viscosity, high viscosity index, anti-wear protection (if needed), good oxidation stability, adequate pour point, good demulsibility, rust inhibition, resistance to foaming, and compatibility with seal materials. Antiwear oils are frequently used in compact, high-pressure, and high-capacity pumps that require extra lubrication protection. Immiscible — incapable of being mixed without separation of phases. Water and petroleum oil are immiscible under most conditions, although they can be made miscible with the addition of a proper emulsifier. Inhibitor — additive that improves the performance of a petroleum product through the control of undesirable chemical reactions. Kinematic viscosity — absolute viscosity of a fluid divided by its density at the same temperature of measurement. It is the measure of a fluid’s resistance to flow under gravity. Lubricity — ability of an oil or grease to lubricate (also called film strength). Miscible — capable of being mixed in any concentration without separation of phases; e.g., water and ethyl alcohol are miscible. Newtonian fluid — fluid, such as a straight mineral oil, whose viscosity does not change with rate of flow. Non-Newtonian fluid — fluid, such as a grease or a polymer containing oil (e.g. multi-grade oil), in which shear stress is not proportional to shear rate. Oxidation inhibitor — substance added in small quantities to petroleum product to increase its oxidation resistance, thereby lengthening its service or storage life; also called anti-oxidant. Polar compound — a chemical compound whose molecules exhibit electrically positive characteristics at one extremity and negative characteristics at the other. Polar compounds are used as additives in many petroleum products.
Pour point — lowest temperature at which an oil or distillate fuel is observed to flow, when cooled under conditions prescribed by test method ASTM D 97. The pour point is 3°C (5° F) above the temperature at which the oil in a test vessel shows no movement when the container is held horizontally for five seconds. Shear rate — rate at which adjacent layers of fluid move with respect to each other, usually expressed as reciprocal seconds. Shear stress — frictional force overcome in sliding one layer of fluid along another, as in any fluid flow. The shear stress of a petroleum oil or other Newtonian fluid at a given temperature varies directly with shear rate (velocity). The ratio between shear stress and shear rate is constant; this ratio is termed viscosity. Surfactant — surface-active agent that reduces interfacial tension of a liquid. A surfactant used in a petroleum oil may increase the oil’s affinity for metals and other material. Vapor pressure — pressure of a confined vapor in equilibrium with its liquid at a specified temperature; thus, a measure of a liquid’s volatility. Viscosity — measurement of a fluid’s resistance to flow. The common metric unit of absolute viscosity is the poise, which is defined as the force in dynes required to move a surface one square centimeter in area past a parallel surface at a speed of one centimeter per second, with the surfaces separated by a fluid film one centimeter thick. In addition to kinematic viscosity, there are other methods for determining viscosity, including, Saybolt Universal viscosity, Saybolt Furol viscosity, Engier viscosity, and Redwood viscosity. Since viscosity varies inversely with temperature, its value is meaningless until the temperature at which it is determined is reported. Viscosity index (V.I.) — empirical, unitless number indicating the effect of temperature changes on the kinematic viscosity of an oil. Liquids change viscosity with temperature, becoming less viscous when heated; the higher the V.I. of an oil, the lower its tendency to change viscosity with temperature.
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Hydraulic hose T
o say that hose is an important part of a hydraulic system is a huge understatement. Hydraulic hose provides a basic means for transporting fluid from one component to another, and at the same time it supplies an inherent versatility to designers. The flexibility of hose enables components to be positioned in the most efficient or convenient places, because the hose has the ability to bend around corners, through tight spaces, or across long distances. Yet these days, there seems to be as many different types of hose as there are telephone long-distance carriers. How does a designer tell one from the other? Isn’t there an easy way to choose or compare hoses? The SAE standards SAE answers those questions with its J517 hydraulic hose standard. This hose standard serves as the most popular benchmark in the realm of industrial hydraulics today. More specifically, J517 is a set of guidelines that applies to the current SAE 100R series of hoses. Currently, 16 such hose styles exist, and they are designated as 100R1 through 100R16 (see descriptions, pages A105 and 106). Each of the styles must meet a set of dimensional and performance characteristics as set forth by SAE. However, SAE issues no approval source lists, certification, or letters of approval—conformance to these standards by manufacturers is strictly voluntary. In short, the standards only assure a similarity of products among different manufacturers. Hydraulic hose construction Modern hydraulic hose typically consists of at least three parts: an inner tube that carries the fluid, a reinforcement layer, and a protective outer layer.
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The inner tube must have some flexibility and needs to be compatible with the type of fluid it will carry. Commonly used compounds include synthetic rubber, thermoplastics, and PTFE, sometimes called Teflon. The reinforcement layer consists of one or more sheaths of braided wire, spiralwound wire, or textile yarn. The outer layer is often weather-, oil-, or abrasionresistant, depending upon the type of environment the hose is designed for. Not surprisingly, hydraulic hoses have a finite life. Proper sizing and use of the correct type of hose will certainly extend the life of a hose assembly, but there are many different factors that affect a hose’s lifespan. SAE identifies some of the worst offenses as: ● flexing the hose to less than the specified minimum bend radius ● twisting, pulling, kinking, crushing, or abrading the hose ● operating the hydraulic system above maximum or below minimum temperature
● exposing the hose to rapid or transient rises (surges) in pressure above the maximum operating pressure, and ● intermixing hose, fittings, or assembly equipment not recommended as compatible by the manufacturer or not following the manufacturer’s instructions for fabricating hose assemblies. Selecting the proper hose Here are seven recommended steps the system designer should follow during the hose and coupling selection process. To help determine the proper hose for an application, use the acronym STAMPED — from Size, Temperature, Application, Materials, Pressure, Ends, and Delivery. Here is what to consider in each area: Size — In order to select the proper hose size for replacement, it is important to measure the inside and outside hose diameters exactly using a precision-engineered caliper, as well as the length of the hose. Hose OD is particularly important when hose-support
Type of fluid
Pressure range
SAE# Petroleum Synthetic High-water Temp.* oil oil content 100R1 100R2 100R3 100R4 100R5 100R6 100R7 100R8 100R9 100R10 100R11 100R12 100R13 100R14 100R15 100R16
x x x x x x x x x x x x x x x x
*Temperatures... 1 = -40° to 100° C
x x
x
x x x x x x x x x x x x x x x
2 = -40° to 93° C
1 1 1 1 1 1 2 2 1 1 1 3 3 4 3 1
psi 3000 5000 1500 300 3000 500 3000 5000 4500 10,000 12,500 4000 5000 1500 6000 5000
3 = -40° to 121° C
ID, in.
psi
ID, in.
3/16 3/16 3/16 3/4 3/16 3/16 3/16 3/16 3/8, 3/16 3/16 3/8 3/4 1/8 3/8 1/4
375 1000 375 35 200 300 1000 2000 2000 2500 2500 2500 5000 600 6000 1625
2 2 1/2 1 1/4 4 3 3/4 1 1 2 2 2 1/2 2 2 1 1/8 1 1/2 1 1/4
4 = -54° to 204° C
HYDRAULIC HOSE
clamps are used or when hoses are routed through bulkheads. Check individual hose specification tables for ODs in suppliers’ catalogs. When replacing a hose assembly, always cut the new hose the same length as the one being removed. Moving components of the equipment may pinch or even sever too long a hose. If the replacement hose is too short, pressure may cause the hose to contract and be stretched, leading to reduced service life. Changes in hose length when pressurized range between +2% to –4% while hydraulic mechanisms are in operation. Allow for possible shortening of the hose during operation by making the hose lengths slightly longer than the actual distance between the two connections. Temperature —All hoses are rated with a maximum working temperature ranging from 200° to 300° F based on the fluid temperature. Exposure to continuous high temperatures can lead to hoses losing their flexibility. Failure to use hydraulic oil with the proper viscosity to hold up under high temperatures can accelerate this problem. Always follow the hose manufacturer’s recommendations. Exceeding these temperature recommendations can reduce hose life by as much as 80%. Depending on materials used, acceptable temperatures may range from –65° F (Hytrel and winterized rubber compounds) to 400° F (PTFE). External temperatures become a factor when hoses are exposed to a turbo manifold or some other heat source. When hoses are exposed to high external and internal temperatures concurrently, there will be a considerable reduction in hose service life. Insulating sleeves can help protect hose from hot equipment parts and other high temperature sources that are potentially hazardous. In these situations, an additional barrier is usually required to shield hydraulic fluid from a potential source of ignition. Application — Will the selected hose meet bend radius requirements? This refers to the minimum bend radius (usually in inches) that a hydraulic hose must meet. Exceeding this bend radius (using a radius smaller than recommended) is likely to injure the hose reinforcement and reduce hose life. Route high-pressure hydraulic lines
parallel to machine contours whenever possible. This practice can help save money by reducing line lengths and minimizing the number of hard-angle, flow-restricting bends. Such routing also can protect lines from external damage and promote easier servicing. Materials — It is mandatory to consult a compatibility chart to check that the tube compound is compatible with the fluid used in the system. Elevated temperature, fluid contamination, and concentration will affect the chemical compatibility of the tube and fluid. Most hydraulic hoses are compatible with petroleum-based oils. Note that new readily biodegradeable or green fluids may present a problem for some hoses. Pressure capabilities — Hose working pressure must always be chosen so that it is greater than or equal to the maximum system pressure, including pressure spikes. Pressure spikes greater than the published working pressure will significantly shorten hose life. Hose ends — The coupling-to-hose mechanical interface must be compatible with the hose selected. The proper mating thread end must be chosen so that connection of the mating components will result in leak-free sealing. There are two general categories of couplings to connect most types of hose: the permanent type (used primarily by equipment manufacturers, largescale rebuilders, and maintenance shops) and the field-attachable type. Permanently attached couplings are cold-formed onto the hose with powered machinery. They are available for most rubber and thermoplastic hoses and offer a wide range of dependable connections at low cost. Assemblies made in the field with portable machines are relatively simple; these machines are economical and easy to operate. In most cases, it is not necessary to skive the cover. These couplings are less complicated to install than other types. Field-attachable couplings are classified as screw-together and clamp-type. The screw-together coupling attaches to the hose by turning the outer coupling shell over the outside diameter of the hose. The coupling insert is then screwed into the coupling shell. A clamp-type coupling has a 2-piece outer shell that clamps onto the hose OD with either two or four bolts and nuts.
In either case, the coupling has limited potential for reuse because the threads distort during attachment. To ensure the correct-size coupling is used when replacing an assembly, the number of threads per inch and thread diameter of the original coupling must be determined. Thread pitch gages are available for identifying the number of threads per inch. A caliper can measure both inside and outside dimensions of the threads. ODs are measured on male couplings, while IDs are measured on female couplings. In most situations, the only differences between an SAE coupling and an imported coupling are the thread configuration and the seat angle. International thread ends can be metric, measured in mm, but also include BSP (British Standard Pipe) threads, which are measured in inches. Knowing the country of origin provides a clue as to what type of thread end is used. DIN (Deutsche Industrial Norme) fittings began in Germany and now are found throughout Europe, while BSP is found on British equipment. Japanese Komatsu machinery uses Komatsu fittings with metric threads, while other Japanese equipment most likely uses JIS (Japanese Industrial Standard–BSP threads), or, in some cases, BSP with straight or tapered threads. Three determinations are required to identify these couplings correctly: • type of seat — inverted (BSPP & DIN), regular (JIS & Komatsu) or flat (flange, flat-face) • seat angle — 30° (JIS, BSP, DIN and Komatsu) or 12° (DIN), and • type of threads — metric (DIN or Komatsu), BSP (BSPP, BSPT or JIS), or tapered (BSPT or JIS Tapered) SAE standards relating to hydraulic/pneumatic fittings and assemblies specifically designed to eliminate leakage include: • J514 — straight thread ports/fittings • J518c — 4-bolt flange ports/fittings, and • XJ1453 — the number provisionally assigned to O-ring face seal fittings. Delivery — How available is the product? Is it unique? How soon can it be delivered to the distributor or end user? It may be preferable to consider several options to maximize flexibility and avoid the delays that can result from relying on components that are
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HYDRAULIC HOSE
SAE hose standards — description and construction SAE 100R1
semble with fittings which do not require removal of the cover or any portion of it.
SAE 100R6
SAE 100R3 This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. Type A—Consists of an inner tube of oil-resistant synthetic rubber, a single wire braid reinforcement, and an oiland weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube or over the wire reinforcement (or both) to anchor the synthetic rubber to the wire. Type AT—Same construction as Type A, except it has a cover designed to assemble with fittings which do not require removal of the cover or any portion of it. SAE 100R2
This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. It consists of an inner tube of oil-resistant synthetic rubber, steel-wire reinforcement according to hose type, as detailed below, and an oil- and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. Type A — This type has two braids of wire reinforcement Type B — This type has two spiral plies and one braid of reinforcement Type AT — This type is the same as Type A, but has a cover designed to assemble with fittings which do not require removal of the cover or any portion of it. Type BT — This type is the same as Type B, but has a cover designed to as-
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This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. It is constructed with an inner tube of oil-resistant synthetic rubber, two braids of suitable textile yarn, and an oil- and weather-resistant synthetic rubber cover.
This hose should be used with petroleum- and water-based hydraulic fluids within a temperature range from140° to 100° C. It consists of an inner tube of oil-resistant synthetic rubber, one braided ply of suitable textile yarn, and an oiland weather-resistant synthetic rubber cover. SAE 100R7
SAE 100R4
This hose should be used in low pressure and vacuum applications, with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. It is constructed with an inner tube of oil-resistant synthetic rubber, a reinforcement consisting of a ply, or plies, of woven or braided textile fibers with a suitable spiral of body wire, and an oil- and weather-resistant synthetic rubber cover.
This thermoplastic hose should be used for synthetic, petroleum-, and water-based hydraulic fluids in a temperature range from140° to 93° C. It consists of a thermoplastic inner tube resistant to hydraulic fluids with suitable synthetic-fiber reinforcement and a hydraulic fluid- and weather-resistant thermoplastic cover. Nonconductive 100R7 is identified with an orange cover and appropriate lay line. Its pressure capacity is similar to that of 100R1. SAE 100R8
SAE 100R5
This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. It is constructed with an inner tube of oil-resistant synthetic rubber reinforced with two textile braids separated by a high-tensile-strength steel-wire braid. All of the braids are impregnated with an oil- and mildew-resistant synthetic rubber compound.
This high-pressure thermoplastic hose should be used with synthetic, petroleum- and water-based hydraulic fluids within a temperature range from140° to 93° C. It consists of a thermoplastic inner tube resistant to hydraulic fluids with suitable synthetic-fiber reinforcement and a hydraulic fluid- and weather-resistant thermoplastic cover. Nonconductive 100R8 is identified with an orange cover and appropriate lay line. Its pressure capacity is similar to that of 100R2.
HYDRAULIC HOSE
SAE 100R9
This hose should be used with petroleum- and water-based hydraulic fluids within a temperature range from 140° to 100° C. Type A—This type consists of an inner tube of oil-resistant synthetic rubber, four spiral plies of wire wrapped in alternating directions, and an oiland weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. Type AT—This type is the same construction as Type A, but has a cover designed to assemble with fittings which do not require removal of the cover or any portion of it. SAE 100R10
This hose should be used with petroleum- and water-based hydraulic fluids within a temperature range from140° to 100° C. Type A—This type consists of an inner tube of oil-resistant synthetic rubber, four spiral plies of heavy wire wrapped in alternating directions, and an oil- and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. Type AT—This type has the same construction as Type A, but its cover is designed to assemble with fittings which do not require removal of the cover or any portion of it. SAE 100R11
This hose should be used with petroleum- and water-based hydraulic fluids within a temperature range from 140° to 100° C. It consists of an inner tube of oilresistant synthetic rubber, six spiral plies of heavy wire wrapped in alternating directions, and an oil- and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. SAE 100R12
SAE 100R14
This hose should be used with petroleum-, synthetic-, and waterbased hydraulic fluids within a temperature range from 154° to 204° C. Type A—This type consists of an inner tube of polytetrafluorethylene (PTFE) reinforced with a single braid of type 303XX stainless steel. Type B—This type has the same construction as Type A, but has the additional feature of an electrically-conductive inner surface to prevent buildup of an electrostatic charge. SAE 100R15
This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from 140° to 121° C. It consists of an inner tube of oil-resistant synthetic rubber, four spiral plies of heavy wire wrapped in alternating directions, and an oil- and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. SAE 100R13
This hose should be used with petroleum-based hydraulic fluids within a temperature range from 140° to 121° C. It consists of an inner tube of oil-resistant synthetic rubber, multiple spiral plies of heavy wire wrapped in alternating directions, and an oil- and weather-resistant rubber cover. A ply, or braid, of suitable material may be used over or within the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire. SAE 100R16
This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from 140° to 121° C. It is constructed with an inner tube of oil-resistant synthetic rubber, followed by multiple spiral plies of heavy wire wrapped in alternating directions, and concluding with an oil- and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire.
This hose should be used with petroleum- and water-based hydraulic fluids, within a temperature range from140° to 100° C. It consists of an inner tube of oil-resistant synthetic rubber, steel wire reinforcement of one or two braids, and an oil-and weather-resistant synthetic rubber cover. A ply, or braid, of suitable material may be used over the inner tube and/or over the wire reinforcement to anchor the synthetic rubber to the wire.
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Hydraulic motors A
ll types of hydraulic motors have these common design features: a driving surface area subject to pressure differential; a way of timing the porting of pressure fluid to the pressure surface to achieve continuous rotation; and a mechanical connection between the surface area and an output shaft. The ability of the pressure surfaces to withstand force, the leakage characteristics of each type motor, and the efficiency of the method used to link the pressure surface and the output shaft determine the maximum performance of a motor in terms of pressure, flow, torque output, speed, volumetric and mechanical efficiencies, service life, and physical configuration. Motor displacement refers to the volume of fluid required to turn the motor output shaft through one revolution. The most common units of motor displacement are in.3 or cm3 per revolution. Displacement of hydraulic motors may be fixed or variable. A fixed-displacement motor provides constant torque. Speed is varied by controlling the amount of input flow into the motor. A variable-displacement motor provides variable torque and variable speed. With input flow and pressure constant, the torque speed ratio can be varied to meet load requirements by varying the displacement. Torque output is expressed in inchpounds or foot-pounds, and is a function of system pressure and motor displacement. Motor torque ratings usually are given for a specific pressure drop across the motor. Theoretical figures indicate the torque available at the motor shaft assuming no mechanical losses. Breakaway torque is the torque required to get a stationary load turning. More torque is required to start a load moving than to keep it moving. Running torque can refer to a motor’s load or to the motor. When it refers to a load, it indicates the torque required to keep the load turning. When it refers to the motor, running torque indicates the
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actual torque which a motor can develop to keep a load turning. Running torque considers a motor’s inefficiency and is a percentage of its theoretical torque. The running torque of common gear, vane, and piston motors is approximately 90% of theoretical. Starting torque refers to the capacity of a hydraulic motor to start a load. It indicates the amount of torque which a motor can develop to start a load turning. In some cases, this is considerably less than the motor’s running torque. Starting torque also can be expressed as a percentage of theoretical torque. Starting torque for common gear, vane, and piston motors ranges between 70 and 80% of theoretical. Mechanical efficiency is the ratio of actual torque delivered to theoretical torque. Torque ripple is the difference between minimum and maximum torque delivered at a given pressure during one revolution of the motor. Motor speed is a function of motor displacement and the volume of fluid delivered to the motor. Maximum motor speed is the speed at a specific inlet pressure which the motor can sustain for a limited time
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Driver
Driven
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Fig. 1. External gear motors have one driving gear and one idler gear enclosed in single housing. Output torque is a function of pressure on one tooth because pressure on other teeth is in hydraulic balance.
without damage. Minimum motor speed is the slowest, continuous, uninterrupted rotational speed available from the motor output shaft. Slippage is the leakage through the motor — or fluid that passes through the motor without performing work. Gear motors External gear motors, Figure 1, consist of a pair of matched gears enclosed in one housing. Both gears have the same tooth form and are driven by pressure fluid. One gear is connected to an output shaft; the other is an idler. Pressure fluid enters the housing at a point where the gears mesh. It forces the gears to rotate, and follows the path of least resistance around the periphery of the housing. The fluid exits at low pressure at the opposite side of the motor. Close tolerances between gears and housing help control fluid leakage and increase volumetric efficiency. Wear plates on the sides of the gears keep the gears from moving axially and help control leakage. Internal gear motors fall into two categories. A direct-drive gerotor motor consists of an inner-outer gear set and an output shaft, Figure 2. The inner gear has one less tooth than the outer. The shape of the teeth is such that all teeth of the inner gear are in contact with some portion of the outer gear at all times. When pressure fluid is introduced into the motor, both gears rotate. The motor housing has integral kidney-shaped inlet and outlet ports. The centers of rotation of the two gears are separated by a given amount known as the eccentricity. The center of the inner gear coincides with the center of the output shaft. In Figure 2(a), pressure fluid enters the motor through the inlet port. Because the inner gear has one less tooth than the outer, a pocket is formed between inner teeth 6 and 1, and other socket A. The kidney-shaped inlet port is designed so that just as this pocket’s volume reaches its
HYDRAULIC MOTORS
Center of outer gear Eccentricity
A
High pressure
Low pressure
A
B 1
G
B
1
6
6
F 2
F 5 F
3
5 4
D
Outlet Center of inner gear port (a)
B
1
5
C
2
E 4
3
4 E
E
Inlet port
6
C
2
A
G
G
C
3 D
D
(b)
(c)
Fig. 2. Direct-drive gerotor motor has internal and external gear sets. Both gears rotate during operation.
maximum, fluid flow is shut off, with the tips of inner gear teeth 6 and 1 providing a seal, Figure 2(b). As the pair of inner and outer gears continues to rotate, Figure 2(c), a new pocket is formed between inner teeth 6 and 5, and outer socket G. Meanwhile, the pocket formed between inner teeth 6 and 1 and outer socket A has moved around opposite the kidney-shaped outlet port, steadily draining as the volume of the pocket decreases. The gradual, metered volume change of the pockets during inlet and exhaust provides smooth, uniform fluid flow with a minimum of pressure variation (or ripple). Because of the extra tooth in the outer gear, the inner gear teeth move ahead of the outer by one tooth per revolution. In Figure 2(c), inner tooth 4 is seated in outer socket E. On the next cycle, inner tooth 4 will seat in outer socket F. This produces a low relative differential speed between the gears. An orbiting gerotor motor, Figure 3, consists of a set of matched gears, a coupling, an output shaft, and a commutator or valve plate. The stationary outer gear has one more tooth than the rotating inner gear. The commutator turns at the same rate as the inner gear and always provides pressure fluid and a passageway to tank to the proper spaces between the two gears. In operation, Figure 3(a), tooth 1 of the inner gear is aligned exactly in socket D of the outer gear. Point y is the center of the stationary gear, and point x is the center of the rotor. If there were no fluid, the rotor would be free to pivot about socket D in either direction. It could move toward seating tooth 2 in socket E or conversely, toward seating tooth 6 in socket J.
When pressure fluid flows into the lower half of the volume between the inner and outer gears, if a passageway to tank is provided for the upper-half volume between the inner and outer gears, a moment is induced which rotates the inner gear counterclockwise and starts to seat tooth 2 in socket E. Tooth 4, at the instant shown in Figure 3(a), provides a seal between pressure and return fluid. However, as rotation continues, the locus of point x is clockwise. As each succeeding tooth of the rotor seats in its socket, Figure 3(b), the tooth directly opposite on the rotor from the seated tooth becomes the seal between pressure and return fluid. The pressurized fluid continues to force the rotor to mesh in a clockwise diX
F
E 2
D
Y
3
2
T = tank P = pressure
(a) F
T2 P1
T3 P3
D T1 P6 J
6
J I
T5
P5 I
T6 (d)
E
T4 P4 H
Y
4
G 5
1
H
H
6
J I
I (b) T3
P2 T2
G
3 D 2
5
H
J
F
E
4 G
1
5
P2
X
D
4
E
F 3
E
G
1 6
Y
rection while it turns counterclockwise. Because of the one extra socket in the fixed gear, the next time tooth 1 seats, it will be in socket J. At that point, the shaft has turned 1/7 of a revolution, and point x has moved 6/7 of its full circle. In Figure 3(c), tooth 2 has mated with socket D, and point x has again become aligned between socket D and point y, indicating that the rotor has made one full revolution inside of the outer gear. Tooth 1 has moved through an angle of 60° from its original point in Figure 3(a); 42 (or 627) tooth engagements or fluid cycles would be needed for the shaft to complete one revolution. The commutator or valve plate, shown in Figures 3(d), (e), and (f), contains pressure and tank passages for each tooth of
(c)
F P3 T4
E
P1 T1 J P6T6
P5
H
T4 P4
T3 P2
G
P4 T5
D
F P3
G T5 P5
D T2 P1 J
T6 P6
T1
I
(e)
H
I (f)
Fig. 3. Orbiting gerotor motor has a stationary outer gear and a rotating inner gear. Rotor and shaft turn in counter clockwise direction, but locus of point X is clockwise. Commutator or valve plate, shown below illustration of each stage of motor rotation, provides pressure and tank passage for pressure fluid.
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HYDRAULIC MOTORS
the rotor. The passages are spaced so they do not provide for pressure or return flow to the appropriate port as a tooth seats in its socket. At all other times, the passages are blocked or are providing pressure fluid or a tank passage in the appropriate half of the motor between gears. A roller-vane gerotor motor, Figure 4, is a variation of the orbiting gerotor motor. It has a stationary ring gear (or stator) and a moving planet gear (or rotor). Instead of being held by two journal bearings, the eccentric arm of the planetary is held by the meshing of the 6-tooth rotor and 7-socket stator. Instead of direct contact between the stator and rotor, roller vanes are incorporated to form the displacement chambers. The roller vanes reduce wear, enabling the motors to be used in closed-loop, high-pressure hydrostatic circuits as direct-mounted wheel drives. Vane motors Vane motors, Figure 5, have a slotted rotor mounted on a drive shaft that is driven by the rotor. Vanes, closely fitted into the rotor slots, move radially to seal against the cam ring. The ring has two major and two minor radial sections joined by transitional sections or ramps. These contours and the pressures introduced to them are balanced diametrically. In some designs, light springs force the vanes radially against the cam contour to assure a seal at zero speed so the motor can develop starting torque. The springs are assisted by centrifugal force at higher speeds. Radial grooves and holes through the vanes equalize radial hydraulic forces on the vanes at all times. Pressure fluid enters and leaves the motor housing through openings in the side plates at the ramps. Pressure fluid entering at the inlet ports moves the rotor counterclockwise. The rotor transports the fluid to the ramp openings at the outlet ports to return to tank. If pressure were introduced at the outlet ports, it would turn the motor clockwise. The rotor is separated axially from the side plate surfaces by the fluid film. The front side plate is clamped against the cam ring by pressure, and maintains optimum clearances as temperature and pressure change dimensions. Vane motors provide good operating efficiencies, but not as high as those of piston motors. However, vane motors generally cost less than piston motors of
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corresponding horsepower ratings. The service life of a vane motor usually is shorter than that of a piston motor. Vane motors are available with displacements of 20 in.3/rev; some low-speed/high-torque models come with displacements to 756 in.3/rev. Except for the high-displacement, low-speed models, vane motors have limited low-speed capability.
Inlet Vane Outlet
Rotor
Piston-type motors Radial-piston motors, Figure 6, have a cylinder barrel attached to a driven shaft; the barrel contains a number of pistons that reciprocate in radial bores. The outer piston ends bear against a thrust ring. Pressure fluid flows through a pintle in the center of the cylinder barrel to drive the pistons outward. The pistons push against the thrust ring and the reaction forces rotate the barrel. Motor displacement is varied by shifting the slide block laterally to change the piston stroke. When the centerlines of the cylinder barrel and housing coincide, there is no fluid flow and therefore the cylinder barrel stops. Moving the slide past center reverses direction of motor rotation. Radial piston motors are very efficient. Although the high degree of precision required in the manufacture of radial piston motors raises initial costs, they generally have a long life. They provide high torque at relatively low shaft speeds and excellent low speed operation with high efficiency; they have limited high speed capabilities. Radial piston motors have displacements to 1000 in.3/rev. Axial-piston motors also use the reciprocating piston motion principle to rotate the output shaft, but motion is axial, rather than radial. Their efficiency characteristics are similar to those of radial-piston motors. Initially, axial-piston motors cost Stationary ring gear (stator)
Eccentric arm
,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,
Planet gear (rotor)
Fig. 4. Roller vane gerotor motor uses rolling vanes to reduce wear.
,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,, ,, ,,, ,, ,,, ,,, ,,,,,,,,,,,,,,,, ,,, ,, ,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,, ,,, ,, ,,, ,,,,,,,,,,,,,,,, ,, ,,, ,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,
Cam ring Outlet
Inlet Drive shaft
Fig. 5. Vane motors (balanced type shown) have vanes in a slotted rotor.
Cylinder block centerline
Centerline Case
Outlet ,,, Pintle ,,, ,,, ,,, ,,,,, ,,, ,,, ,,, ,,,,, ,,, ,,,,, ,,, ,,, ,,,,, ,,,,, ,,, ,,, ,,,,, ,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,, ,,,,, ,,, ,, ,,,,,,, ,,, ,, ,, ,,,,,, ,,,,,, ,,, ,,,,,,, ,, ,,,,,, ,,,,,, ,,, ,,, ,, ,, ,,,,,, ,,,,,,, ,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,, ,,,,,, ,,, ,,,,, ,,,,,, ,,, ,,,, ,, ,,,,, ,,,,,, ,,, ,,,, ,,,, ,, ,,,,, ,,,, ,, ,,,,, ,,,, ,,,, ,,,, ,,,, ,,, ,,,, ,,,, ,,,, ,,, ,,,, ,,,, ,,,, ,,, Inlet ,,,,
,, ,,,,,,, ,,,,,,, ,,, ,,,
Cylinder block
Piston Reaction ring
Fig. 6. Typical radial piston motor.
more than vane or gear motors cost more than vane or gear motors of comparable horsepower, and, like radial piston motors, have a long operating life. Because of this, their higher initial cost may not truly reflect the expected overall costs during the life of a piece of equipment. In general, axial piston motors have excellent high speed capabilities, but, unlike radial piston motors, they are limited at low operating speeds: the inline type will operate smoothly down to 100 rpm and the bent-axis type will give smooth output down to the 4- rpm range. Axial piston motors are available with displacements from a fraction to 65 in.3/rev. Inline-piston motors, Figure 7, generate torque through pressure exerted on the ends of pistons which reciprocate in a cylinder block. In the inline design, the motor drive-shaft and cylinder block are centered on the same axis. Pressure at the ends of the pistons causes a reaction against a tilted swashplate and rotates the
HYDRAULIC MOTORS
Piston
Valve plate slot
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,,,,,,, ,,,,,,, ,,,,,,, ,,,,,,,, ,, ,,,,,,,, ,,,,,,,,,, ,, Outlet port ,,,,,,, ,,,,,,,,, ,, ,,,,,,, ,,,,,,,,,,,,, ,,,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,
Inlet port
Cylinder block
Drive shaft
Swashplate
Fig. 7. Cutaway drawing of inline axial-piston hydraulic motor.
cylinder block and motor shaft. Torque is proportional to the area of the pistons and is a function of the angle at which the swashplate is positioned. These motors are built in fixed- and variable-displacement models. The swashplate angle determines motor displacement. In the variable model, the swashplate is mounted in a swinging yoke, and the angle can be changed by various means — ranging from a simple lever or hand-wheel to sophisticated servo controls. Increasing the swashplate angle increases the torque capacity but reduces drive shaft speed. Conversely, reducing the angle reduces the torque capacity but increases drive shaft speeds (unless fluid pressure decreases). Angle stops are included so that torque and speed stay within operating limits. A compensator varies motor displacement in response to changes in the work load. A spring-loaded piston is connected to the yoke and moves it in response to variations in operating pressure. Any load increase is accompanied by a corresponding pressure increase as a result of the additional torque requirements. The control then automatically adjusts the yoke so that torque increases when the load is light. Ideally, the compensator regulates displacement for maximum performance under all load conditions up to the relief valve setting. Bent-axis piston motors, Figure 8, develop torque through a reaction to pressure on reciprocating pistons. In this design, the cylinder block and drive shaft are mounted at an angle to each other; the reaction is against the drive-shaft flange.
Speed and torque change with changes in the angle—from a predetermined minimum speed with a maximum displacement and torque at an angle of approximately 30˚ to a maximum speed with minimum displacement and torque at about 71/2˚. Both fixed- and variable-displacement models are available. Rotary abutment motors Rotary abutment motors, Figure 9, have abutment A, which rotates to pass rotary vane B, while second abutment C, is in alternate sealing engagement with the rotor hub. Torque is transmitted directly from the fluid to the rotor and from the rotor to the shaft. Timing gears between the output shaft and rotary abutments keep the rotor vane and abutments in the proper phase. A roller in a dovetail groove at the tip of the rotor vane provides a positive seal that is essentially frictionless and relatively insensitive to
wear. Sealing forces are high and friction losses are low because of rolling contact. A screw motor essentially is pump with the direction of fluid flow reversed. A screw motor uses three meshing screws—a power rotor and two idler rotors, Figure 10. The idler rotors act as seals that form consecutive isolated helical chambers within a close-fitting rotor housing. Differential pressure acting on the thread areas of the screw set develops motor torque. The idler rotors float in their bores. The rotary speed of the screw set and fluid viscosity generates a hydrodynamic film that supports the idler rotors, much like a shaft in a journal bearing to permit high-speed operation. The rolling screw set provides quiet, vibration-free operation. Selecting a hydraulic motor The application of the hydraulic motor generally dictates the required horsepower and motor speed range, although the actual speed and torque required may sometimes be varied while maintaining the required horsepower. The type of motor selected depends on the required reliability, life, and performance. Once the type of fluid is determined, the selection of actual size is based on the expected life and the economics of the overall installation on the machine. A fluid motor operating at less than rated capacity will provide a service life extension more than proportional to the reduction in operation below the rated capacity. The maximum horsepower produced by a motor is reached when operating at the maximum system pressure and at the maximum shaft speed. If the motor is alFig. 8. Crosssectional view of bentaxis piston motor.
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HYDRAULIC MOTORS
Rotor
Seal pins
,,,,
Port2
,,
Port1 B
,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,
Tip pin Rotary vane
A
C Rotary abutments
Fig. 9. Abutment A of rotary abutment motor turns past rotary vane B, while second abutment C, contacts seal plate to separate high and low pressure areas. Sealing pins in vane tips and rotor periphery provides nearly frictionless seal. Rotor will turn clockwise with pressure fluid applied to port 1.
ways to be operated under these conditions, its initial cost will be lowest. However, where output speed must be reduced, the overall cost of the motor with speed reduction must be considered — to optimize the overall drive installation costs. Sizing hydraulic motors As an example of how to calculate hydraulic motor size to match an application, consider the following: an application calls for 5 hp at 3000 rpm, with an
available supply pressure of 3000 psi, and a return line pressure of 100 psi; the pressure differential is 2900 psi. The theoretical torque required is calculated from: T = (63,0252 hp) ÷ ∆Pem, where T is torque, lb-in., and N is speed, rpm. For the condition T = 105 lb-in., motor displacement is calculated as: D = 2π T ÷ ∆Pem, where D is displacement, in.3/rev ∆P is pressure differential, psi, and em is mechanical efficiency, %. If mechanical efficiency is 88%, then D is 0.258 in.3/rev. Calculating the required flow: Q = DN/231eV, where Q is flow, gpm, and eV is volumetric efficiency, %. If volumetric efficiency is 93%, then Q is 3.6 gpm. Pressure in these equations is the difference between inlet and outlet pressure. Thus, any pressure at the outlet port reduces torque output of a fluid motor. The efficiency factor for most motors will be fairly constant when operating from half- to full-rated pressure, and over the middle portion of the rated speed range. As speed nears either extreme, efficiency decreases. Lower operating pressures result in
Fig. 10. Screw motor uses power rotor and two idler rotors. Idler rotors act as seals to form consecutive isolated helical chambers in rotor housing.
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lower overall efficiencies because of fixed internal rotating losses that are characteristic of any fluid motor. Reducing displacement from maximum in variable-displacement motors also reduces the overall efficiency. Hydraulic motor malfunctions The majority of motor problems fall into these categories: Improper fluid — The motor is no different than any of the other components of the hydraulic system—it must have clean fluid, in adequate supply, and of the proper quality and viscosity. Poor maintenance — A poor maintenance program runs a close second in the cause of major problems. Typical slips in a program include: • failure to check and repair lines and connections to stop leaks; faulty connections can allow dirt and air into the system, lower pressure, and cause erratic operation. • failure to install the motor correctly. Motor shaft misalignment can cause bearing wear which can lead to lost efficiency. A misaligned shaft also can reduce the torque, increase friction drag and heating, and result in shaft failure. • failure to find the cause of a motor malfunction. If a motor fails, always look for the cause of the failure. Obviously, if the cause is not corrected, failure will recur. Improper operation — Exceeding a motor’s operating limits promotes motor failure. Every motor has design limitations on pressure, speed, torque, displacement, load, and temperature. Excessive pressure can generate heat because of motor slippage, and can cause the motor to exceed torque limits. Excessive speed can cause heating and can cause wear of bearings and other internal parts. Excessive torque can cause fatigue and stress to bearings and the motor shaft, especially on applications that require frequent motor reversing. Excessive load can create bearing and shaft fatigue. And finally, excessive temperature can cause loss of efficiency because the oil becomes thinner, and can produce rapid wear because of lack of lubrication.
Hydraulic pumps W
hen a hydraulic pump operates, it performs two functions. First, its mechanical action creates a vacuum at the pump inlet which allows atmospheric pressure to force liquid from the reservoir into the inlet line to the pump. Second, its mechanical action delivers this liquid to the pump outlet and forces it into the hydraulic system. A pump produces liquid movement or flow: it does NOT generate pressure. It produces the flow necessary for the development of pressure which is a function of resistance to fluid flow in the system. For example, the pressure of the fluid at the pump outlet is zero for a pump not connected to a system (load). Further, for a pump delivering into a system, the pressure will rise only to the level necessary to overcome the resistance of the load. Classification of pumps All pumps may be classified as either positive displacement or non-positive displacement. Most pumps used in hydraulic systems are positive displacement. A non-positive displacement pump produces a continuous flow. However, because it does not provide a positive internal seal against slippage, its output varies considerably as pressure varies. Centrifugal and propeller pumps are examples of non-positive displacement pumps. If the output port of a non-positive displacement pump were blocked off, the pressure would rise, and output would decrease to zero. Although the pumping element would continue moving, flow would stop because of slippage inside the pump. In a positive-displacement pump, slippage is negligible compared to the pump’s volumetric output flow. If the output port were plugged, pressure
would increase instantaneously to the point that the pump’s pumping element or its case would fail (probably explode, if the drive shaft did not break first), or the pump’s prime mover would stall. Positive displacement principle A positive displacement pump is one that displaces (delivers) the same amount of liquid for each rotating cycle of the pumping element. Constant delivery during each cycle is possible because of the close-tolerance fit between the pumping element and the pump case. That is, the amount of liquid that slips past the pumping element in a positive displacement pump is minimal and negligible compared to the theoretical maximum possible delivery. The delivery per cycle remains almost constant, regardless of changes in pressure against which the pump is working. Note that if fluid slippage is substantial, the pump is not operating properly and should be repaired or replaced. Positive displacement pumps can be of either fixed or variable displacement. The output of a fixed displacement pump remains constant during each pumping cycle and at a given pump speed. The output of a variable displacement pump can be changed by altering the geometry of the displacement chamber. Other names to describe these pumps are hydrostatic for positive displacement and hydrodynamic pumps for non-positive displacement. Hydrostatic means that the pump converts mechanical energy to hydraulic energy with comparatively small quantity and velocity of liquid. In a hydrodynamic pump, liquid velocity and movement are large; output pressure actually depends on the velocity at which the liquid is made to flow. Reciprocating pumps The positive displacement principle is well illustrated in the reciprocating-
A hydraulic pump is a mechanical device which converts mechanical energy into hydraulic energy. It provides the force necessary to move a liquid, thus transmitting power.
type pump, the most elementary positive displacement pump, Figure 1. As the piston extends, the partial vacuum created in the pump chamber draws liquid from the reservoir through the inlet check valve into the chamber. The partial vacuum helps seat firmly the outlet check valve. The volume of liquid drawn into the chamber is known because of the geometry of the pump case, in this example, a cylinder. As the piston retracts, the inlet check valve reseats, closing the valve, and the force of the piston unseats the outlet check valve, forcing liquid out of the pump and into the system. The same To system Outlet
Inlet
,, ,, ,, ,, ,, Check valve
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Fig. 1. Reciprocating pump. Outlet
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Driver
Driven
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Fig. 2. Spur gear pump.
1998/1999 Fluid Power Handbook & Directory
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HYDRAULIC PUMPS
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Outlet ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, Inlet
,,,,,,,, ,,,,,,,, ,,,,,,,,
Fig. 3. Lobe pump.
amount of liquid is forced out of the pump during each reciprocating cycle. All positive displacement pumps deliver the same volume of liquid each cycle (regardless of whether they are reciprocating or rotating). It is a physical characteristic of the pump and does not depend on driving speed. However, the faster a pump is driven, the more total volume of liquid it will deliver. Rotary pumps In a rotary-type pump, rotary motion carries the liquid from the pump inlet to the pump outlet. Rotary pumps are usually classified according to the type of element that transmits the liquid, so that we speak of a gear-, lobe-, vane-, or piston-type rotary pump. External-gear pumps Gear pumps can be divided into external and internal gear types. A typical external gear pump is shown in Figure 2. These pumps come with a straight spur, helical, or herringbone gears. Straight spur gears are easiest to cut and are the most widely used. Helical and herringbone gears run more quietly, but cost more. A gear pump produces flow by carrying fluid in between the teeth of two meshing gears. One gear is driven by the drive shaft and turns the idler gear. The chambers formed between adjacent gear teeth are enclosed by the pump housing and side plates (also called wear or pressure plates). A partial vacuum is created at the pump inlet as the gear teeth unmesh. Fluid flows in to fill the space and is carried around the outside of the gears. As the teeth mesh again at the outlet end, the fluid is forced out. Volumetric efficiencies of gear pumps run as high as 93% under optimum con-
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ditions. Running clearances between gear faces, gear tooth crests and the housing create an almost constant loss in any pumped volume at a fixed pressure. This means that volumetric efficiency at low speeds and flows is poor, so that gear pumps should be run close to their maximum rated speeds. Although the loss through the running clearances, or “slip,” increases with pressure, this loss is nearly constant as speed and output change. For one pump the loss increases by about 1.5 gpm from zero to 2000 psi regardless of speed. Change in slip with pressure change has little effect on performance when operated at higher speeds and outputs. External gear pumps are comparatively immune to contaminants in the oil, which will increase wear rates and lower efficiency, but sudden seizure and failure are not likely to occur. Lobe pumps The lobe pump is a rotary, external gear pump, Figure 3. It differs from the conventional external gear pump in the way the “gears” are driven. In a gear pump, one gear drive the other; in a lobe pump, both lobes are driven through suitable drives gears outside of the pump casing chamber. Screw pumps A screw pump is an axial-flow gear pump, Figure 4. Three types of screw pumps are the single-screw, two-screw, and three-screw. In the single-screw pump, a spiraled rotor rotates eccentrically in an internal stator. The two-screw
pump consists of two parallel intermeshing rotors rotating in a housing machined to close tolerances. The three-screw pump consists of a central-drive rotor with two meshing idler rotors; the rotors turn inside of a housing machined to close tolerances. Flow through a screw pump is axial and in the direction of the power rotor. The inlet hydraulic fluid that surrounds the rotors is trapped as the rotors rotate. This fluid is pushed uniformly with the rotation of the rotors along the axis and is forced out the other end. Note that the fluid delivered by screw pumps does not rotate, but moves linearly. The rotors work like endless pistons which continuously move forward. There are no pulsations even at higher speed. The absence of pulsations and the fact that there is no metal-to-metal contact results in very quietly operating pumps. Larger pumps are used as low-pressure, large-volume prefill pumps on large presses. Other applications include hydraulic systems on submarines and other uses where noise must be controlled. Internal-gear pumps Internal gear pumps, Figure 5, have an internal gear and an external gear. Because these pumps have one or two less teeth in the inner gear than the outer, relative speeds of the inner and outer gears in these designs are low. For example, if the number of teeth in the inner and outer gears were 10 and 11 respectively, the inner gear would turn 11 revolutions, while the outer would turn 10. This low relative speed means a low wear rate. These pumps are small, com-
Fig. 4. Non-pulsating axial flow three-screw pump.
HYDRAULIC PUMPS
Case Gerotor element
Internal gear
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Inlet
External gear Outlet
Internal
,,,,,,,,, gear ,,,,,,,,, ,,,,,,,,, ,,,,,,,,, ,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,, ,,,,,,,,, Inlet,,,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,, Crescent ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Fig. 5. Internal gear pumps — gerotor and crescent. Rotor
Cam ring surface
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Eccentricity ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,, ,,,,,,,,, ,,,,,,,,,,,,,,,,,,, ,,,,,,,,,, Housing
Vane
Inlet Vane Outlet
Rotor
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Cam ring Outlet
Inlet Drive shaft
Fig. 6. Unbalanced vane pump.
Fig. 7. Balanced vane pump.
pact units.
the teeth start to mesh again. The seal is provided by the sliding contact. Generally, the internal gear pump with toothcrest pressure sealing has higher volumetric efficiency at low speeds than the crescent type. Volumetric and overall efficiencies of these pumps are in the same general range as those of external gear pumps. However, their sensitivity to dirt is somewhat higher.
Crescent-seal pumps The crescent seal internal gear pump consists of an inner and outer gear separated by a crescent-shaped seal. The two gears rotate in the same direction, with the inner gear rotating faster than the outer. The hydraulic oil is drawn into the pump at the point where the gear teeth begin to separate and is carried to the outlet in the space between the crescent and the teeth of both tears. The contact point of the gear teeth forms a seal, as does the small tip clearance at the crescent. Although in the past this pump was generally used for low outputs, with pressures below 1000 psi, a 2-stage, 4000-psi model has recently become available. Gerotor pumps The gerotor internal gear pump consists of a pair of gears which are always in sliding contact. The internal gear has one more tooth than the gerotor gear. Both gears rotate in the same direction. Oil is drawn into the chamber where the teeth are separating, and is ejected when
Vane pumps In these pumps, a number of vanes slide in slots in a rotor which rotates in a housing or ring. The housing may be eccentric with the center of the rotor, or its shape may be oval, Figure 6. In some designs, centrifugal force holds the vanes in contact with the housing, while the vanes are forced in and out of the slots by the eccentricity of the housing. In one vane pump, light springs hold the vanes against the housing; in another pump design, pressurized pins urge the vanes outward. During rotation, as the space or chamber enclosed by vanes, rotor, and housing increases, a vacuum is created, and atmospheric pressure forces oil into this space, which is the inlet side of the
pump. As the space or volume enclosed reduces, the liquid is forced out through the discharge ports. Balanced and unbalanced vane pumps The pump illustrated in Figure 6 is unbalanced, because all of the pumping action occurs in the chambers on one side of the rotor and shaft. This design imposes a side load on the rotor and drive shaft. This type vane pump has a circular inner casing. Unbalanced vane pumps can have fixed or variable displacements. Some vane pumps provide a balanced construction in which an elliptical casing forms two separate pumping areas on opposite sides of the rotor, so that the side loads cancel out, Figure 7. Balanced vane pumps come only in fixed displacement designs. In a variable-volume unbalanced design, Figure 8, the displacement can be changed through an external control such as a handwheel or a pressure compensator. The control moves the cam ring to change the eccentricity between the ring and rotor, thereby changing the size of the pumping chamber and thus varying the displacement per revolution. When pressure is high enough to overcome the compensator spring force, the cam ring shifts to decrease the eccentricity. Adjustment of the compensator spring determines the pressure at which the ring shifts. Because centrifugal force is required to hold the vanes against the housing and maintain a tight seal at those points, these pumps are not suited for low-speed service. Operation at speeds below 600 rpm is not recommended. If springs or other means are used to hold vanes out against the ring, efficient operation at speeds of 100 to 200 rpm is possible. Vane pumps maintain their high efficiency for a long time, because compensation for wear of the vane ends and the housing is automatic. As these surfaces wear, the vanes move further out in their slots to maintain contact with the housing. Vane pumps, like other types, come in double units. A double pump consists of two pumping units in the same housing. They may be of the same or different sizes. Although they are mounted and driven like single pumps, hydraulically, they are independent. Another variation is the series unit: two pumps of equal capacity are connected in series, so that the
1998/1999 Fluid Power Handbook & Directory
A/121
HYDRAULIC PUMPS
Eccentricity
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Compensator
Slide block Outlet
Maxium volume stop screw Inlet
Compensator spring
Fig. 8. Variable displacement, pressure compensated vane pump.
output of one feeds the other. This arrangement gives twice the pressure normally available from this pump. Vane pumps have relatively high efficiencies. Their size is small relative to output. Dirt tolerances is relatively good. Piston pumps The piston pump is a rotary unit which uses the principle of the reciprocating pump to produce fluid flow. Instead of using a single piston, these pumps have many piston-cylinder combinations. Part of the pump mechanism rotates about a drive shaft to generate the reciprocating motions, which draw fluid into each cylinder and then expels it, producing flow. There are two basic types, axial and radial piston; both area available as fixed and variable displacement pumps. The second variety often is capable of variable reversible (over-
center) displacement. Most axial and radial piston pumps lend themselves to variable as well as fixed displacement designs. Variable displacement pumps tend to be somewhat larger and heavier, because they have added internal controls, such as handwheel, electric motor, hydraulic cylinder, servo, and mechanical stem. Axial piston pumps The pistons in an axial piston pump reciprocate parallel to the centerline of the drive shaft of the piston block. That is, rotary shaft motion is converted into axial reciprocating motion. Most axial piston pumps are multi-piston and use check valves or port plates to direct liquid flow from inlet to discharge. Inline piston pumps The simplest type of axial piston
Piston
Valve plate slot
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,,,,, ,, ,,,,, ,,,,,,, ,,,, ,,,,,,,, ,,,,,,,, ,, ,,,,,,,, Outlet port ,,,,,,,,,,, ,,,,,,,,, ,,,,,,,,, ,,,,,,,,,,,,, ,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,, ,,,, ,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,
Inlet port
Fig. 9. Inline, axial piston pump.
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Cylinder block
Drive shaft
Swashplate
pump is the swashplate design in which a cylinder block is turned by the drive shaft. Pistons fitted to bores in the cylinder block are connected through piston shoes and a retracting ring, so that the shoes bear against an angled swashplate. As the block turns, Figure 9, the piston shoes follow the swashplate, causing the pistons to reciprocate. The ports are arranged in the valve plate so that the pistons pass the inlet as they are pulled out and the outlet as they are forced back in. In these pumps, displacement is determined by the size and number of pistons as well as their stroke length, which varies with the swashplate angle. In variable displacement models of the inline pump, the swashplate swings in a movable yoke. Pivoting the yoke on a pintle changes the swashplate angle to increase or decrease the piston stroke. The yoke can be positioned with a variety of controls, i.e., manual, servo, compensator, handwheel, etc. Bent axis pumps This pump consists of a drive shaft which rotates the pistons, a cylinder block, and a stationary valving surface facing the cylinder block bores which ports the inlet and outlet flow. The drive shaft axis is angular in relation to the cylinder block axis. Rotation of the drive shaft causes rotation of the pistons and the cylinder block. Because the plane of rotation of the pistons is at an angle to the valving surface plane, the distance between any one of the pistons and the valving surface continually changes during rotation. Each individual piston moves away from the valving surface during one-half of Cylinder block centerline
Centerline Case
Outlet Pintle
,,, ,,, ,,, ,,, ,,, ,,, ,,,
,,, ,,,,, ,,, ,,, ,,,,, ,,,,, ,,, ,,, ,,,,, ,,,,, ,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,, ,, ,,, ,,,,,,, ,,, ,,,,,, ,, ,, ,,,,,, ,,, ,,,,,,, ,, ,,,,,, ,,, ,,, ,,,,,, ,, ,, ,,,,,, ,,,,,,, ,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,, ,,,,,, ,,, ,,,,, ,,,,,, ,,, ,, ,,,,, ,,,, ,,,, ,,,,,, ,,, ,, ,,,,, ,,,, ,,,, ,, ,,,,, ,,,, ,,,, ,,,, ,,,, ,,, ,,,, ,,,, ,,,, ,,, ,,,, ,,,, ,,,, ,,, ,,,,
,, ,,, ,, ,,, ,,,, ,, ,,, ,,,
Cylinder block
Inlet Piston Reaction ring
Fig. 10. Radial piston pump.
HYDRAULIC PUMPS
Pressure
Pressure
Load flow
Maximum flow
X Load pressure
Pump pressure = load pressure Pump flow = load flow
X Maximum flow
Relief valve setting Maximum pressure = relief valve setting
Flow Flow Useful power Useful power Wasted power
Pump operating point
X
Pump operating point
X
Maximum power
Maximum power
,,,, ,,,,
Fig. 11. Pressure-flow curve of hydraulic system with fixed displacement pump.
Fig. 12. Pressure flow curve of hydraulic system with variable displacement pump equipped with displacement control.
Pressure compensator
Yoke control
Load/pump pressure
Yoke
Metered pressure Exhaust
Load pressure
Pressure at pump outlet
Fig. 13. Schematic of typical proportional pump pressure compensator control.
the shaft revolution and toward the valving surface during the other half. The valving surface is so ported that its inlet passage is open to the cylinder bores in that part of the revolution where the pistons move away. Its outlet passage is open to the cylinder bores in the part of the revolution where the pistons move toward the valving surface. Therefore, during pump rotation the pistons draw liquid into their respective cylinder bores through the inlet chamber and force it out through the outlet chamber. Bent axis pumps come in fixed and variable displacement configurations, but cannot be reversed. Radial piston pumps In these pumps, the pistons are arranged radially in a cylinder block; they move perpendicularly to the shaft center-
line. Two basic types are available: one uses cylindrically shaped pistons, the other ball pistons. They may also be classified according to the porting arrangement: check valve or pintle valve. They are available in fixed and variable displacement, and variable reversible (overcenter) displacement. In pintle-ported radial piston pump, Figure 10, the cylinder block rotates on a stationary pintle and inside a circular reacting ring or rotor. As the block rotates, centrifugal force, charging pressure, or some form of mechanical action causes the pistons to follow the inner surface of the ring, which is offset from the centerline of the cylinder block. As the pistons reciprocate in their bores, porting in the pintle permits them to take in fluid as they move outward and discharge it as
they move in. The size and number of pistons and the length of their stroke determine pump displacement. Displacement can be varied by moving the reaction ring to increase or decrease piston travel, varying eccentricity. Several controls are available for this purpose. Plunger pumps These reciprocating pumps are somewhat similar to rotary piston types, in that pumping is the result of pistons reciprocating in cylinder bores. However, the cylinders are fixed in these pumps; they do not rotate around the drive shaft. Pistons may be reciprocated by a crankshaft, by eccentrics on a shaft, or by a wobble plate. When eccentrics are used, return stroke is by springs. Because valving cannot be supplied by covering and uncovering ports as rotation occurs, inlet and outlet check valves may be used in these pumps. Because of their construction, these pumps offer two features other pumps do not have: one has a more positive sealing between inlet and outlet, permitting higher pressures without excessive leakage of slip. The other is that in many pumps, lubrication of moving parts other than the piston and cylindrical bore may be independent of the liquid being pumped. Therefore, liquids with poor lubricating properties can be pumped. Volumetric and overall efficiencies are close to those of axial and radial piston pumps. Measuring pump performance Volume of fluid pumped per revolution is calculated from the geometry of the oil-carrying chambers. A pump never quite delivers the calculated, or theoretical, amount of fluid. How close it comes is called volumetric efficiency. Volumetric efficiency is found by comparing the calculated delivery with actual delivery. Volumetric efficiency varies with speed, pressure, and the construction of the pump. A pump’s mechanical efficiency is also less than perfect, because some of the input energy is wasted in friction. Overall efficiency of a hydraulic pump is the product of its volumetric efficiency and the mechanical efficiency. Pumps are generally rated by their maximum operating pressure capability
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HYDRAULIC PUMPS
Maximum pressure
Load pressure Pump flow = load flow
Maximum flow
Pressure
X
Flow Useful power Wasted power
Pump operating point.
X Maximum power
cause the pump puts out full pressure and flow regardless of load demand. A relief valve prevents excessive pressure buildup by routing high-pressure fluid to tank when the system reaches the relief setting. As Figure 11 shows, power is wasted whenever the load requires less than full flow or full pressure. The unused fluid energy produced by the pump becomes heat that must be dissipated. Overall system efficiency may be 25% or lower. Variable displacement pumps, equipped with displacement controls, Figure 12, can save most of this wasted hydraulic horsepower when moving a single load. Control variations include hand wheel, lever, cylinder, stem servo,
,,, ,,,,,, ,,,, , ,,,, , ,,,,
Fig. 14. Pressure-flow curve of hydraulic system with variable displacement pump equipped with a pressure compensator.
Two - stage pressure compensator Orifice
Load pressure
Relief valve
Stroking piston
Yoke position + 80 %
Bias position
Pressure at pump outlet
Pump/load pressure Pilot/control pressure Metered pressure Exhaust
Fig. 15. Schematic of pump two-stage compensator control.
and their output, in gpm, at a given drive speed, in rpm.
Matching pump power with the load Pressure compensation and load sensing are terms often used to describe pump features that improve the efficiency of pump operation. Sometimes these terms are used interchangeably, a misconception that is cleared up once you understand the differences in how the two enhancements operate. To investigate these differences, consider a simple circuit using a fixed-displacement pump running at constant speed. This circuit is efficient only when the load demands maximum power be-
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and electrohydraulic servo controls. Examples of displacement control applications are the lever-controlled hydrostatic transmissions used to propel windrowers, skid-steer loaders, and road rollers. While matching the exact flow and pressure needs of a single load, these controls have no inherent pressure or power-limiting capabilities. And so, other provisions must be made to limit maximum system pressure, and the prime mover still must have corner horsepower capability. Moreover, when a pump supplies a circuit with multiple loads, the flow and pressure-matching characteristics are compromised.
A design approach to the system in which one pump powers multiple loads is to use a pump equipped with a proportional pressure compensator, Figure 13. A yoke spring biases the pump swashplate toward full displacement. When load pressure exceeds the compensator setting, pressure force acts on the compensator spool to overcome the force exerted by the spring. The spool then shifts toward the compensator-spring chamber, ports pump output fluid to the stroking piston, and decreases pump displacement. The compensator spool returns to neutral when pump pressure matches the compensator spring setting. If a load blocks the actuators, pump flow drops to zero. Using a variable-displacement, pressure-compensated pump rather than a fixed-displacement pump reduces circuit horsepower requirements dramatically, Figure 14. Output flow of this type of pump varies according to a predetermined discharge pressure as sensed by an orifice in the pump’s compensator. Because the compensator itself operates from pressurized fluid, the discharge pressure must be set higher — say, 200 psi higher — than the maximum load-pressure setting. So if the load-pressure setting of a pressurecompensated pump is 1100 psi, the pump will increase or decrease its displacement (and output flow) based on a 1300-psi discharge pressure. A two-stage pressure-compensator control, Figure 15, uses pilot flow at load pressure across an orifice in the main stage compensator spool to create a pressure drop of 300 psi. This pressure drop generates a force on the spool which is opposed by the main spool spring. Pilot fluid flows to tank through a small relief valve. A spring chamber pressure of 4700 psi provides a compensator control setting of 5000 psi. An increase in pressure over the compensator setting shifts the main stage spool to the right, porting pump output fluid to the stroking piston, which overcomes bias piston force and reduces pump displacement to match load requirements. The earlier stated misconception stems from an observation that output pressure from a pressure-compensated pump can fall below the compensator setting while an actuator is moving. This does not happen because the pump is
HYDRAULIC PUMPS
80
4
5
1
2
3
4
5
1
2
3
4
Typical loader & excavator systems Pressure overshoot - %
60 Typical proportional compensator 40
Two - stage compensator 20
10,000
100,000 Rate of pressure rise - psi/sec
Fig. 16. Typical performance of a single- and two-stage compensator control.
sensing the load, it happens because the pump is undersized for the application. Pressure drops because the pump cannot generate enough flow to keep up with the load. When properly sized, a pressure-compensated pump should always force enough fluid through the compensator orifice to operate the compensator.
sator setting, Figure 16. Note that in system equipped with an accumulator, a two-stage compensator control provides little advantage. In excavator hydraulic systems, however, superiority of the two-stage compensator is evident: it provides system components much greater protection against pressure transients.
Superior dynamically With respect to its matching function, a two-stage compensator is identical to the proportional compensator control shown in Figure 13. The dynamic performance of the two-stage control is superior, however. This becomes obvious when one analyzes a transient which involves a sudden decrease in load flow demand, starting from full stroke at low pressure. The single-stage control spool ports pressure fluid to the stroke piston only when pump discharge pressure reach the compensator setting. The main-stage spool of the two-stage control starts moving as soon as pump discharge pressure minus spring chamber pressure exceeds the 300-psi spring setting. Because pilot fluid flows through the orifice and because of the flow needed to compress the fluid in the spring chamber, the spring chamber pressure lags pump discharge pressure. This causes the spool to become unbalanced and shift to the right. Pump destroking starts before pump discharge pressure reaches the compen-
Load sensing: the next step A similar control, which has recently become popular, is the load sensing control, sometimes called a power matching control, Figure 17. The single-stage valve is almost identical to the singlestage compensator control, Figure 13, except that the spring chamber is connected downstream of a variable orifice rather than directly to tank. The loadsensing compensator spool achieves equilibrium when the pressure drop across the variable orifice matches the 300-psi spring setting. Any of three basic load-sensing signals control a load-sensing pump: unloaded, working, and relieving. In the unloaded mode, the lack of load pressure causes the pump to produce zero discharge flow at bias or unload pressure. When working, load pressure causes the pump to generate discharge flow in relation to a set pressure drop, or bias pressure. When the system reaches maximum pressure, the pump maintains this pressure by adjusting its discharge flow. Like the pressure-compensated
pump, a load-sensing pump has a pressure-compensation control, but the control is modified to receive two pressure signals, not just one. As with pressure compensation, the load-sensing control receives a signal representing discharge pressure, but it also receives a second signal representing load pressure. This signal originates from a second orifice downstream from the first. This second orifice may be a flow-control valve immediately beyond the pump outlet, the spool opening of a directional control valve, or it may be a restriction in a fluid conductor. Comparison of these two pressure signals in the modified compensator section allows the pump to sense both load and flow. This reduces power losses even further, Figure 18. Output flow of the pump varies in relation to the differential pressure of the two orifices. Just as the pressure-compensated pump increased its discharge pressure by the amount required to run the pressure compensator, the load- and flow-sensing pump’s discharge pressure typically is between 200 and 250 psi higher than actual load pressure. Furthermore, a load-sensing pump can follow the load and flow requirements of a single circuit function or multiple simultaneous functions, relating horsepower to maximum load pressure. This consumes the lowest possible horsepower and generates the least heat. Operator control If the variable orifice is a manually operated flow control valve, the system can operate in a load-matched mode at the direction of an operator. As he opens the flow control valve, flow increases proportionally (constant pressure drop across an increasing-diameter orifice), at a pressure slightly above load pressure. As suggested in Figure 18, wasted power is very low with a load-sensing variable volume pump compensator. Since the control senses pressure drop and not absolute pressure, a relief valve or other means of limiting pressure must be provided. This problem is solved by a load-sensing/pressure-limiting control, Figure 19. This control functions as the load-sens-
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,,, ,, ,,, ,,, ,,,, ,, , , , ,, ,, HYDRAULIC PUMPS
Load sensing compensator
Pressure limiting compensator
Load sensing compensator
Yoke control
Yoke control
Yoke
Pump pressure
Yoke
Load pressure Metered pressure
Load pressure
Exhaust
Load pressure
Pressure at pump outlet
Metered pressure Pump pressure
Fig. 17. Schematic of proportional pump compensator which provides load-sensing capability.
Maximum flow
Pressure
Load pressure
Priority
Hydrostat with adjustable spring
Pump operating point
X Maximum popwer
Fig. 18. Pressure-flow curve of pump with load-sensing compensator control.
ing control previously described, until load pressure reaches the pressure limiter setting. At that point, the limiter portion of the compensator overrides the loadsensing control to destroke the pump. Again, the prime mover must have corner horsepower capability. Load-sensing gear pumps Piston and vane pumps rely on their variable-displacement capability to accomplish load sensing. How, then, can a gear pump accomplish load sensing if its displacement is fixed? Like standard gear pumps, load-sensing gear pumps have low initial cost when compared to other designs with equivalent flow and pressure capabilities. However, load-sensing gear pumps offer the versatility of variable-displacement axial-piston and
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,, ,,, ,,,,,,,,, ,,,,,,, ,,,,,, ,,,,, ,,,,,, ,,,,, ,,,,,, ,,,,, ,, ,,,
Load-sensing signal
Hydrostat Adjustable spring
Fig. 20. Load sensing gear pumps with two different types of hydrostats installed. The spring adjustment allows tuning pressure drop for different manufacturers’ valves or line lengths.
vane pumps but without the high complexity and high cost of variable-displacement mechanisms. A load-sensing gear pump is a versatile pump of patented design that can: ● provide the high efficiency of load sensing without the high cost associated with piston or vane pumps ● produce zero to full output flow in less than 40 milliseconds with little or no pressure spiking and without pump
,, ,,,,,,,,, ,,, ,,,,,, ,,,,,,,,, ,,,,,,
Useful power Wasted power
Load-sensing signal Unloader control
Load flow Flow
,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,,,,, ,,
~ 200
X
,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,,,,, ,,
Pump pressure
Load-sensing signal Unloader control
Load-sensing signal
,,
Relief valve setting
Fig. 19. Schematic of pump control which provides load sensing and pressure limiting.
,,, ,,,,,,, ,,, ,,,,,,, ,,,,, ,,,,,
Load pressure
Exhaust
Variable orifice
,, ,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,
Variable orifice
Pressure at pump outlet
Fig. 21. Unloader control has been added to the load-sensing gear pump. The control uses a poppet or a plunger to allow maximum flow at the minimum pressure drop across the unloader with minimal control movement.
inlet supercharging ● drive circuits with low (approaching atmospheric) unload relief pressures ● provide priority flow and secondary flow with a low unload pressure to reduce standby and secondary loaded power draw, and ● interchange with load-sensing vane or piston pumps without having to change line or component sizes. Load-sensing piston pumps use a pressure compensator and a hydrostat to vary volumetric output to a system in reference to load pressure and flow re-
HYDRAULIC PUMPS
,, ,,,,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,
Pilot relief
,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,,,,, ,, Fig. 22. Combined control is achieved by incorporating a pilot relief, which causes the hydrostat to act as the main stage of a pilot-operated relief valve.
,
LOW UNLOAD POPPET
LOAD SENSE PORT
,,, ,,,, ,, ,,,,
,, ,,,,,, ,, ,,,, ,,,,,,,,,, ,,
,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,,
,,, ,,,,,,,,,,,,,
RELIEF VALVE
A load-sensing gear pump, on the other hand, uses a hydrostat in combination with an unloader to vary its volumetric output in response to load and flow requirements. Because load-sensing piston and gear pumps both use a single load-sensing signal to control pump discharge pressure and flow, they are interchangeable in load-sensing circuits. Both types have much in common and offer substantial power savings over systems using fixed-displacement pumps. Both offer reduced power consumption in the running mode — when flow and pressure are required to operate a function. They also conserve
BYPASS PORT
RELIEF TO INLET
G 20 LS PUMP HEAD
PUMP DISCHARGE
HYDROSTAT
Fig. 23. This cutaway shows combined control, which has an adjustable hydrostat contained within the unloader control. Locating the hydrostat within the low-unload control allows all piston areas to operate from a single load-response signal. It is intended for applications using large pumps where secondary flow bypasses to tank.
quirements. A hydrostat is a springloaded device that meters flow according to the spring force across its equal but opposing effective areas. It may be restrictive, as in a series circuit, or it may bypass primary load pressure to a secondary or tank pressure. In simple terms, a hydrostat separates the total flow into two flows: one represents the required flow and the other represents the required pressure of the primary circuit. A load-sensing piston pump uses its hydrostat to regulate output flow relative to load pressure and bypasses the excess pump flow to a secondary route, which may be ported to tank or to a secondary circuit.
power in the standby mode — when the system is idling or in a non-operational mode. Furthermore, they can reduce the required size — and, therefore, cost— of valves, conductors, and filters needed for the circuit. The load-sensing gear pump illustrated in Figure 20 minimizes power consumption in the running mode by separating total discharge flow according to a remote primary function pressure and a primary flow. This is accomplished through a single load-sensing signal originating from the priority circuit and routed as close as possible to the discharge side of the pump’s gears. Adding an unloader control to the
pump circuit, Figure 21, allows the system to conserve power in the standby mode of operation as well as in the running mode. This control must be installed in parallel with the inlet port of the hydrostat and as close as possible to the discharge side of the gears. It must be piloted by the same load-sensing signal as in Figure 20. This signal causes the pump to dump all flow from the outlet to the secondary circuit and at a pressure well below the hydrostat’s pressure-drop setting in the standby mode. The unloader control must operate off the same remote load-sensing signal that controls the hydrostat. Unlike the hydrostat, the unloader poppet of the unloader control is designed with opposing areas having a ratio of at least 2:1. Any line pressure sensed that exceeds 50% of pump discharge pressure will close the unloader control. The ability of the unloader control to unload the pump to near atmospheric discharge pressure is controlled by the poppet or plunger spring force. The unloader control is set to the lowest value to maintain the internal pressure loading of the gear pump. When compared to a standard fixed-displacement gear pump circuit, this control can reduce standby power consumption by 90%. Dual and combined controls The load-sensing signal can be conditioned by limiting pressure in the remote sensing line or taking it to 0 psig. Doing so causes the hydrostat and the unloader control of the load-sensing gear pump to respond to the conditioned signal according to the discharge pressure. This is accomplished by providing a pilot relief, Figure 22, which causes the hydrostat to act as the main stage of a pilot-operated relief valve. The ability to condition the load-sensing line is patented and makes the loadsensing gear pump useful for functions other than just load sensing. The combined-control load-sensing gear pump, Figure 23, is intended for large-displacement pumps and bypasses secondary flow to tank. It also is patented, and can be used in the same applications as the dual-control pump. However, because secondary flow must be routed to tank, it cannot be used when the secondary circuit drives a load.
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Hydrostatic transmissions T
he primary function of any hydrostatic transmission (HST) is to accept rotary power from a prime mover (usually an internal combustion engine) having specific operating characteristics and transmit that energy to a load having its own operating characteristics. In the process, the HST generally must regulate speed, torque, power, or, in some cases, direction of rotation. Depending on its configuration, the HST can drive a load from full speed in one direction to full speed in the opposite direction, with infinite variation of speed between the two maximums — all with the prime mover operating at constant speed. The operating principle of HSTs is simple: a pump, connected to the prime mover, generates flow to drive a hydraulic motor, which is connected to the load. If the displacement of the pump and motor are fixed, the HST simply acts as a gearbox to transmit power from the prime mover to the load. The overwhelming majority of HSTs, however, use a variable-displacement pump, motor, or both so that speed, torque, or power can be regulated. HSTs offer many important advantages over other forms of power transmission. Depending on its configuration, an HST: ● transmits high power in a compact size ● exhibits low inertia ● operates efficiently over a wide range of torque-to-speed ratios ● maintains controlled speed (even in reverse) regardless of load, within design limits ● maintains a preset speed accurately against driving or braking loads ● can transmit power from a single prime mover to multiple locations, even if position and orientation of the
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Hydrostatic transmissions (HSTs) boast a long list of characteristics that make them the first choice for countless applications.
locations changes efficient the entire power system. Ulti● can remain stalled and undamaged mately the power system should be deunder full load at low power loss signed for a balance between efficiency and productivity. A machine designed ● does not creep at zero speed ● provides faster response than me- for maximum efficiency usually has chanical or electromechanical trans- sluggish response, which robs producmissions of comparable rating, and tivity. Conversely, a machine designed for quick response usually exhibits low ● can provide dynamic braking. Either of two types of construction is efficiency because a high degree of enused for HSTs: integral and non-inte- ergy must be available at all times to gral. Non-integral construction is by far perform work — even when there is no the most common, because power can immediate need for work. be transmitted to one or more loads in areas that would otherwise be difficult Four functional types of HSTs to access. In this technique, the pump is The configuration of an HST — coupled to the prime mover, the motor whether it has a fixed- or variable disis coupled to the load, and they are con- placement pump, motor, or both —denected through hose and tubing assem- termines its performance characterisblies, Figure 1. Integral construction tics. Figure 3 summarizes these combines pump, motor, and all other configurations and the performance hydraulic components into a single characteristics of each. housing, Figure 2. The advantage here The simplest form of hydrostatic is an economical, compact package that transmission uses a fixed-displacemay contain axles, mounting surfaces, ment pump driving a and other components in addition to the HST. Whatever its task, the HST must be designed for an optimum match between the engine and the load. This allows the engine to operate at its most efficient speed and the HST to make adjustments to operating conditions. The better the match be- Fig. 1. A typical hydrostatic transmission consists of a variable-displacement tween input and pump and fixed-displacement motor connected through metal tubing, hose asoutput character- semblies, or both. Providing a reservoir (and usually a heat exchanger and filtraistics, the more tion system) between the pump and motor forms an open-circuit HST.
H Y D R O S TAT I C T R A N S M I S S I O N S
Closed circuit or closed loop?
T
here are four basic HST configurations; two open-circuit and two closed-circuit configurations. Both refer to how the hydraulic lines in the system are connected. In an open circuit, fluid is drawn into the pump through a reservoir, is routed to the motor, then re-enters the reservoir after passing through the hydraulic motor. In a closed circuit, the flow path is uninterrupted — fluid flows in a continuous path from the pump discharge port to the fluid motor inlet port, out the motor discharge port and back into the pump inlet. The two types of open and closed circuit systems are open loop and closed loop, which refer to the control arrangement of the HST. An open-loop system has no means of feedback for speed, pressure, flow, or torque regulation. Any variable control settings are accomplished manually by the operator. Closedloop control, however, incorporates feedback devices that provide communication between the pump and motor, so the HST automatically adjusts to variations in operating conditions of the load, engine, or both.
fixed-displacement motor, Figure 3A. Although this transmission is inexpensive, its applications are limited, primarily because alternative forms of power transmission are much more energy efficient. Because pump displacement is fixed, the pump must be sized to drive the motor at a fixed speed under full load. When full speed is not required, fluid from the pump outlet passes over the relief valve. This wastes energy in the form of heat. Using a variable-displacement pump instead of one with a fixed displacement creates a constant torque transmission, Figure 3B. Torque output is constant at any speed because torque depends only on fluid pressure and motor displacement. Increasing or decreasing pump displacement increases or decreases motor speed, re-
spectively, while torque remains fairly constant. Power, therefore, increases with pump displacement. Using a variable-displacement motor with a fixed-displacement pump produces a transmission that delivers constant power, Figure 3C. If flow to the motor is constant, and motor displacement is varied to maintain the product of speed and torque constant, then power delivered is constant. Decreasing motor displacement increases motor speed but decreases torque, a combination that maintains constant power. The most versatile HST configuration teams a variable-displacement pump with a variable-displacement motor, Figure 3D. Theoretically, this arrangement provides infinite ratios of torque and speed to power. With the motor at maximum displacement, varying pump output directly varies speed and power output while torque remains constant. Decreasing motor displacement at full pump displacement increases motor speed to its maximum; torque varies inversely with speed, and horsepower remains constant. The curves in Figure 3D illustrate two ranges of adjustment. In Range 1, motor displacement is fixed at maximum; pump displacement is increased from zero to maximum. Torque remains constant as pump displacement increases, but power and speed increase. Range 2 begins when the pump reaches maximum displacement, which is maintained while the motor’s displacement decreases. Throughout this range, torque decreases as speed increases, but power remains constant. (Theoretically, motor speed could be increased infinitely, but from a practical stand-
point, it is limited by dynamics.) Application example #1 Assume that a 3116-lb-in. torque load must be driven at 1000 rpm with a fixed-displacement HST. Power required is determined from: P = (T 3 N) 4 63,024 where: P is power in hp T is torque in lb-in., and N is speed in rpm. Therefore, P = (3116 3 1000)4 63,024 = 50 If we choose a 2000-psi pump (based on experience for providing a good combination of size, weight, performance, and cost) rated at 50 hp, we then calculate the flow it must deliver: q = (1714 3 P)4 p where: q is flow in gpm, and p is pressure in psi Therefore, q = (1714)(50) / 2000 = 43 gpm We then select a hydraulic motor with a displacement of 10 in.3/rev to deliver 3116 lb-in. of torque at 2000 psi — approximately 43 gpm at 1000 rpm. Figure 3A shows the power/torque/ speed characteristics for the pump and motor, assuming the pump operates at constant speed. Pump flow is maximum at this operating speed, and the pump attempts to deliver this quantity of oil to the fixeddisplacement hydraulic motor. Load inertia makes it impossible to accelerate instantaneously to full speed, so part of the pump output flows over the relief valve. (Figure 3A also illustrates the power loss during acceleration.) As the motor increases speed, it transmits more of the pump’s output, and less oil flows over the relief valve. At rated speed, all oil flows through the motor. Fig. 2. Packaged HST encloses pump, motor, controls, conducting system, and all auxiliary components into a single housing. The unit shown accepts input power from a V-belt drive and transmits power to the load through its output shaft. Packaged HSTs are available in a variety of configurations, many of which bolt directly to an engine.
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H Y D R O S TAT I C T R A N S M I S S I O N S
Torque is constant because system pressure builds to the relief valve setting immediately after the control valve shifts. Power lost over the relief valve is the difference between the constant power delivered by the pump and the variable power delivered by the motor. The area under this curve represents the power wasted when the transmission starts or stops. It also shows the low ef-
Torque/speed ratio Theoretically, the maximum power a hydrostatic transmission can trans-
P4, Q4
NI, TI
NO, TO
NI, TI
Efficiency - % and torque - lb-ft 310
PR, QR Motor torque at relief valve setting Max. Max. Pump power
Motor power
NO, TO
250
100 80
200
60
150
40
Output torque
Overall efficiency 100 50
20 Output power 0
Power
Torque
mit is a function of flow and pressure. However, in constant-power transmissions with variable output speeds, theoretical power divided by the torque/speed ratio determines actual power output. The greatest constant power that can be transmitted is determined by the lowest output speed at which this constant power must be transmitted.
Power - hp
P1, Q1
ficiency for any operating speed below maximum. A fixed-displacement transmission is not recommended for applications requiring frequent starts and stops or when less than full load torque occurs frequently.
Power lost over relief valve
400
800 1200 Speed - rpm
1600
0
C
0
0
Max.
Motor speed
NO, TO
A
NO TO
NI, TI
300
100 80
Overall effieiency
200
60 40
100
Power - hp
Efficiency - % and torque - lb-ft 3 10
Output torque
Torque, power, and flow
Power
Flow
Torque
Output power 20 0
0 400
800 1200 Speed-rpm B
1600
0
Max. Range 1
Range 2 Speed D
Fig. 3. Functional hydrostatic transmissions summarized according to types of pumps and motors involved: Fig. A shows HST with fixeddisplacement pump and motor; Fig. B has fixed motor and variable-displacement pump; Fig C has fixed pump and variable-displacement motor, and Fig. D has a variable-displacement pump and motor.
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H Y D R O S TAT I C T R A N S M I S S I O N S
rpm. Above critical speed, torque decreases as speed increases, which provides constant power. Torque
Building a closed-circuit HST The descriptions of closed-circuit hydrostatic transmissions in Figure 3 concentrate on parametric considerations only. Additional functions must be provided to achieve a practical HST. Pump-end components — Consider, for example, a constant-torque HST — the type used most commonly — with a servo-controlled, variabledisplacement pump driving a fixed-displacement motor, Figure 5A. Because this is a closed-circuit HST, slip flow accumulates in the pump and motor cases and is removed through a case drain line, Figure 5B. The combined
Torque and power
For example, if the minimum speed represented by point A on the power curve in Figure 4 is half the maximum speed, the torque-to-speed ratio is 2:1. The maximum power that can be transmitted is half the theoretical maximum. At point B, corresponding to output speed A, the torque curve decreases as speed increases. At maximum output speed, it has dropped to point C. At output speeds less than half the maximum, torque remains constant at its maximum value, but power decreases in proportion to speed. The speed at point A is the critical speed and is determined by the dynamics of the HST’s components. Below critical speed, power decreases linearly (with constant torque) to zero at zero
B
A
r
we
C
Po
0
Maximum42
Maximum
Speed
Fig. 4. Critical speed (indicated by point A) in a constant-power HST is the lowest speed at which maximum constant power can be transmitted.
NO, TO To or from pump
NO, TO
A
E
D
To or from pump
Charge pump
D
B
To or from pump
B
F
To or from motor
G C A To or from motor C
To or from pump E
Fig. 5. Progression of constant-power HST circuits — from a bare pump and motor to an assembly with basic accessories.
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H Y D R O S TAT I C T R A N S M I S S I O N S
case drains flow to the reservoir then through the drain line to the V is volume of fluid on pressure side in through a heat exchanger. pump case. Fluid then returns to the in.3, and One of the most important features charge pump reservoir through the Qcp is charge pump output in in.3/sec. of a closed-circuit HST is a charge heat exchanger. pump. The charge pump is usually an Application example #2 integral part of the main pump pack- Cavitation control Assume that the HST of Figure 5 is age but can also be an independent The stiffness of an HST depends on connected with 2 ft of 11/2-in. ID steel pump gang-mounted with the drive the compressibility of the fluid and the tubing. Neglecting the volumes of pumps it serves, Figure 6. Whatever compliance of system components, pump and motor, V is about 30 in.3 For the arrangement, the charge pump namely, tubing, and hoses. The influ- oil in steel tubing, B e is 200,000 psi. performs two functions. First, it pre- ence of these components can be com- Assuming the charge pump delivers 6 vents cavitation of the main pump by pared to the effect a spring-loaded ac- gpm (28 in.3/sec), then the rate of presreplenishing the fluid lost by the cumulator would have if connected to sure rise is: closed system through pump and mo- the supply line through a tee fitting. dp/dt = (200,000 328)4 30 tor slip. It also provides pressurized Under light loads, the effective accu= 190,000 psi/sec. fluid required by the variable-dis- mulator spring compresses slightly; unNow consider the effect of plumbing placement control mechanism. der heavy loads, the accumulator un- the system with 20 ft of 1 1 / 2 -in. ID, Referring now to Figure 5C, low- dergoes substantial compression, and three-wire braided hose. The hose manpressure relief valve A on the dis- there is more fluid in the accumulator. ufacturer would have to provide the charge side of the charge pump sets This additional fluid volume must be volumetric expansion coefficient in control pressure. Although charge supplied by the charge pump. in.3/ 1000-psi to calculate the effective pressures vary from one pump manuThe critical factor is the rate of bulk modulus. Assume, for this examfacturer to another, they typically pressure rise in the system. If pres- ple, that Be is about 84,000 psi. Then: dp/dt = (84,000 328) 4 294.5 range between 250 and 300 psi. sure rises too rapidly, the rate of vol= 7986 psi/sec Back-to-back replenishing check ume increase on the supply side (soIncreasing the output of valves B and C supply the charge pump would be make-up fluid to the approthe most effective way to priate low-pressure line. prevent the tendency of such Motor-end components a system to cavitate. Alter— A typical, closed-circuit nately, if changes in the exHST also requires crossternal load are not continuover relief valves D and E, ous, an accumulator can be Figure 5D. These usually added to the charge circuit. are integrated into the moIn fact, some HST manufactor package. Two crossturers provide a port for conover relief valves are innecting an accumulator to stalled to prevent excessive the charge circuit. pressure from developing Fig. 6. Gang mounting multiple pumps provides a single, compact assembly If the stiffness of the in either supply line due to that supplies two or more independent circuits from the rear drive pad of a HST is low, and it is shock-load feedback gasoline or diesel engine. In this example, two variable-displacement axialequipped with automatic through the motor, an over- piston pumps are visible at left; a fixed-displacement vane pump, at right, controls, the HST should running load, or similar serves as a charge pump. be started with pump disconditions. These valves limit pressure in either pressure sup- called compressibility flow) may placement at zero. In addition, accelply line by routing high-pressure exceed the flow capacity of the eration of the displacement mechafluid to the low-pressure line. These charge pump, and the main pump nism should be limited to prevent relief valves perform the same func- may cavitate. Perhaps circuits pow- jerky starts, which, in turn, could tion as a system relief valve in an ered by variable-displacement pumps generate excessive pressure surges. open circuit. However, they are lo- with automatic controls pose the most Some HST manufacturers provide cated at the fluid motor end because serious threat of danger. When such a damping orifices in the stroking cirthis is where overpressures originate system cavitates, pressure drops or cuit for this very purpose. This discussion demonstrates the in closed-circuit HSTs. disappears altogether. The automatic In addition to cross-over relief controls attempt to respond, resulting multi-faceted role of the charge pump system in a closed-circuit HST. Therevalves, shuttle valve F is included. in an unstable system. The shuttle valve is always shifted by Mathematically, the rate of pressure fore, system stiffness and control of the rate of pressure rise may be the primary high-pressure fluid, which connects rise can be expressed as: considerations for determining charge the low-pressure line to low-pressure dp/dt = BeQcp 4V relief valve G. Valve G routes excess where: Be is effective bulk modulus of pump delivery, rather than simply main pump and motor slip flows. charge pump flow to the motor case, the system, psi
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Intensifiers I
ntensifiers, or boosters, convert low-pressure fluid power into higher-pressure fluid power. One type of intensifier resembles a cylinder in appearance and operation. Another type is commonly known as a multiplesection, internal-gear flow divider; when it is appropriately connected in a circuit, pressure intensification results. Cylinder-type intensifiers, Figure 1, have a large and a small piston mounted on a common rod or shaft. Each piston is housed in bores of appropriate diameter. The end of the rod serves as the small piston in many models. A source of low-pressure fluid (usually shop air although oil is not uncommon) is made available to the large piston, while the smaller piston most often receives oil at low pressure. When activated, the low-pressure fluid powers the large piston to increase or intensify the pressure output from the smaller piston chamber. This fluid is pumped to a work cylinder for the job at hand. Area ratios The ratio of intensification is the same as the ratio of the two piston ar-
eas. If these areas are in a ratio of 10:1, for example, and 100 psi shop air strokes the larger piston, the resulting pressure can be 1000 psi, depending on the load that is to be moved. These intensifiers can be single or double acting. The single-acting intensifier, often called a one shot, must have some mechanical means to return the pistons once they have operated. Double-acting intensifiers use external, manually-operated valving to cycle the intensifier. Another variation uses the external valve to control the beginning and end of the intensification cycle, and also has internal valving mechanically operated by the larger piston as it bottoms, to provide reciprocating motion of the piston and pumping of intensified fluid to the work cylinder. An additional variation of cylinder intensifiers uses a double-acting, doublerod-end cylinder in which each rod-end alternately feeds intensified-pressure fluid to the work actuator, thus providing intensified-pressure in each direction of intensifier stroke. Generally, cylinder intensifier circuits are either single- or dual-pressure circuits. In a single-pressure circuit, the Low-pressure piston
High-pressure inlet passage
High-pressure pistons Seals
Low-pressure ports
Fig. 1. Double-acting intensifier uses rod end as small piston area, discharges intensified fluid out port at left end.
intensifier produces high-pressure oil for operation of the entire work cylinder stroke. These circuits are recommended when the work cylinder approach stroke is short compared to its higher-pressure stroke. The dual-pressure circuit uses an air-oil tank to extend the work cylinder through its lower-pressure approach stroke before the intensifier begins operation. This circuit can save up to 90% of the air required for single-pressure operation. Flow dIvider intensification An internal gear pump with multiple gear sections mounted on a common shaft usually can be a multi-sectioned pump or a rotary flow divider. When used as a flow divider, the amount of energy expended across the divider as it operates remains the same, minus a small efficiency loss. When output of one or more of the divider sections is returned to tank, the same energy situation remains true: the energy expended remains the same, minus a small efficiency loss. When output of one or more of the divider sections is returned to tank, the same energy situation remains true: the energy expended remains the same. This means that the input horsepower of the gear sections returning fluid to tank is applied through the common shaft to the working gear sections to drive them as motors to increase their power output, again minus the small efficiency losses. When the flow divider is connected in a circuit in this manner, pressure fluid downstream of the working gear sections can be raised perhaps to a level above the pressure capability of the main pump and relief valve setting. This use also can reduce time that the main system pump must operate at maximum pressure.
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Low-speed/hightorque motors B
asic low-speed/high-torque (LSHT) motor designs include: internal gear, vane, radial piston, axial piston, axial ball piston, rolling vane, and radial-piston/constant-acceleration cam. Many factors influence the operating performance of LSHT motors; consequently, direct comparison is virtually impossible. Here are some general points to consider: Gerotor motors are more economical, but their leakage rates tend to lower volumetric efficiency, making them better-suited for low pressure operation. Their mechanical efficiency is reasonable. Vane motors have a large number of leakage paths and tend to have lower volumetric efficiency at low speeds. These motors are radially-balanced, which improves their mechanical efficiency and extends their operating life. Vane motor operation tends to improve at lower pressures. Rolling-vane motors require precisely-controlled tolerances and tend to cost more. However, their volumetric efficiency is nearly constant at all speeds. These motors are also radially balanced. Radial-piston motors exhibit good leakage characteristics and consequently, good volumetric efficiency Reduced ,,,,,,,,,,,,,,,,, Rotor A B rise ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, A B ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, B A ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,, A-Inlet A B ,,,,,,,,,,,,,,,,, B-Outlet ,,,,,,,,,,,,,,,,, Vane ,,,,,,,,,,,,,,,,,
Ring
(a) Basic
ball pistons are small, so volumetric efficiency can be good. Torque efficiencies are about 80%.
throughout their speed range. Starting torques are good; starting torque efficiencies for motors with an eccentric crankshaft are about 85%. Motors with constant-acceleration cams have starting torque efficiencies to 95%. Radial piston motors using eccentric crankshafts or eccentric, circular cam rings may exhibit some torque and speed variation caused by harmonic piston motion. Effects which may occur at very high speeds include intense whine and flow pulsation. At very low speeds there may be torque or speed flutter, or cogging of the output shaft. Close attention to the manufacturer’s operating recommendations for maximum and minimum speed limits is essential. Constant-acceleration cam radial piston motors eliminate these harmonic difficulties because at any moment, the sum of piston velocities is always zero. However, these motors are more expensive than eccentric crankshaft types. Axial-piston motors have good volumetric efficiencies, particularly at lower pressures, and usually have good starting torque characteristics. Axial-ball piston motors with multiple wave cams are pressure-balanced to operate without pulsation or vibration. Operating clearances around the B
Normal rise
A1
Split Rise
B
A1
A2
B
A2
B
B
A2
B
A2
A1 (b) Reduced Rise
Construction There are many variations among basic LSHT motor designs. The following are representative: Gear motors usually are of Gerotor design and consist of a Gerotor set, a splined drive coupling, and a commutator valve. The Gerotor set has a stationary outer ring which is part of the motor housing, and a rotor. The outer ring has integral gear teeth which mesh with mating teeth on the rotor. The rotor has one less tooth than the outer ring. A 6-lobe/7-lobe gear set has a 6:1 mechanical advantage. Pressure fluid forces the rotor to revolve inside the outer ring while orbiting around the center of the outer ring. A coupling transmits the motion of the rotor to the output shaft. Each tooth of the rotor is in sliding contact with the outer ring at all times. The commutator valve, connected to and rotating with the output shaft, ports pressure fluid to the spaces between the gear teeth. Pressure and return passages in the commutator valve are connected to the motor ports through the housing. As a valve rotates, the fluid passages which keep the
B
(c) Even Split
A1
B
(d) Even Split
Fig. 1. Variations of standard vane motor is available in (a) basic, (b) reduced rise, (c) even split rise, and (d) uneven split rise designs.
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LSHT MOTORS
Fig. 2. Rolling-vane motor is hydraulically balanced because pressure fluid always acts on equal and opposite areas.
Fig. 3. Pistons and control valve in radialpiston motor are hydrostatically balanced.
pressure fluid in phase with the opening and closing of the spaces between the gear teeth are subject to pressure fluid, three are connected to the return line, and the seventh is blocked. The motor is bidirectional, depending on which port is pressurized. Models for heavy-duty applications, and those which require higher shaft load capacities use precision-machined rollers to form the displacement chambers. The rollers provide support and rolling contact as the inner gear rotates to minimize friction. In reduced-rise models, Figure 1b, the cam-rise of the ring or stroke of the vane is reduced to provide a smaller displacement in the same basic frame size. In split rise models, Figure 1c, two inlet ports feed individual opposing pairs of rises. When all rises are the same, it is termed an even split, Figure 1c; when one set of opposed ports has reduced rises, it is termed an uneven split, Figure 1d. An uneven split gives three torque/speed combinations at any
one given flow and pressure. With the split concept, top speed can be obtained at half pump capacity if maximum torque is not required at the same time. Full torque is obtained at half rated speed. All basic, reduced rise, and split rise models may be stacked to produce from two to four times the basic torque in each frame size. Vane motors resemble traditional hydraulic vane motors, Figure 1, where vanes ride a ring cam and slide in and out of rotor slots. The rotor is independent of the ring and is centered by a shaft. The rotor puts only pure torque into the shaft. There are no side loads. Rolling-vane motors, Figure 2, have fluid flow through the motor body to the rotor shaft through a combined body-and-shaft groove. Passages within the rotor shaft port highpressure fluid radially into pressure chambers; symmetrical passages port low-pressure fluid back through the rotor shaft and body to the outlet port. The rolling vanes function as timing valves to sequence the fluid to insure that each pair of rotor vanes always has high pressure against its trailing surface, and low pressure at its leading surface. Because the pressure fluid always acts on equal and opposite areas, the rotors are always in hydraulic balance. Radial-piston motors have a great variety of designs within the basic radial configuration. For example, one design of this type motor produces hydraulic thrust on an eccentric crankshaft which produces output torque, Figure 3. The hydraulic load on the piston crown is transmitted to the connecting rod through a passage to a pressure-compensated hydrostatic bearing. This creates a high-pressure column of oil which is trapped in a relieved area of the connecting rod foot. The trapped static oil column transmits piston force to the crankshaft with only a small percentage of the load being carried on the connecting rod foot. As system pressure increases, pressure fluid trapped in the connecting rod pocket increases, automatically compensating for increased piston loading and keeping the percentage of mechanical loading constant at a low level.
Fig. 4. Rotary valve distributes pressure fluid from inlet to each piston in sequence, forcing rollers against cam ring, causing it and outer case of motor to rotate.
The radially mounted pistons in motors of another design, Figure 4, reciprocate inside bores in a stationary cylinder housing and are connected to rollers that bear on a cam ring. A rotary valve distributes pressure fluid to each cylinder in sequence. Fluid pressure pushes the pistons outward, forcing the rollers against the cam ring, causing it and the outer casing of the motor to rotate. Fixed side guides absorb reaction forces on the rollers, “unloading” the pistons from any tangential forces. A two-speed valve that attaches directly to the motor with no intermediate plumbing permits operation at two different speeds without changing flow to the motor. In the higher speed range, the torque is halved. A third design also produces thrust on an eccentric crankshaft through connecting rods, a heat-treated ring, and spherical roller bearing. The ring follows the turning crank without rotating itself. Since there is no friction between the ring and the connecting rod shoes, hydrostatic balancing is not necessary. The seven pistons have a working frequency of three and four; before the first of the three pistons under pressure reaches the top of its stroke, a fourth piston pressurizes to smooth rotation at low speed.
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Manifolds M
anifolds come in two basic types. One is a single-piece design which supports all necessary valving and contains all the passages for an entire system. The other is the modular-block design. Each modular block usually supports only one valve and contains internal passages for that valve’s functions as well as flowthrough provisions. It normally is connected to a series of similar modular blocks to make up a complete system. Both types have their advantages. Which is best suited for your system will depend on a variety of actors, such as application, specific function, cost, space, and system longevity. Some manifold manufacturers supply only manifolds; others provide manifolds that go only with their valves; still others supply manifolds and valves, but are willing to sell you their manifold and permit you to select valves of your choice. As more valves are built with standard mounting patterns, interchangeability becomes feasible for more and more systems.
Single-piece manifolds These are available in two basic designs: laminar and drilled metal block. Laminar design — In a laminar-type manifold, several layers of metal have appropriate passages machined or milled through them. These plates, usually steel, are stacked or sandwiched with the various fluid paths determined by the shape of the overlapping passages. Solid-metal end pieces are added, and the whole stack is brazed together.
For more information on manifolds, see the chapter on cartridge valves, beginning on page A/35. A/136
Manifolds provide a convenient and economical means for locating multiple valves into a centralized location. Benefits include lower cost, lighter and more compact design, fewer components, less leakage, and simpler maintenance.
Because the internal passages can be cut in contoured shapes and as large as necessary, nearly any flow rate can be accommodated with virtually no pressure drop. Because the stack is brazed together, these manifolds can handle pressures to 10,000 psi, and there is no limit to the number or size of the valves which can be mounted on the manifold. Laminar manifolds are custom-designed. Valves and other connections can be located where appropriate for a specific application. But because of the permanently shaped flow passages and brazed construction, this type manifold cannot be modified easily if future circuit changes become necessary. Drilled metal-block design — Drilled metal block manifolds, Figure 1, also can be custom-designed for specific applications. Usually made from a slab of steel, aluminum, or cast iron, the blocks are drilled to provide flow passages for design requirements. This network of drilled passages also enables you to locate valves as desired, with some limitations because the drilled passages must be straight. Other drilled-block manifolds accept cartridge valves into cavities drilled into the manifold surface. Interconnecting flow passages travel through the manifold from the valve cavities. Some cartridge valves have threaded bodies that hold them in threaded cavities; others slip into smooth cavities where they are retained by plates on the manifold surface. Modular manifolds Modular manifold systems, Figures 2 and 3, allow relatively easy modification of existing manifolds. This erector-set approach to manifold construction consists of cast iron, aluminum, or steel blocks which permit you to design
and build your own manifold. They also can be ordered ready-to-install. Most modular systems can be benchassembled horizontally and stacked. End plates usually seal the ends of the assembled manifold, but these plates also can be drilled for pump and tank connections. Interconnecting, divider, and spacer plates are usually installed between the basic building blocks. Interconnecting plates divert flow from one passage to another between blocks, or stop flow between blocks by plugging a passage. Divider plates allow flow to continue or to be blocked by plugging.
Fig. 1. (Top view) Free-standing drilledblock manifold accepts three different valves; (Bottom view) Drilled metal-block manifolds with 1- and 11⁄4-in. ports accommodate high flows.
MANIFOLDS
Spacer plates serve to increase dimensions between basic blocks when an outsized valve must be accommodated on the mounting surface. The tops of the basic modular blocks are ported and drilled to accept subplate-mounted valves. Blocks with different ports usually are available for each type of valve and subplate. Each type is identified for the valve it will accept. One company has color-coded their blocks so that block function can be determined according to color. Some modular systems will accept cartridge valves as well as subplatemounted valves. When specifying a manifold system, be sure to examine the interchangeability of valves, subplates, and blocks. Manifold assembly The methods of manifold-block connection vary. Some modular manifolds
Fig. 2. Compact modular manifold system includes directional, sequence, and relief valves, plus other control components.
Fig. 3. Horizontal manifold stacking system can mount directly on tank with integrated filter and pressure-relief valve, includes 5 symmetrical channels, adaptable to any type of circuit. Valve-mounting patterns conform to DIN standards.
use tie rods which extend through the blocks and are secured with nuts. Some have external flanges on each block which bolt together for connection. Others have a connecting hardware system which uses socket-headed studs which are also threaded in the socket to accept the threads of the next stud. All blocks and plates have provisions in their fluid-conducting passages for O-rings to provide a seal between sections when they are drawn together. Some manifold systems have pump and tank connections in the bottom of the block, others locate them in the end plates. One advantage of the bottom port is that pressure fluid can be introduced in the center of the circuit, or at the division between two parallel circuits, thereby reducing the distance through which pressure fluid must flow. An advantage of the endplate connection is that they permit base-mounting of the manifold, which is convenient on some equipment. Some manifold designs arrange to close-couple the pump to the manifold, eliminating the pressure line to the manifold. Pneumatic manifolds, Figure 4, often have separate exhausts at each block or work station because there is no need to return this fluid to tank. This permits exhaust metering which in turn, allows individual workstation speed control. Restrictions Flow-rates, pressures, and manifold length may be limiting factors with some modular manifold systems. One system is limited to 20 gpm at 3000 psi, and the manifold package cannot be more than 3 ft long. Another is restricted to 35 gpm and 3000 psi. A third accommodates fluids at pressures to 6000 psi and has no length limit. These restrictions can be designed around, but be sure the system you specify can adequately handle your needs. Electrical connections to manifolds may be made with wires which lead directly from the power source to the appropriate solenoid. Some manifold systems have built-in electrical troughs or channels for interior runs of electric cable. Some pneumatic systems have electrical connections at the valve subplate, which provide plug-in connec-
Fig. 4. Low pressures in pneumatic systems allow manifolds to be made from easily shaped materials such as clear acrylic.
tions for the valve. This reduces the maze and clutter of the installation. Advantages of manifolds There are many advantages in using a manifold system. For example, one manifold which could replace approximately 300 lb of tubing and valving, occupies about 1 ft3 of space. Design engineers list these advantages and tips for the the use of manifolds. • assembly and installation costs reduced 30% to 50% • space required for installation reduced by 33% • single-pressure-type pneumatic manifolds require only one filter-regulatorlubricator • placement of all valves and manifolds should be considered during the design stage; manifolds should be located as close as possible to the equipment they are to control • if four of five valves are to be operated simultaneously, connect the supply to both manifold ends • when designing and specifying a manifold, the system designer should work with the manifold manufacturer. It is nearly impossible to catalog all manifold options or substitute catalog information for the detailed knowledge and experience of the manufacturer. Modular fitting manifolds Modular manifolds for fittings provide rapid connect and disconnect capability with leak-free connections. Some are made of sealed pipe with fittings or valve ports added at intervals along the length of the pipe. These are commonly used in fluid, instrument air, and some pneumatic applications.
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Pressurecontrol valves
R
elief, unloading, sequence, and counterbalance valves are normally-closed. They open partially and/or fully, while performing their intended functions. A reducing valve is normally-open. It restricts, and ultimately blocks flow to a secondary circuit. In either case, a restriction is usually necessary to produce the required pressure control. One exception is the externally-piloted unloading valve, which depends on an external signal for its actuation . Relief valves Most fluid power systems are designed to operate within a preset pressure range. This range is a function of the forces the actuators in the system must generate to do the required work.
In their simplest form, pressure controls are 2-way valves that are either normally-closed (no flow through the valves) or normally-open (flow through the valve) Most pressure control valves have infinite positioning: they can assume an infinite number of positions between their fully open and fully closed positions, depending on flow rates and pressure differentials.
Being able to control and limit these forces could damage the fluid power components and expensive equipment. Relief valves avoid this hazard. They are the safeguards which limit maximum pressure in a system by diverting excess oil when pressures get too high. Relief valves can be divided in two categories: direct-acting and pilotoperated. Direct-acting — A direct-acting valve may consist of a poppet or ball held exposed to system fluid pressure on one side and opposed by a spring of preset force on the other. In a fixed, non-adjustable relief valve, Figure 1, when the valve is normally closed, the force exerted by the compression spring exceeds the force exerted by system fluid pressure acting on the ball or poppet. The spring holds the ball or poppet
System pressure
tightly seated. A reservoir port on the spring side of the valve returns leakage fluid to reservoir. When system pressure begins to exceed the setting of the valve spring, the fluid unseats the ball or poppet, allowing a controlled amount of fluid to bypass to reservoir, keeping system pressure at the valve setting. The spring re-seats the ball or poppet when enough fluid is released (by-passed) to drop system pressure below the setting of the valve spring. Because the usefulness of a fixed relief valve would be limited to the one setting of its spring, most relief valves are adjustable. This is commonly achieved with an adjusting screw acting on the spring, Figure 2. By turning the screw in or out, the operator compresses or decompresses the spring respec-
System pressure
Spring adjustment screw
Fig. 1. Simple, direct-acting relief valve has no adjusting screw and therefore opens a fixed, pre-set pressure as controlled by setting of compression spring.
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Fig. 2. Adjustable, direct-acting relief valve prevents flow through the valve until force of system pressure on the poppet overcomes the adjustable spring force and downstream pressure.
P R E S S U R E - C O N T R O L VA LV E S
Main relief valve
Vent port Pilot valve
Axial orifice
Pilot
Pilot spring
Main valve
System pressure Drain port
Reduced pressure (secondary circuit)
Fig. 3. Pilot-operated relief valve has an orifice through the piston and is held closed by the force of a light spring and system pressure acting on the larger piston area at the spring end. When system pressure rises to overcome the force of the pilot valve spring, the pilot valve opens. Flow through the orifice creates a pressure drop across the piston. Pressure differential unseats piston and opens valves.
tively. The valve can be set to open at any pressure within a desired range. Aside from the adjustable feature, this valve works just like the fixed valve, Figure 1. Pressure override and cracking pressures — The pressure at which a relief valve first begins to open to allow fluid to flow through is known as cracking pressure. When the valve is bypassing its full rated flow, it is in a state of full flow pressure. The difference between full flow pressure and cracking pressure is sometimes known as pressure differential, also known as pressure override. In some cases, this pressure override is not objectionable. However, it can be a disadvantage as it can result in wasted power because of the fluid lost through the valve before the maximum setting is reached. This can further permit maximum system pressure to exceed the ratings of other components. To minimize override, use a pilot-operated relief valve, which is described in the pressure-control valves section. Poppet design — Spring-loaded poppet valves are generally used for small flows. They don’t leak below cracking pressure and have a fast response, making them ideal for relieving shock pressures. They are often used as safety valves to prevent damage to components from high surge pressures, or to relieve pressure caused by thermal expansion in locked cylinders.
Fig. 4. Direct-acting, pressure reducing valve is held open by spring force. As reduced pressure at outlet port increases above the loading of the spring, it gradually moves the spool to the right, closing the valve, and increasing pressure drop across it. When valve closes completely, a small quantity of fluid drains through the spool to tank, preventing reduced pressure from increasing because of valve leakage.
If poppet relief valves must operate frequently, chatter caused when the valve relieves may damage the seat. The following is one theory on the cause of chatter. When the poppet unseats, fluid starts flowing and accelerates near the valve opening. The area subject to full fluid pressure when the valve was closed is now subject to a changed pressure at its periphery. A combination of fluid pressure and velocity combined with some over travel of the poppet in the opening direction, results in too much fluid flow, subsequent pressure drop, and return of the poppet toward its seat. The rapid repetition of this cycle may produce anything from a pulsation to a high pitched whistle. Some reverse acting relief valves use a dashpot to cushion the valve’s closing and eliminate chatter. The differential between cracking and full open pressure on spring-loaded poppet relief valves is high. For this reason they are not recommended for close pressure control. Reverse flow and guided piston designs — Relief valves are also made to relieve flow in either direction. Fluid pressure at the other port acts on a shoulder on the plunger to open the valve. Another type of direct-acting relief valve is the guided piston. In this valve a sliding piston, rather than a poppet, connects the pressure and reservoir ports.
System pressure acts on the piston and moves it against a spring force. As the piston moves, it uncovers a reservoir port in the valve body. These valves have a fast response but may be prone to chatter. They can be damped to eliminate chatter, but this also slows valve reaction time. They are reliable and can operate with good repetitive accuracy if flow does not vary widely. Valves with hardened steel pistons and sleeves have a very long service life. They may leak slightly below cracking pressure unless seals are used to seal the pistons. Guided piston relief valves are generally used for pressures below 800 psi, although they can be made with heavier springs for higher pressures. The heavier springs give the valve a greater differential and consequently increase the size of the valve. Differential piston design — A variation of the guided piston relief valve is the differential piston relief valve. It is so called because the pressure acts on an annular area which is the difference between two piston diameters. The annular area is smaller than the valve’s seat area. This permits using a smaller spring than would be needed if pressure acted on the entire seat area. They have a lower pressure differential than poppet or guided piston relief valves. Pilot-operated reliefs — For applications requiring valves that must re-
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P R E S S U R E - C O N T R O L VA LV E S
B is the same as system pressure and is less than the setting of the pilot valve spring. Direct-acting relief valve As system pressure rises, the Pilot-operated relief valve pressure in passage B rises as well, and, when it reaches the setting of the pilot valve, the pilot valve opens. Oil is released behind the main valve through passage B through the drain port. The resulting pressure drop across orifice A in the main relief valve opens it and excess oil flows to Cracking pressure tank, preventing any further rise in 0 inlet pressure. The valves close 100 Pressure, % again when inlet oil pressure Fig. 5. Comparison of action of relief valves at drops below the valve setting. Pilot-operated relief valves have cracking and full flow pressure. less pressure override than directlieve large flows with small pressure acting relief valves, such as in Figure 2. differential, pilot-operated relief valves Because these valves do not start are often used, Figure 3. The pilot-oper- opening until the system reaches 90% of ated relief valve operates in two stages. full pressure, the efficiency of the sysA pilot stage which consists of a small, tem is protected because less oil is respring-biased relief valve (generally leased. These valves are best suited for built into the main relief valve) acts as a high pressure, high volume applications. trigger to control the main relief valve. Although their operation is slower than However, the pilot may also be located that of direct-acting relief valves, pilotremotely and connected to the main operated relief valves maintain a system valve with pipe or tubing. at a more constant pressure while relievThe main relief valve is normally ing. The operating characteristics of diclosed when the pressure of the inlet is rect acting and pilot-operated relief below the setting of the main valve valves are in Figure 5. spring. An orifice, A, in the main valve, Figure 3, permits system pressure fluid Pressure reducing valves to act on a larger area on the spring side The most practical components for of the poppet so that the sum of this maintaining secondary, lower pressure force and that of the main spring keep in a hydraulic system are pressure rethe poppet seated. At this time, the pilot ducing valves. valve is also closed. Pressure in passage Pressure reducing valves are nor100
Flow capacity, %
Full-flow pressure
mally-open, 2-way valves which receive their actuating signals from downstream fluid pressure to close. There are two types: direct-acting and pilot-operated. Direct-acting — A pressure reducing valve limits the maximum pressure available in the secondary circuit regardless of pressure changes in the main circuit and as long as the work load generates no back flow into the reducing valve port in which case the valve will close, Figure 4. The pressure sensing signal comes from the downstream side (secondary circuit). This valve, in effect, operates in reverse fashion from a relief valve (which senses pressure from the inlet) and is normally closed. As pressure rises in the secondary circuit, Figure 4, hydraulic force acts on area A of the valve, closing it partly. Spring force opposes the hydraulic force, so that only enough oil flows past the valve to supply the secondary circuit at the desired pressure. The spring setting is adjustable. When outlet pressure reaches that of the valve setting, the valve closes except for a small quantity of oil that bleeds from the low-pressure side of the valve, usually through an orifice in the spool, through the spring chamber, to reservoir. Should the valve close fully, leakage past the spool could cause pressure build-up in the secondary circuit. To avoid this, a bleed passage to reservoir keeps it slightly open, preventing a rise in downstream pressure above the valve To primary circuit
Pressure sensing passage
Pressure sensing passage
Reduced pressure (secondary circuit)
Fig. 6. Pilot-operated, pressure reducing valve has reduced pressure on both ends of the spool. A light spring holds the spool open. When reduced pressure increases to the spring setting on the pilot valve, the pilot valve opens.
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To secondary circuit
Fig. 7. Sequence valve is normally-closed, 2-way valve. It is held closed by an adjustable spring. Pressure at the inlet port acting on the left of the spool opens the valve against the force of the spring. The spring chamber drains to tank.
P R E S S U R E - C O N T R O L VA LV E S
External sensing line
Fig. 8. Counterbalance valve stops flow from its inlet port to its outlet port until pressure at the inlet port overcomes adjusting spring force. An integral check valve permits free flow through the valve in the opposite direction.
setting. The drain passage returns leakage flow to reservoir. Valves with built-in relieving capability also are available to eliminate the need for this orifice. Constant and fixed pressure reduction — Constant pressure reducing valves supply a preset pressure regardless of main circuit pressure as long as pressure in the main circuit is higher than that in the secondary. Here is how this valve works: it balances secondary circuit pressure against the force exerted by an adjustable spring which wants to open the valve. When pressure in the secondary circuit drops, spring force opens the valve enough to increase pressure and keep a constant reduced pressure in the secondary circuit. Fixed pressure reducing valves supply a fixed amount of pressure reductions regardless of the pressure in the main circuit. For instance, assume a valve is set to provide reduction of 250 psi. If main system pressure is 2750 psi, reduced pressure will be 2500 psi; if main pressure is 2000 psi, reduced pressure will be 1750 psi. This valve operates by balancing the force exerted by the pressure in the main circuit against the sum of the forces exerted by secondary circuit pressure and the spring. Because the pressurized areas on both sides of the poppet are equal, the fixed reduction is that exerted by the spring. Pilot-operated pressure reducing valves — The spool in a pilot-operated, pressure-reducing valve is balanced hy-
Fig. 9. Unloading valve is spring-loaded to the closed position. When system pressure, transmitted to the valve though the pilot port, is sufficient to overcome force of the adjustable spring, the valve opens. Pump delivery unloads to the reservoir at low pressure.
draulically by downstream pressure at both ends, Figure 6. A light spring holds the valve open. A small pilot relief valve, usually built into the main valve body, relieves fluid to tank when reduced pressure reaches the pilot valve’s spring setting. This fluid flow causes a pressure drop across the spool. Pressure differential then shifts the spool toward its closed position against the light spring force. The pilot valve relieves only enough fluid to position the main valve spool or poppet so that flow through the main valve equals the flow requirements of the reduced pressure circuit. If no flow is required in the low pressure circuit during a portion of the cycle, the main valve closes. Leakage of high pressure fluid into the reduced pressure section of the valve then returns to the reservoir though the pilot operated relief valve. Pilot-operated pressure reducing valves generally have a wider range of spring adjustment than direct-acting valves. They generally provide more repetitive accuracy. However, oil contamination can block flow to the pilot valve and the main valve will fail to close properly. Pilot-operated valves with built-in reduced pressure system relieving capability also are available. Sequence valves In circuits with more than one actuator, it is often necessary to move the actuators, such as cylinders, in a definite order or sequence. One way to do this is
with limit switches, timers, or other electrical controls. Sometimes, this result can also be achieved by sizing cylinders according to the load they must displace. The cylinder requiring the least pressure to move its load extends first. At the end of its stroke, system pressure increases and extends the second cylinder. This continues until all cylinders are actuated. However, in many installations, space limitations and force requirements determine the size cylinder needed to do the job. In this case, sequence valves can be used to actuate the cylinders in the required order. Sequence valves are normallyclosed, 2-way valves. They regulate the sequence in which various functions in a circuit occur, Figure 7. They resemble direct-acting relief valves except that their spring chambers are generally drained externally to reservoir, instead of internally to the outlet port as in a relief valve. Normally, a sequence valve permits pressure fluid to flow to a second function only after an earlier, priority function has been completed and satisfied. When normally closed, a sequence valve allows fluid to flow freely to the primary circuit, to perform its first function until the pressure setting of the valve is reached. When the primary function is satisfied, pressure in the primary circuit rises and is sensed in pressure-sensing passage A. This pressurizes the spool and overcomes the force exerted by the
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P R E S S U R E - C O N T R O L VA LV E S
To system
To system
From pump
From pump Control piston
Fig. 10. Unloading valve for accumulator circuit opens at a set unloading pressure and closes at a lower pressure. The valve opens when the system reaches a pressure determined by the adjustable spring and pump pressure on the right of the control piston. The pump unloads to tank and relieves pressure on the right of the control spool. The valve closes at a lower pressure because force from system pressure on the left of the control spool must only overcome force of the adjusting spring.
Fig. 11. Piloted unloading valve has piston with pump pressure at both ends. The valve is initially held closed by a light spring. When system pressure on the pilot piston overcomes force of the pilot spring the pilot valve opens. Fluid from behind the main valve piston drains to tank, causing a pressure difference across the piston which opens the valve. When system pressure fails, the pilot valve closes, restoring equal pressure to both ends of main valve spool. The light spring closes the valve.
spring. The spring is compressed, the valve spool shifts, and oil flows to the secondary circuit. Sequence valves sometimes have check valves which permit reverse flow from the secondary to the primary circuit. However, sequencing action is provided only when the flow is from the primary to the secondary circuit. In some applications, it is desirable to provide an interlock so that sequencing does not occur until the primary actuator reaches a certain position. This is done with remote operations.
allow the cylinder to retract freely. If it is necessary to relieve back pressure at the cylinder, and increase the force at the bottom of the stroke, the counterbalance valve can be operated remotely. Counterbalance valves are usually drained internally. When the cylinder extends, the valve must open and its secondary port is connected to reservoir. When the cylinder retracts, it matters little that load pressure is felt in the drain passage because the check valve bypasses the valve’s spool.
Counterbalance valves These normally closed valves are primarily used to maintain a set pressure in part of a circuit, usually to counterbalance a weight or external force or counteract a weight such as a platen or a press and keep it from falling. The valve’s primary port is connect to the cylinder’s rod end, the secondary port to the directional control valve, Figure 8. The pressure setting is slightly higher than that required to keep the load from falling. When pressure fluid flows to the cylinder’s cap end, the cylinder extends, increasing pressure in the rod end, and shifting the main spool in the counterbalance valve. This creates a path which permits fluid to flow through the secondary port to the directional control valve and to reservoir. As the load is raised, the integral check valve opens to
Unloading valves These valves are normally used to unload pumps. They direct pump output flow (often the output of one of the pumps in a multi-pump system) directly to reservoir at low pressure, after system pressure has been reached. The force exerted by the spring keeps the valve closed, Figure 9. When an external pilot signal acting on the opposite end of the valve spool exerts a force large enough to exceed that exerted by the spring, the valve spool shifts, diverting pump output to reservoir at low pressure. High-low circuits which use two pumps for traverse and speed, or clamping, depend on unloading valves to improve efficiency. Output from both pumps is needed only for fast traverse. During feed or clamping, output
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from the large pump is unloaded to reservoir at low pressure. Unloading valves for accumulator circuits — An unloading valve can be used in an accumulator circuit to unload the pump after the accumulator has been charged, Figure 10. The valve remains closed while the pump is charging the accumulator. After the accumulator is charged the unloading valve opens, unloading the pump at low pressure while the accumulator supplies pressure fluid to the system. Every time pressure in the accumulator drops below a preset level (controlled by the setting of the spring) the charge/unload cycle repeats. Piloted unloading valves — Unloading valves are also made with a pilot to control the main valve, Figure 11. A port through the main valve plunger allows system pressure to act on both ends of the plunger. A light spring plus system pressure acting on the larger area at the spring end of the plunger holds the valve closed. A built-in check valve maintains system pressure. When system pressure drops to a preset value, the pilot valve closes. Pump flows through the port in the main valve spool closes the valve. In most pump unloading valve for accumulator circuits, only the opening pressure is adjustable and the closing pressure is a fixed percentage of it. However, a pilot-operated unloading valve can adjust both pressures.
Pressure gages & flow meters T
he majority of gages for measuring pressure have one characteristic in common: the pressure being measured is the only source of energy required to provide a visual indication of static pressure. Some form of elastic chamber inside the gage case converts the pressure to motion, which is translated through suitable links, levers, and gearing into movement of a pointer across an indicating scale. Three types of elastic chambers are commonly used in gages for fluid power systems: ● C-shaped, spiral, and helical Bourdon tubes ● bellows, and ● single- and multi-capsule stacks. Bourdon-tube designs Since the invention of the Bourdontube gage more than a century ago, pressure gage manufacturers have been developing different types of gages to meet specific needs without ever changing the basic principle of the Bourdon tube’s operation. Bourdontube gages, Figure 1, are now commonly available to measure a wide range of gage, absolute, sealed, and differential pressures, plus vacuum. They are manufactured to an accuracy as high as 0.1% of span and in dial diameters from 11/2 to 16 in. A variety of accessories can extend their performance and usefulness. For example, snubbers and gage isolators can be installed to protect the sensitive internal workings of the gage from pressure spikes. The availability of Bourdon-tube pressure gages to meet specific needs, coupled with their inherent ruggedness, simplicity, and low cost has resulted in their wide use in many applications. Gages using C-shaped Bourdon tubes as the elastic chamber — the type shown in Figure 1 — are by far the
ment then transmits this tip motion to a gear train that rotates an indicating pointer over a graduated scale to display the applied pressure. Often, a movement is incorporated to provide mechanical advantage to multiply the relatively short movement of the tube tip. Spiral and helical Bourdons Bourdon tubes also may be made in
A
A
Pressure
Section A-A
Fig. 1. Cut-away view of C-shaped Bourdon-tube pressure gage. Pressure-induced strain in the Bourdon tube causes it to deform. Transmitting this deformation to a pointer through a movement linkage provides a visual indication of pressure.
Fig. 2. Simplified view of spiral Bourdontube pressure gage and movement.
21/4-turn helical Bourdon
most common. Pressurized fluid enters the stem at the bottom (which is sometimes center-back-mounted instead) and passes into the Bourdon tube. The tube has a flattened cross section and is sealed at its tip. Any pressure in the tube in excess of the external pressure (usually atmospheric) causes the Bourdon tube to elastically change its shape to a more circular cross section. This change in shape of the cross section tends to straighten the C-shape of the Bourdon tube. With the bottom stem end fixed, the straightening causes the tip at the opposite end to move a short distance — 1/16 to 1/2 in., depending on the size of the tube. A mechanical move-
Pinion gear
Fig. 3. Simplified view of helical Bourdontube pressure gage and movement.
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PRESSURE GAGES & FLOW METERS
the form of a spiral, Figure 2, or a helix, Figure 3. Each uses a long length of flattened tubing to provide increased tip travel. This does not change the operating principle of the Bourdon tube, but produces tip motion equal to the sum of the individual motions that would result from each part of the spiral or helix considered as a Cshape. Small-diameter spirals and helices can be manufactured to provide enough motion to drive the indicating pointer directly through an arc up to 270° without having to use a multiplying movement. Alternatively, they may be manufactured to be used in conjunction with a multiplying movement. In this case, the required motion is distributed over several turns, resulting in lower stress in the Bourdon material. This improves fatigue life when Link to pointer
Bellows
Coil spring Pressure
Sealed shell
Fig. 4. Cross-sectional view of springloaded bellows pressure gage.
compared to a C-shaped Bourdon tube in the same pressure range. Bellows and diaphragms Low-pressure applications do not generate enough force in the Bourdon tube to operate the multiplying mechanism; therefore, Bourdon-tube gages are not generally used for pressure spans under 12 psi. For these ranges, some other form of elastic chamber must be used, a metallic bellows, Figure 4, for example. These bellows generally are made by forming thinwall tubing. However, to obtain a reasonable fatigue life and motion that is more linear with pressure, a coil spring supplements the inherent spring rate of the bellows. These spring-loaded bellows gages generally are used in pressure ranges having spans to 100 psi and to 1 in. Hg. Metallic diaphragms also are used as the elastic chamber in low-pressure gages. A diaphragm plate is formed from thin sheet metal into a shallow cup having concentric corrugations. To make an element with a low spring rate that generates substantial deflection from a small change in pressure, two plates can be soft soldered, brazed, or welded at their periphery to form a capsule, and additional capsules can be joined at their centers to form a stack, Figure 5. Generally, the measured pressure is applied to the interior of the element and no supplemental coil springs are used. A 2-in. diameter capsule (two plates) will provide about 0.060 in. of
Calibrated scale
Capsules
Pressure
Fig. 5. Cross-sectional view of metallic diaphragm pressure gage with stacked capsules.
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motion without exceeding the elastic limit of the material. This is usually enough to operate a high-ratio multiplying movement because diaphragm deflection can transmit high force. Diaphragm elements often are used in gages to indicate absolute pressure. In this form, the diaphragm element is evacuated. sealed, and mounted within a closed chamber. The pressure to be measured is admitted to the closed chamber and surrounds the diaphragm element. Changes in the measured pressure cause the element to deflect, but because atmospheric pressure is excluded and has no effect on the indication, the gage may be calibrated in terms of absolute pressure. If the applied pressure is atmospheric pressure, the gage is known as a barometer. Diaphragm elements also may be used in an opposing arrangement. By evacuating one side of the assembly, the gage can indicate absolute pressure. If a pressure is applied to one side of the assembly, and a second pressure is applied to the other side, then the differential pressure will be indicated. The differential pressure is limited with respect to the static pressure that can be applied. That is, the gage may be suitable to indicate between 10 psi and 12 psi, but not be suitable to indicate between 100 psi and 102 psi. Also, the consequence of inadvertently applying full pressure to one side of the element and no pressure to the other side of the element must be considered. Selection Specifying a pressure gage involves a number of considerations: ● connection size — nominal size of the port or fitting into which the gage will be threaded, male or female, and thread size ● mounting configuration — bottom or back-center stem mounted or panel mounted ● dial size — large enough to be seen clearly from a distance but small enough to prevent taking up excessive space ● units of measure — determine whether the dial should be calibrated in psi, bar, kPa, etc. Many manufacturers offer gages with dual-dimensioned scales ● materials of construction — gages may have a glass or plastic crystal, metal
PRESSURE GAGES & FLOW METERS
or plastic case, and usually a brass connection. Ensure that materials are compatible with the environment and fluid ● dry or liquid filled — liquid-filled gages generally contain glycerin to dampen effects of shock and vibration, and provide continuous lubrication of the movement to extend life, and ● pressure range — as a rule of thumb, select a gage with a maximum pressure reading twice that of the anticipated measured pressure. This provides a safety margin to prevent temporary high-pressure pulsations or spikes from damaging the gage. Options and accessories A variety of options and accessories are available to enhance life and operation of gages. Digital readout is accomplished by mounting a strain gage to the sensing element and using on-board electronics to convert the strain induced by pressure into digital readout on an LED or LCD panel. Digital gages require a power source — generally a long-life battery — and may use a switch so that power is consumed only when a button is pushed to read the pressure. A gage isolator, mounted between the gage and circuit, prevents the gage from being exposed to fluid pressure unless a button is pushed. In this manner, the gage is not exposed to pressure spikes and pulsations unless they occur when pressure is being read. Orifices or snubbers protect gages by smoothing out pressure fluctuations seen by the gage. Snubbers may cause gages to respond sluggishly, but can extend life by damping rapid pressure fluctuations. To help protect the gage from external physical shock, case protectors can be used, which encapsulate the gage in rubber. A wide variety of other useful options — such as an integral adjustable pressure switch — are available from manufacturers to make pressure gages even more versatile. Flow meters Unlike pressure gages, which have been permanently mounted on the vast majority of hydraulic and pneumatic systems for decades, flow meters continue to be used primarily for testing to assess the performance of a system, Figure 6. Systems requiring continuous
monitoring of flow generally use electronic flow sensors rather than flow meters, which require no power. Electronic flow sensors use a variety of sensing elements (turbines, positivedisplacement chambers, differentialpressure measurement, etc.) to generate an electronic signal proportional to or otherwise representative of flow. This signal is then routed to an electronic display panel or control circuit. However, flow sensors produce no visual indication of flow by themselves, and they need some source of external power to transmit a signal to an analog or digital display. Self-contained flow meters, on the other hand, rely on the dynamics of flow to provide a visual indication of flow. Although design details differ from one manufacturer to another, flow meters operate on the principle of dynamic pressure. The main components are a tapered shaft and spring-loaded piston, Figure 7. With no fluid flow, the actuating spring pushes the piston to its left-most position. As fluid enters from the left side, pressure acts against the spring and builds to open the orifice formed between the ID of the piston and OD of the tapered shaft by pushing the piston to the right. As the piston is pushed farther to the right, the orifice area increases because the effective area of the tapered shaft decreases. Eventually, the orifice area will be large enough so that dynamic pressure from flow equals the opposing spring force. The position of the piston in equilibrium, then, provides an indication of flow. For some applications, flow can be measured directly by comparing the pis-
Piston
Fig. 6. Flow meters, unlike pressure gages, generally are not permanently installed in hydraulic or pneumatic equipment and must be piped temporarily in series with the circuit before use.
ton position to a calibrated scale marked on the flow meter’s transparent outer case. For most hydraulic applications, however, the piston usually has a magnet embedded that moves a follower collar. Position of the collar can then be compared to a calibrated scale. Because flow indication depends on fluid dynamics, changes in a fluid’s physical properties can affect readings. This is because a flow meter is calibrated to a fluid having a certain specific gravity within a range of viscosities. A wide deviation in temperature can change a hydraulic fluid’s specific gravity and viscosity, so if a flow meter is used when the fluid is very hot or very cold, flow readings may not conform to manufacturers’ specifications. However, because most equipment is tested under operating conditions, readings generally should fall within manufacturers’ specifications for accuracy.
Tapered shaft
Spring
Fig. 7. Cut-away view shows internal components of flow meter that provide visual indication of flow. Fluid entering from left side of meter passes through variable orifice formed between OD of tapered shaft and ID of spring-loaded piston. Dynamic flow pushes the piston to the right until orifice becomes large enough to accommodate flow.
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Pressure switches T
he pressure switch probably is the most common electrical interface device found in fluid power systems. Pressure is inherently present in every system, and solenoidoperated valves remain the most popular in use. Pressure switches act as a low cost alarm, to protect operators, machinery, and work in progress by turning equipment off if pressure becomes too high or too low. These components also play important roles in automation circuits by signaling when sequence steps are completed or noting if they are skipped. The action of a pressure switch is simple. When pressure at the pressure port generates sufficient force against the sensing element to overcome an opposing spring; the element moves a very short distance to actuate an electrical switching element, then is stopped mechanically. When pressure decays, everything moves back to its original position. The designer must realize that these components are subjected to the same operating hazards as any other fluid power device: pressure peaks and pulsations, mechanical vibration, thermal extremes and shocks, and contaminants in the pressure media. Depending on the application, pressure switches may have to perform in a high-frequency duty cycle, or be at rest for long periods and then perform reliably. Careful specification to match the switch to its application is essential for accurate, reliable operation. Construction The basic parts of a typical pressure switch are the pressure-sensing element, the opposing spring, and electrical switching element, and its exterior electrical connection. The three types of sensing elements found most frequently are diaphragms, bourdon tubes, and pistons.
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A switch’s diaphragm may be made of metal, elastomer, or fabric-reinforced elastomer. Usually used in low-pressure circuits (around 700 psi maximum), accuracy and repeatability of diaphragmtype pressure switches is good. By their nature, diaphragms virtually eliminate internal leakage. Potential problems: pressure surges could physically deform a diaphragm, and metal diaphragms can fatigue in high-cycle service. Fatigue may also limit service life. Some models combine diaphragms with pistons to minimize these problems. Bourdon tubes operate under the same principle used in pressure gages. Pressure fluid, introduced into a curved tube with one end fixed, tends to straighten the tube and the opposite free end moves to actuate the electrical switching element. Bourdon tube pressure switches accept higher pressures than diaphragm types and also do not leak. Accuracy is high. Potential problems: excessive machine vibration may move the tube’s free end, causing false signals; and fatigue failure from high cycling. Again, fatigue can limit service life. Piston-type pressure sensors work like rams in hydraulic cylinders. By their nature, they can accommodate the highest pressures. Piston-type pressure switches tolerate pressure spikes and vibration well. Fatigue is not a factor in the piston’s heavier parts. Leakage can eventually become a problem as piston seals wear, and silting from fine, solid contaminants may impede piston movement in some systems — particularly if the piston moves infrequently. The switch may be equipped with special snubbing devices, Figure 1, to mitigate the destructive forces of rapid pressure rises. Performance criteria Setpoint is the term for the desired pressure level for an alarm or control
Pressure switches accept pressure signals and, at a preselected pressure, open or close internal electrical switching elements. They may respond to rising or falling pressure, or a pressure differential. function to occur. When pressure rises or falls in the opposite direction from which the set point was approached, the switching element will deactivate. Often, most pressure switches will include an adjustment to allow the user to change the setpoint while in the field. Some
Figure 1. Pressure snubber in pressure port mitigates spikes and pump ripple.
Figure 2. Watertight enclosure.
manufacturers offer models with two setpoints that can be used for both high and low alarm applications. A pressure switch will actuate at a given setpoint pressure as system pressure increases, but will deactuate at a lower pressure as system pressure decreases. The difference is known as deadband. Deadband is the result of
PRESSURE SWITCHES
inefficiency of the mechanical parts of the switch. Some deadband often is desired, as in the case of a compressor control, where the setpoint might be used to turn the compressor off at a 60 psi increasing setpoint and to turn it back on at a reset point of 35 psi. The 25-psi difference in this example is the deadband. Several models of pressure switches include deadband adjustment for such control applications, but the majority of pressure alarms and shutdowns are satisfied by fixed, nonadjustable deadband models. Repeatability is the measure of accuracy of a pressure switch. This is defined as the closeness of agreement among a number of consecutive measurements of the output setpoint for the same value of the input, under the same operating condition, for fullrange traverses. This is normally expressed as a percentage of range (or span) of the switch and does not include deadband. Most models with diaphragm-sealed piston actuators feature ±1% of range repeatability. The enclosure Many applications require that the pressure switch include an electrical enclosure, with provisions for the user to terminate the wires conveniently and safely. When installed properly, the internals will be protected from the environment, and personnel will be protected from voltage on the electrical terminals. The enclosure also has provisions for mounting. Most industrial applications are satisfied with a watertight enclosure, Figure 2. These enclosures are provided with a gasket and threaded electricalconduit opening that also will be dusttight. Explosion-proof enclosures, Figure 3, are required on applications that are in hazardous locations or on pumps, compressors, turbines, lube systems, etc., that are intended for use in hazardous locations. Electrical considerations Snap-action micro switches used in pressure switches are isolated from system pressure and contaminants, and they are mechanically protected from overtravel, so their service life probably will exceed that of the sensing ele-
Figure 3. Explosion-proof enclosure.
ment. Available designs include singlepole, double-throw (SPDT), and double-pole, double-throw (DPDT). Standard wiring arrangements allow the designer to select whether the switch will function as normally open or normally closed. Manufacturers also offer a variety of standard electrical connectors to match other connectors in the electrical circuits to simplify installation and replacement. Specification considerations To match a pressure switch to a particular application, consider these factors: ● voltage and current requirements of electrical circuit ● normally open or normally closed ● operating pressure ● proof and peak pressures ● set point adjustability — and if required, adjustment range ● repeatability of setpoint ● deadband magnitude ● deadband adjustability ● fluid compatibility ● duty cycle ● speed of cycling ● vibration potential ● fluid, ambient temperature conditions ● mounting requirements ● port size and thread, and
Fig. 4. Solid state pressure switch.
● ambient conditions that could affect housing material. Electronic switches Electronic (or solid state) pressure switches, Figure 4, are now available in several configurations from most major pressure switch suppliers. In these products, the mechanical actuator is replaced by a pressure transducer, and the setpoint adjustment is done electronically. Higher setpoints, improved repeatability, longer cycle life, excellent temperature stability, and easy setpoint and deadband adjustments result from the lack of moving parts. Electronic switches can be engineered to survive cycle rates in excess of 50 million cycles. Designers can choose between metal or ceramic diaphragms. Cost is usually higher than with the mechanical equivalent, but the differential is becoming smaller as the costs of the mechanical parts increase and those of the electronic parts decrease. Changing roles Today, the pressure switch not only functions as a simple alarm. In conjunction with other devices such as transducers, the pressure switch can be a key interface device in system health monitoring and/or servo applications.
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Reservoirs
When it comes to reservoir design, bigger is not necessarily better. In fact, the trend is to provide smaller reservoirs.
I
n addition to holding in reserve enough fluid to supply a hydraulic system’s varying needs, a reservoir, Figure 1, provides: ● a large surface area to transfer heat from the fluid to the surrounding environment ● enough volume to let returning fluid slow down from a high entrance velocity. This lets heavier contaminants settle and entrained air escape ● a physical barrier (baffle) that separates fluid entering the reservoir from fluid entering the pump suction line ● air space above the fluid to accept air that bubbles out of the fluid ● access to remove used fluid and contaminants from the system and to add new fluid ● space for hot-fluid expansion, gravity drain-back from a system during shutdown, and storage of large volumes needed intermittently during peak periods of an operating cycle, and ● a convenient surface to mount other system components, if practical. These are the traditional roles of reservoirs; new trends may present deviations from the norm. For example, new designs for hydraulic systems often call for reservoirs that are much smaller than those based on traditional rules of thumb. Because most systems warrant some special consideration, it is important to consult industry standards for minimum guidelines. Recommended Practice NFPA/T3.16.2* ad* The industry standard for hydraulic reservoirs is contained in NFPA/T3.16.2 R1-1996 (pending approval as ANSI/B93.18), which is published by the National Fluid Power Assn. To order a copy, or for more information, contact NFPA, 3333 North Mayfair Rd., Milwaukee, WI 53222-3219; phone 414/778-3344, fax 414/778-3361, or email
[email protected].
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Return line
Clean-out plate (both ends)
Drain return
Pump inlet line (option)
Air breather and filler
Sealed flange
Baffle plate Strainer Drain plug
Thermometer and sightglass
End plate extension Drain plug
Fig. 1. Cutaway illustrates key features of traditional rectangular reservoir. Baffle separates returning fluid from that being drawn into pump.
dresses basic minimum design and construction features for reservoirs. Reservoir sizing Although the considerations just discussed may be important, the first variable to resolve is, indeed, reservoir volume. A rule of thumb for sizing a hydraulic reservoir suggests that its volume should equal three times the rated output of the system’s fixed-displacement pump or mean flow rate of its variable-displacement pump. This means a system using a 5-gpm pump should have a 15-gal reservoir. The rule suggests an adequate volume to allow the fluid to rest between work cycles for heat dissipation, contaminant settling, and deaeration. Keep in mind that this is only a rule of thumb for initial sizing. In fact, NFPA’s Recommended Practice states, “Previously, three times the pump capacity had been recommended. Due to today’s system technology, design objectives have changed for economic reasons, such as space saving, minimizing oil usage, and overall system cost reductions.” Whether or not you choose to adhere to the traditional rule of thumb or follow
the trend toward smaller reservoirs, be aware of parameters that may influence the reservoir size required. For example, some circuit components — such as large accumulators or cylinders — may involve large volumes of fluid. Therefore, a larger reservoir may have to be specified so fluid level does not drop below the pump inlet regardless of pump flow. Systems exposed to high ambient temperatures require a larger reservoir unless they incorporate a heat exchanger. Be sure to consider the substantial heat that can be generated within a hydraulic system. This heat is generated when the hydraulic system produces more power than is consumed by the load. A system operating for significant periods with pressurized fluid passing over a relief valve is a common example. Reservoir size, therefore, often is determined primarily by the combination of highest fluid temperature and highest ambient temperature. All else being equal, the smaller the temperature difference between the two, the larger the surface area (and, therefore, volume) required to dissipate heat from fluid to the surrounding environment. Of course, if
RESERVOIRS
ambient temperature exceeds fluid temperature, a water-cooled or remotemounted heat exchanger will be needed to cool the fluid. In fact, for applications where space conservation is important, heat exchangers can reduce reservoir size (and cost) dramatically. Keep in mind that the reservoir may not be full at all times, so it may not be dissipating heat through its full surface area. The reservoir should contain additional space equal to at least 10% of its fluid capacity. This allows for thermal expansion of the fluid and gravity drainback during shutdown, yet still provides a free fluid surface for deaeration. In any event, NFPA/T3.16.2 requires that maximum fluid capacity of the reservoir be marked permanently on its top plate. As the NFPA document implies, a trend toward specifying smaller reservoir has emerged as a means of reaping economic benefits. A smaller reservoir is lighter, more compact, and less expensive to manufacture and maintain than one of traditional size. Moreover, a smaller reservoir reduces the total amount of fluid that can leak from a system — important from an environmental standpoint. But specifying a smaller reservoir for a system must be accompanied by modifications that compensate for the lower volume of fluid contained in the reservoir. For example, because a smaller reservoir has less surface area for heat transfer, a heat exchanger may be necessary to maintain fluid temperature within requirements. Also, contaminants will not have as great an opportunity for settling, so high-capacity filters will be required to trap contaminants that would otherwise settle in the sump of the reservoir. Perhaps the greatest challenge to using a smaller reservoir lies with removing air from the fluid. A traditional reservoir provides the opportunity for air to escape from fluid before it is drawn into the pump inlet. Providing too small a reservoir could allow aerated fluid to be drawn into the pump. This could cause cavitation and eventual damage or failure of the pump. When specifying a small reservoir, consider installing a flow diffuser, which reduces the velocity of return fluid (typically to 1 ft/sec), helps prevent foaming and agitation, and reduces potential pump cavitation from flow disturbances at the inlet. Another technique is
Fig. 2. This modular power unit demonstrates a trend in design: mounting the electric motor vertically with the pump submerged in hydraulic fluid. This technique reduces leakage, noise, and floor space required.
to install a screen at an angle in the reservoir. The screen collects small bubbles, which join with others to form large bubbles that readily rise to the fluid’s surface. Perhaps the best way to prevent aerated fluid from being drawn into the pump is to prevent aeration of fluid in the first place by paying careful attention to fluid flow paths, velocities, and pressures when designing the hydraulic system. Design configurations Traditionally, the pump, electric motor, and other components of a hydraulic power unit mount on top of a rectangular reservoir. The reservoir top, therefore, must be structurally rigid enough to support these components, maintain alignments, and minimize vibration. An auxiliary plate may be mounted on the reservoir’s top to meet these objectives. A big advantage of this configuration is that it allows easy access to the pump, motor, and accessories. A current design trend has the electric motor mounted vertically, with the pump submerged in hydraulic fluid, Figure 2. This conserves space, because the reservoir can be made deeper and take up less floor space than one with traditional “bathtub” proportions. The submergedpump design also eliminates external
pump leakage, because any fluid leaking from the pump flows directly into the reservoir. In addition, the power unit is quieter, because the hydraulic fluid tends to damp pump noise. An alternate configuration positions the reservoir above the pump and motor, Figure 3. This overhead configuration provides the advantage of combining atmospheric pressure and the weight of the fluid column to flood (force fluid into) the pump inlet, which helps prevent cavitation. The reservoir’s top cover can be removed to service internal components without disturbing the pump and motor. The overhead reservoir may cause a problem with gravity-return drain lines, so an auxiliary pump may be needed to route fluid up to the reservoir. When noise is a problem, overhead tanks provide the most convenient way to enclose the pump and electric motor within a noise suppression chamber. Many applications use reservoirs that combine characteristics of the different configurations. For example, an Lshaped reservoir, Figure 4, combines the advantages of top- and base-mounted reservoirs — a flooded pump inlet and easy accessibility of components. Reservoirs can also be pressurized to flood the pump. This pressure can come from an external source or from trapped air and fluid thermal expansion. A pressure-control valve allows filtered air to enter the reservoir when the fluid cools but prevents its release unless air inside reaches a threshold pressure. Shape and construction There is no standard reservoir shape. Geometrically, a square or a rectangular prism has the largest heat-transfer surface per unit volume. A cylindrical shape, on the other hand, may be more economical to fabricate. If the reservoir is shallow, wide, and long, it may take up more floor space than necessary and does not take full advantage of the heattransfer surface of the walls. Theoretically, because heat rises, the reservoir top holds the greatest potential for heat transfer to the atmosphere. However, in particularly dirty environments, contaminants often collect on the reservoir top and act as insulation. This reduces the effective heat transfer from the top of the reservoir, so reservoir sides could actually be the most ef-
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RESERVOIRS
fective heat transfer area in some instances. On the other hand, a tall and narrow geometry conserves floor space
and provides a large surface area for heat transfer from the sides. Depending on the application, however, this shape
may not provide enough area at the top surface of the fluid to let air escape. The reservoir should be strong and
Reservoirs for mobile equipment
M
obile hydraulic reservoirs are expected to perform the same functions as their industrial counterparts — but usually under more adverse and less predictable operating conditions. Machine motion (which makes complex baffling systems necessary to prevent fluid sloshing) and extreme ambient temperatures are just two examples of the special problems designers of hydraulic systems for mobile equipment face. Size and weight limitations may require mobile equipment to operate with reservoirs as small as the volume a pump discharges in a minute. This is roughly a third the size of a reservoir traditionally used in an industrial application. The space and shape limitations mobile equipment places on reservoirs requires that they often be custom designed. Cost, size, and weight must be minimized, while still maintaining adequate performance and efficiency. Internal or external filters? Return filters are often placed inside the tank to save space and to provide integral diffusion. One advantage of in-tank return filtration is that filling the tank through the filter helps ensure system cleanliness. However, be sure contaminants cannot fall into the reservoir when a return filter element is changed. Placing filters within the tank provides a neat design but may promote contaminating an area that is difficult to keep clean. While more difficult to plumb, external return filters keep contamination outside the tank, and they are more easily accessible for servicing. Magnets should be placed in the reservoir to trap ferrous particles. Dams and suction strainers also can be added to increase the effectiveness of the reservoir as a contaminant controller. Particle dams, placed between the return and suction areas of the tank, help contain heavier particles that may have bypassed the return filters. Dams commonly consist of an angle plate that extends across the floor of the tank. The dam should be high enough to contain particles until the reservoir is rou-
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tinely cleaned but low enough to prevent fluid from having to cascade over it. Dams also provide ideal mounting surfaces for magnets. Locating a pump at or above fluid level and far away from the tank (more the rule than the exception with mobile equipment) usually prohibits the use of pump inlet filters. Suction strainers or filters should be considered as a form of last-chance pump protection when positive pump inlet conditions can be provided — as with a charge pump or pressurized reservoir. Pay attention to fluid temperature (especially during
Reservoirs for mobile equipment often use a dipstick to check fluid level because sight gages, though preferred, might be inaccessible or subject to damage.
startup) when sizing suction filters if equipment will operate in cold climates and pumps cannot be disengaged during startup. Vented or pressurized reservoir? An important design consideration is whether to specify a vented or pressurized reservoir. The major deciding factors are the location and inlet requirements of the pumps. The fluid level of the reservoir in many mobile applications is below the pump inlet. At best, if there is vacuum at the pump inlet, the pump may have to be derated. If inlet line losses are great enough, cavitation will occur. In these cases, pressurizing the reservoir will help maintain pump performance. There are three ways to pressurize a reservoir on most mobile equipment: 1. Use regulated compressed air from a machine’s pneumatic system — the most
effective method — if available. 2. Trap the air within the reservoir clearance volume (above the fluid) and depend on thermal expansion of the fluid to compress this air, and thus pressurize the reservoir. A reservoir pressure cap holds pressure within the tank and relieves excess pressure. 3. Tap pressurized air from the scavenge pump of a two-cycle diesel engine. With pressurized reservoirs, consideration must be given to calculate stresses on reservoir walls, because even low pressures can exert substantial loads. For example, an internal pressure of only 3 psi applies a force of 1800 lb on a 20- 2 30-in. wall. This force, combined with weight of hydraulic fluid, plus G forces involved in mobile equipment, can produce stresses high enough to actually work harden a metal reservoir. Work hardening makes the metal more brittle, which eventually will cause leakage when the metal is exposed to continued stress. Wall stresses should also be calculated for vented reservoirs. High stresses develop quickly in large areas of flat plate. And again, weight of the fluid can cause large deflections. Furthermore, mounting peripheral equipment, such as ladders, to a reservoir increases the need to specify stiffening members and thicker plate. Cleaning and maintenance Reservoir servicing must also be taken into account. There must be provisions to drain both return and suction areas of the tank, especially if a dam is installed to separate them. Pipe couplings often are used, but SAE O-ring ports provide better sealing. Valving should also be provided to close off inlet lines when replacing pumps or other components that are mounted below fluid level. This is often wishful thinking, but access should be provided for cleaning and maintaining the interior of the tank. Ideally, hatches should be large enough to provide enough room for service personnel to maneuver cleaning tools. There should also be means for lighting each portion of the tank for inspection.
RESERVOIRS
rigid enough to allow lifting and moving while it is full of fluid. Appropriate lift rings, lugs, or forklift provisions should be included. Accessories Reservoir accessories are used for: ● straining new fluid as it enters a system ● filtering air drawn into the reservoir as hydraulic fluid level rises and falls during system operation ● indicating fluid level in the reservoir ● indicating fluid temperature ● routing return fluid to minimize potential pump cavitation and improve heat transfer ● heating cold or low-viscosity fluids to necessary operating temperature, and ● removing ferrous contaminant particles from the fluid. Fluid must be added to the reservoir at startup, after cleanout, and to make up for losses. Two filler openings should permit reasonably rapid filling (at least 5 gpm each), intercept large contaminant particles from the new fluid, and either seal when closed or filter incoming air if vented as a breather. The openings should be on opposite sides or ends of the reservoir. Metal strainer screens of 30mesh or finer should have internal metal guards and be attached so tools are necessary for removal. The filler cover should be permanently attached, and if it does not include a breather, a separate breather should be specified. In either case, 40-µm air filtration should be provided. In addition to slowing down fluid returning to the reservoir, reducing foaming and pump cavitation from flow dis-
Fig. 3. This industrial hydraulic power unit consists of five pump-motor assemblies supplied by an overhead reservoir. The overhead mounting provides pressurized fluid to each pump’s inlet, and mounting pump-motor assemblies offset from reservoir provides access for lifting pump-motor assemblies from overhead.
turbances at the inlet, and providing fluid mixing without agitation, flow diffusers also reduce noise and the need for baffling. They are especially effective in small reservoirs with relatively high flows and in deep reservoirs with a small floor area. A fluid-level indicator should be located at each filler. Indicators should have high and low levels marked against a contrasting background to help maintain appropriate fluid level. An electronic level indicator can serve as a more sophisticated alternative. Electronic devices use a variety of means to measure liquid level. Transducers produce a continuous output, and switches signal when liquid reaches a predetermined high or low level. Fluid temperature measurement is not required by the NFPA standard, but a selection of thermometers is available, many in the same housing as the fluidlevel indicator. (If high fluid temperature is a continuing problem, the heat source in the circuit should be identified and removed.) As with level indicators, a variety of electronic temperature indicators are available. In either case, signals generated by these devices are routed to a display or control panel to provide operators with an indication of fluid status. Wiring a level or temperature switch into the machine’s control can prevent equipment damage by shutting down the machine if fluid reaches a dangerously low level or high temperature. After shutdown, or when the reservoir is exposed to colder temperatures, the fluid may be too cold for immediate operation. Cold fluid may become viscous or thick enough to prevent it from being drawn into the pump, causing pump cavitation or other problems that can damage components or cause system malfunctions. A thermostatically controlled heater to warm fluid until its viscosity becomes compatible with the system solves this problem. Again, by wiring this thermostat into the system control, machine operation can be prevented until fluid reaches a minimum temperature. Magnets can be placed in the reservoir to capture and remove metallic particles from the fluid stream. Fluid returning to the reservoir should be routed past intank magnets to collect as many ferrous particles as possible. Magnets should be
Sightglass
Filler/breather/strainer
Tank Baffle plate
Drain Baffle hole
Fig. 4. An L-shaped reservoir combines the advantages of base- and top-mounted reservoirs by providing not only easy access to the pump, motor, and other components but a flooded pump inlet as well.
checked periodically and cleaned to ensure continued maximum performance. Although hydraulic filters are usually not considered reservoir accessories, almost all pump inlet strainers are located within the reservoir, and many other filters mount on or through reservoir surfaces. Because the inlet strainer is out of sight, a pressure gage will help indicate when cleaning is necessary. Integral reservoirs In some systems, the hydraulic reservoir is built as an integral part of the equipment it serves. Because of the diversity of designs and special design practices, integral reservoirs are not addressed in the NFPA/ANSI standard. Integral reservoirs are used most often with mobile equipment, and their placement often is an afterthought, which necessitates custom-designed shapes to fit into irregular areas. A number of potential problems exist with integral reservoirs that require special consideration. These include: ● available space may limit size. Because heat transfer capacity is a function of size, external oil coolers or heat exchangers may be needed ● irregular shape may require special baffling to properly route fluid ● surrounding equipment may limit convectional heat transfer ● service accessibility may be poor, and ● special heat shielding may be needed to isolate components or the operator from reservoir heat. For more information on integral reservoirs, refer to the box, “Reservoirs for Mobile Equipment.”
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Rotary actuators R
otary actuators can mount right at the equipment joint without taking up the long stroke lengths required of cylinders. Rotary actuators also are not limited to the 90° pivot arc typical of cylinders; rotary actuators can achieve arc lengths of 180°, 360° or even 720° or more, depending on the configuration. Rotary actuator designs In a helical-spline actuator, Figure 1a, fluid entering the right-hand port causes the piston sleeve to move to the left. Internal helical teeth in a stationary ring gear mate with external teeth on the piston sleeve, causing it to rotate clockwise (as viewed from the output shaft) as it moves to the left. As the piston sleeve rotates and moves to the left, helical teeth in its inner circumference mate with external teeth on the output shaft. Because the output shaft cannot move axially, rotational motion generated by the piston sleeve’s external teeth and internal teeth cause the output shaft to rotate. This rotation is proportional to the axial distance traveled by the piston sleeve and twice the arc distance traveled by the piston sleeve. Because there is no internal leakage past its piston, the helical-spline actuator can maintain selected positions with only about 50% of the normal forward driving pressure. The actuator can be stopped at any point along its stroke, and its arc of rotation can exceed 360°. Torque is directly proportional to pressure, and output torque can reach approximately 700,000 lb-in. with 3000-psi fluid pressure. A variation of the helical-spline actuator is the planetary helical actuator. In this design, the piston carries a set of planetary rollers (rather than gears) that engage helical grooves in the shaft and housing. The rollers enable the use of very high helix angles, which produces a more-compact design. Actuators with rotation exceeding 360° and output torque to 27,000 lb-in. from 3000-psi fluid pressure are readily available. An enclosed piston crank actuator, Figure 1b, has an adjustable arc of up to about 100°. This actuator is compact and
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has few mechanical problems. Built-in bearing support overcomes side thrust forces. Fail-safe versions are equipped with a spring that returns the shaft to a safe position in case of power failure or loss of fluid. A scotch-yoke actuator, Figure 1c, has two pistons connected rigidly by a common rod. The central drive pin on the rod engages a double yoke keyed to the output shaft which turns through arcs to 90° maximum. Torque outputs at the beginning and end of the stroke (breaking torque) is twice that at mid-point (running torque). This characteristic is efficient because many applications require high breaking torques to move and accelerate the load. Fail-safe, single-acting, and double-acting models are available. Efficiencies range from 70 to 95%. In a rack-and-pinion actuator, Figure 1d, a long piston with one side machined into a rack engages a pinion to turn the output shaft. This gearset principle is adaptable for use in fail-safe, single-acting and double-acting models. Where balanced loading on the bearings is required, two bi-directional pistons with parallel racks are used; both racks engage the one centrally-located pinion. Rotation to 1800° and torque to 50 million lb-in. are available. Torque is equal in both directions. Multiple-position rack-and-pinion actuators are available that rotate the output shaft to several positions by varying the pressure porting. Output positions can be in any sequence, allowing the actuator to stop at or pass any intermediate position. Rack-and-pinion actuators are particularly useful for heavy-duty applications because they tolerate heavy side and end loads and can accommodate large bearings. Because of their constant torque output characteristics and resistance to drift, they are often used for precision control. Efficiencies range from 85 to 92% in single rack models and from 92 to 97% for double rack models. In a piston-chain actuator, Figure 1e, a circular drive chain is held taut over two sprockets. One sprocket converts linear motion into torque output; the idler
Rotary actuators generally are used to pivot a joint when a conventional cylinder mounting proves impractical due to space, weight, or motion requirements.
sprocket provides automatic tensioning. Two piston-shaped links are located at equal distances on the chain; one piston is larger than the other. The housing containing the mechanism has two parallel piston chambers and a port on each of the two opposite ends. Pressurized fluid entering a port acts against both pistons; the chain moves in the same direction as the larger piston because of the differential forces being exerted. The smaller piston seals the return side of the chain to prevent fluid leakage. Rotation is reversed by reversing porting. A piston-chain actuator provides rota-
(a) Helical spline
(b) Enclosed piston crank
(c) Scotch yoke
(d) Rack and pinion
(e) Piston chain
Fig. 1. Five designs of piston-type rotary actuators
R O TA RY A C T U AT O R S
tion to five complete turns and torques to 23,500 lb-in. The design is limited by the strength of the chain and sprocket, and by its bulk for applications requiring super high torques. Torque is constant throughout the stroke. Bladder and vane actuators In a bladder actuator, Figure 2a, the output shaft is driven by the expansion and contraction of a pair of bladders. When one bladder is pressurized, it pushes against a cup-shaped lever arm, rotating the shaft through an arc of up to 100°. A 4-way valve controls oil or air flow so that while one bladder is pressurized the other is exhausted, and vice versa. Because this type of actuator has zero internal leakage, it is highly accurate, insensitive to contamination. Almost any
Bladder
Lever arm
(a) Blladder
(B) Single vane
(c) Double vane
Fig. 2. Cross-sections of bladder-, single-, and double-vane-type rotary actuators
fluid medium can be used if the bladder is chemically compatible; special housing materials are not necessary. This actuator can transmit torque to 2750 lb-in. Another bladder design uses a rackand-pinion mechanism driven back and forth by expanding and contracting air bags. The bags hold precharges indefinitely and provide rotation to 90° or 180°. Highest torque is at the beginning of the rotation cycle. Torques to 45,000 lb-in. A single-vane actuator, Fig. 2b, has a cylindrical chamber in which a vane connected to a drive shaft rotates through an arc to 280°. Two ports are separated by a stationary barrier. Differential pressure applied across the vane rotate the drive shaft until the vane meets the barrier. Rotation is reversed by reversing pressure fluid at the inlet and outlet ports. A double-vane actuator, Figure 2c, has two diametrically opposed vanes and barriers. This construction provides twice the torque in the same space as a singlevane actuator, however, rotation is generally limited to 100°. Vane actuators are easy to service because they have fewer parts and lesscritical fits than many other types of rotary actuators. Their mechanical efficiencies range from 80 to 95%, depending on construction and application. The performance of a vane-type actuator depends on the ability of the vane seal to prevent leakage at higher pressures. Vane-type actuators transmit torques to 700,000 lb-in. Mounting To avoid excessive wear and premature failure, it is essential that very little or no thrust or overhung load be imposed on the actuator’s output shaft unless it is equipped with bearings (such as taperedroller bearings) to accommodate these loads. Use a flexible shaft coupling to eliminate side loading from shaft misalignment. When side loading is unavoidable, support the output shaft with auxiliary bearings if the actuator is not equipped with adequate bearings to support such a load. Actuators coupled to gear trains belong in this category. Many actuators are available with an integral bearing package as an option. To bleed trapped air, mount the actuator so the supply ports are on top. Or provide a suitable air bleeding device for the system. Larger rotary actuators often
have built-in bleed. In applications that involve continuous cycling and where hot hydraulic fluid may collect near the actuator, arrange for greater fluid circulation. Heat exchangers may be required. Do not install rotary actuators where contaminants are likely to collect — for example, at the system’s low point. For linear motion? An actuator that rotates at constant speed can move heavy loads very efficiently in a linear direction by using a harmonic motion mechanical linkage. The harmonic motion produced offers a maximum mechanical advantage at the beginning of the stroke to accelerate the load quickly. Halfway through the stroke the load is at maximum velocity. The deceleration half of the stroke is a mirror image of the acceleration half. Heavy loads are slowed automatically and stopped with a force equal to that originally used to accelerate the load. Automatic advance and return of a load at maximum speed can be obtained by using a 360° rotary actuator connected to a linkage rather than a 180° actuator. During deceleration, energy is not transmitted back into the hydraulic system but is used by the actuator to work the linked load. Less hydraulic fluid and horsepower is necessary. To size actuators for such applications, determine friction losses and the force needed to accelerate and decelerate the load. Applications Rotary actuators are used for mixing, dumping, intermittent feeding, screw clamping, continuous rotation, turning over, automated transfer, providing constant tension, and material handling. They are also suitable for turning, toggle clamping, indexing, positioning, oscillating, lifting, opening, closing, pushing, pulling, and lowering. For example, in the steel industry they up-end coils, turnstiles, and rollover devices, and tilt electric furnaces. In material handling they switch conveyors, turn and position container clamps on lift trucks, tension, guide, operate valves, and brake. In marine operations they open and close hatches, swing cargo handling gear, operate booms and all types of large valves, position hydrofoils, and control steering.
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Seals and sealing technology H
igh-pressure sealing generally refers to confining fluids at pressures above 5000 psi. Below these pressures, standard energized urethane lip seals and U-cup seals function satisfactorily without special provisions. Above them, some sort of special sealing devices are necessary. To be effective, seals must perform three basic functions: Seal — sealing elements must conform closely enough to the microscopic irregularities of the mating surfaces (rod to seal groove and/or piston groove to cylinder bore, for example) to prevent pressure fluid penetration or passage, Figure 1 Adjust to clearance-gap changes — the seal must have sufficient resilience to adjust to changes in the distance between mating surfaces during a cylinder stroke. This clearance gap changes size because of variations in the roundness and diameter of the cylinder parts. The clearance gap also may change size in response to side loads. As the size of the gap changes, the seal must match the size change to maintain compressive sealing force against adjacent mating surfaces, and Resist extrusion — the seal must resist shear forces that result from the pressure differential between the pressurized and unpressurized sides of the seal. These shear forces attempt to push the elastomeric seal into the clearance gap between adjacent metal surfaces, Figure 2. The seal must have sufficient strength and stiffness to resist becoming deformed into the gap and damaged or destroyed. Higher pressure improves sealing Elastomeric materials also must seal while accommodating dimensional variations caused by manufacturing tolerances, side loads, and cylinder de-
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Successful sealing involves containment of fluid within hydraulic or pneumatic components while excluding contaminants.
,,,,, , , , , , ,,,,, ,,,,, ,,,,, ,, System fluid
Extrusion gap
(b)
(c)
O-ring
(a)
formations under pressure. Understand that in general, sealing improves as fluid pressures increase. System pressure on the seal surface attempts to compress the seal axially. This compression forces the seal more tightly into the gland and helps improve conformability of the seal with its contacting metal surfaces. If the clearance gap increases during the stroke, resilience of the compressed elastomeric seal causes it to expand radially and maintain sealing force against the metal surfaces. System pressure combines with seal resilience to increase compressive sealing forces when the clearance gap increases. It
Fig. 1 (a) Seal material must conform to microscopic irregularities in metal surfaces to block fluid passage; (b) to adjust to clearance gap size changes, the seal must expand or compress rapidly to follow dimensional variations; (c) to resist being forced into the extrusion gap, the seal must have sufficient modulus and hardness to withstand shear stress produced by system pressure.
generally is true that, as system pressure increases, sealing force and the resulting sealing effectiveness also increase if the seal is correctly designed. The seal’s internal shear stresses increase as system pressure increases. With increasing pressure, the stresses eventually exceed the physical limits of the seal elastomer, and it extrudes into the gap. Difficulties presented by high pressure are not primarily sealing problems but are problems of keeping the seal in its gland while maintaining its structural integrity as increasing system pressures force the seal into the gap. Almost all of the design and in-service technology of high-pressure seal-
,,, ,,, , , ,,, , , ,,, ,,,, ,,,, , ,, , ,,, , SEALING TECHNOLOGY
System fluid
Extrusion gap
(a)
System fluid
Extrusion gap
(b)
System fluid
Extrusion gap
(c)
Fig. 2. As system fluid pressure increases, (a) to (b), an O-ring seal is progressively forced into the extrusion gap. Finally, (c), the physical limits of the seal material have been exceeded.
ing deals with protecting the elastomeric seal from the potentially destructive distortion caused by high system pressures. With proper backup to reduce the size of the gap, relatively fragile elastomers can successfully seal extremely high pressures. When handling a 90-durometer energized urethane lip seal or U-cup at room temperature, the seal seems to be made of an extraordinarily stiff, tenacious material. It requires well-designed experiments and/or sophisticated computer simulations to visualize the state of such a seal inside a hydraulic cylinder at normal operating temperatures and pressures. At pressures as low as 600 psi for 70-durometer nitrile rubber and 1500 psi for 90durometer urethane, the seal cross section is significantly deformed. It changes shape almost instantaneously in response to pressure spikes or changes in the size of the clearance gap. Literally, the seal becomes an annular glob in the seal gland. Seal extrusion The ability of a seal to resist extrusion into the gap depends on the inter-
action of these five factors: ● system operating pressure ● system operating temperature ● size and type of clearance gap ● seal material, and ● seal design. System operating temperature is especially important in high-pressure applications because most elastomers soften and lose their ability to resist extrusion at higher temperatures. Some design methods that help lower high system temperatures include the use of low-friction materials, an increase in fluid volume, and a decrease in the cycle rate of the system. However, when ambient temperature is high, and operating conditions are extreme, it is possible for system temperatures to exceed design parameters. Under such conditions, it often becomes necessary to upgrade seals and for anti-extrusion devices to be more temperature-resistant. The size of the extrusion gap can be controlled throughout the design and manufacture of the cylinder, piston, rod, and end cap. Decreasing manufacturing tolerances increases cylinder cost, however, and also may increase the probability of metal-to-metal interference. Additionally, reducing the extrusion gap size is inherently limited by differential thermal expansion of mating metal components. The actual size of the extrusion gap is a function of:
,,,,, ,,,,, ,,,,, ,,,,,
Fig. 3. Standard PolyPak in modified urethane such as Molythane can be used at pressures to 5000 psi. Higher modulus elastalloys such as PolyMyte, in the basic PolyPak configuration, operate successfully to 7000 psi at moderate temperatures and standard tolerances.
● the nominal gap designed into the cylinder ● manufacturing tolerances, including diametrical variation and ovality ● diametrical expansion of the cylinder caused by system pressure ● side loads, and ● wear on radial load-bearing surfaces. Because all these factors vary, and because the variances can be cumulative, seal design and material must resist extrusion through the largest gap likely to be encountered at design pressure and temperature. Material is the key The key to high-pressure sealing is the use of a material or a combination of materials that has sufficient tear strength, hardness, and modulus to prevent extrusion through the gap. At pressures of 5000 to 7000 psi, the strongest elastomeric materials in standard seal configurations resist the extrusion without reinforcement. At higher pressures, the elastomeric sealing element must be backed by a higher modulus and harder material. Various more-or-less standard backup configurations have demonstrated their effectiveness over many years. At pressures in excess of 20,000 psi, the extrusion gap must be closed and the elastomeric seal must be protected by a sequence of progressively harder, higher-modulus materials. Properly designed, this progression of materials prevents extrusion, tearing, cutting, or other destructive deformation of the elastomeric seal and distributes loads more uniformly to the element that bridges the gap. Designs, materials In high-pressure applications, material characteristics, such as high modulus, tear strength, self-lubrication, and abrasion resistance, become increasingly important. The following seal configurations and materials are specially suited to high-pressure applications. Although these examples cite proprietary compounds as typical, other manufacturers offer their own proprietary compounds, which generally have similar properties. Abrasion-resistant and self-lubricating materials should always be used at these pressures because friction in-
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,,,, , ,, , ,,, , ,,,, , ,, , ,,, ,
SEALING TECHNOLOGY
,,,,, ,,,,, ,,,,, ,,,,,
Fig. 4. A standard PolyPak with a modular backup ring of elastalloy such as PolyMyte will seal successfully to 12,000 psi.
,,,,, ,,,,, ,,,,, ,,,,,
Fig. 5. A positively activated PolyPak of modified urethane such as Molythane with a backup ring of Nylatron can seal to 10,000 psi.
creases at high pressure . Some of these materials are: Enhanced polyurethane — At the lower end (5000 psi) of the high-pressure continuum, a standard PolyPak configuration of modified polyurethane energized by a resilient O-ring elastomer, Figure 3, is sufficient. Polyurethane-based materials — such as Molythane (impregnated with 1
Service recommendations based on test conditions of 100,000 pressure cycles at 60 cycles/min from zero to 5000 psi at 160°F (71°C) with maximum 0.010-in. diametral and 0.005-in. radial clearances.
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molybdenum disulfide to provide dry lubrication plus good compatibility with lubricating properties of working fluids) — are suitable for application pressures up to 5000 psi without backups.1 Molythane comes in a 90-Shore A durometer formulation for PolyPak seals and in a 65-Shore D durometer formulation with a higher modulus for increased extrusion resistance for antiextrusion devices. Ultrathane K-24, a high-tensile, reduced-friction, enhanced-urethane material also is suitable for applications to 5000 psi without reinforcement. Elastalloy co-polymers — Various elastoplastic or elastalloy copolymers, such as PolyMyte — a material with exceptionally high tear strength, abrasion resistance, hardness (Shore D 65), and modulus — offer high pressure performance capabilities. PolyMyte configured as PolyPak and energized by a resilient elastomeric O-ring is suitable for applications up to 7000 psi without backups.2 A high-durometer PolyMyte modular backup, Figure 4, used in conduction with a Molythane PolyPak, can withstand pressures up to 12,000 psi or more.3 Non-elastomeric materials. Nonelastomers include polyamide resins such as nylons and modified nylons and metal backup rings, typically ductile bronze or brass. One non-elastomer is Nylatron, a glass-filled polyamide resin. A Molythane PolyPak with a positively actuated Nylatron backup ring inserted to bridge the extrusion gap, Figure 5, can be used successful at pressures to 10,000 psi. For extreme pressures in one direction, a three-part sealing system, Figure 6, is recommended. The seal is made of a Type B PolyPak, backed by a filledpolyamide modular backup beveled at 308. A wedge-shaped, skive-cut splitring, machined from ductile bronze or brass, is placed behind the beveled modular backup. The metal backup and seal groove are mated at a 458 angle. 2
Service recommendations based on test conditions of 100,000 pressure cycles at 60 cycles/min from zero to 7000 psi at 160°F (71°C) with maximum 0.010-in. diametral and 0.005-in. radial clearances.
,,, , ,,,, , ,, , ,,,, , ,, , ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, Type B (deep) Poly-Pak
,,,, ,,,, ,,,, ,,,, ,,,, High-strength Molythane modular backup beveled at 30°
Ductile bronze or brass anti-extrusion ring with 45° bevel
Fig. 6. At extreme pressures, a metal antiextrusion device of ductile bronze or brass and a high-strength, high-durometer modular backup ring is required.
Under pressure, the wedge-shaped metal ring expands to close the extrusion gap. This design has operated successfully at pressures to 100,000 psi in a specialized application for making synthetic diamonds. Compressed by the elastomeric urethane PolyPak, this elastoplastic modular backup expands radially to fill the groove and prevent sealing-element extrusion. Without an additional anti-extrusion device, the elastoplastic modular backup would experience plastic flow into the gap at 100,000 psi. A softer, lower tear-strength urethane back-up element would be nibbled or cut by the metal backup ring especially where the metal ring is split. 3
The standard PolyPak fits a gland width equal to nominal depth (that is, square). The modular backup also is square; it occupies a space identical in size to the PolyPak it backs up.
SEALING TECHNOLOGY
These proven designs and materials are typical of those available to increase the pressure capabilities of elastomeric seals in dynamic applications. Many other materials suitable for high-pressure applications are available. Often, the choice of seal materials is dictated by the fluid medium, system operating temperatures, cost, or system pressure. The potentially higher efficiency of high-pressure systems comes at a slight cost premium. Sealing materials for high pressures are more expensive, and seal designs often are more complicated. Higher sealing pressures increase sealing force and friction. Increased friction causes higher wear rates and may require more frequent seal replacement, but frictional force and wear rates typically increase more slowly than pressure. Hydraulic system design today often seems to focus on dramatic high-pressure applications. For example, the aerospace industry is presently evaluating 8000-psi systems for future aircraft in special test beds, such as LockheedGeorgia’s HTTB. Many successful high-pressure systems incorporate innovative seal designs in both static and dynamic modes of operation.
Table 1: Average dynamic and static squeeze levels (dimensions in inches) Seal Percent compression Squeeze or compression Gland packing space cross-section
0.070 0.103 0.139 0.210 0.275
Dynamic
Static
Dynamic
Static
Dynamic
Static
0.0565 0.0900 0.1225 0.1870 0.2400
0.0510 0.0820 0.1120 0.1715 0.2275
0.0135 0.0130 0.0165 0.0230 0.0350
0.0190 0.0210 0.0270 0.0385 0.0475
19.3 12.6 11.9 11.0 12.7
27.1 27.1 19.4 18.3 17.3
equipment is not operating, and the pressure falls to zero. Or, in some applications, the system’s pressure may never exceed 100 psi. These are typical of the types of operations defined as low pressure; that is, when the confined fluid media exert little or no pressure force on the sealing element to affect or augment a seal. Within the framework of low-pressure sealing, several primary design considerations affect sealability: ● seal squeeze ● compression set ● sealing force ● gland surface finish conditions, and ● molding flash.
Squeeze A seal component is generally installed in a groove machined into one of the surfaces to be sealed. As the two Low-pressure considerations surfaces are brought together to form a Almost every hydraulic system, gland, they squeeze the diametral cross however, will face occasions when the section of the seal. The mechanical squeezing action deforms the seal cross section; the degree of deformaVoid Crown tion obviously is a function of the squeezing force. In low-pressure applications, the tendency of the squeezed elastomer to maintain its original shape creates a Chemically bonded seal. As the elastomer shape is deformed in its gland, it exerts a counter force against the mating surfaces Flange equal to the force squeezing it, Figure 7, and hence, provides the available sealing force. We can see then that squeeze is a major low-pressure consideration. The recommended squeeze levels are Flange a function of seal cross section, the Squeeze application conditions and whether Metal-to-metal load path the application is dynamic or static. Typical dynamic compression is Fig. 7. Simple representation of how squeeze lower than static compression, due to force compresses a combination gasket/O-ring seal wear and friction considerations. Table 1 summarizes dynamic squeeze seal during gland assembly.
levels as defined by MIL-G-5514F — a document which is a good guide to those parameters. Static data in the table are summarized from common industrial practice. Compression set Compression set reflects the partial loss of memory due to the time effect. In hydraulic systems operating over extreme temperature ranges, it is not uncommon for compression-type seals, such as O-rings, to leak fluid at low pressure because they have deformed permanently or taken a set after used for a period of time. The term compression set refers to the permanent deflection remaining in the seal after complete release of a squeezing load while exposed to a particular temperature level. As related to low-pressure sealing, set—the loss of memory—reduces the compressive sealing force. Compression set is expressed as a fraction of the initial squeeze. Thus, a 0% compression set value indicates complete recovery from a compressive load, producing the maximum possible compressive sealing force. A 100% set value indicates no recovery or rebound at all. A seal in this condition will no longer provide a sealing force and hence, has no ability to act as a lowpressure seal. The bar graph in Figure 8 depicts the range of typical compression set values for various sealing elastomers. Of course, compression set properties are a major but not the only factor affecting elastomer choice for low-pressure sealing. Compatibility with various hydraulic fluids must be considered as well. Sealing force There are several factors affecting the sealing force: ● material hardness ● percentage squeeze, and
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SEALING TECHNOLOGY
Table 2: Surface finishes for special media Ethylene propylene diene monomer ASTM Test Method D395 70 hours at 212°F in air 25% deflection
Ethylene propylene Nitrile, Buna-N Fluoroelastomer Polyurethane
Tetrafluoroethylene/propylene Perfluoroelastomer Epichlorohydrin Neoprene 20
30
40
50
60
70
Percent compression set
Fig. 8. Percentage compression set exhibited by typical families of sealing elastomers used in fluid power systems.
● seal cross-section size For a certain amount of squeeze, the sealing force is directly related to the hardness or elastic moduli of seal materials for low-pressure applications. The harder the material, the larger the initial sealing force. A seal material has a nonlinear stress-strain curve and needs to be described by special material models. For simplicity, linear moduli, such as Young’s modulus and shear modulus, are usually used due to their direct relations to the material hardness. The modulus commonly used for specifica-
Compressive load per linear inch — pounds
tion purpose is tensile stress at a specified elongation. For example, modulus at 100% elongation is the tensile stress corresponding to that elongation. Hardness generally is measured with a durometer gage — typically using the Shore A scale. The gage measures the force required to deflect the flat surface of a rubber specimen with a pointed indicator. The A scale ranges from 0 to 100; a 90 Shore A compound would be designated as a hard (or high-viscosity) material, and would exhibit much higher compressive force than a 60 Shore A compound, which would be classified as soft. For a specific material, seal comRoom temperature, NBR material pression force of the elastomeric 70 durometer, Shore A 80 durometer, Shore A material increases as the percentage 90 durometer, Shore A deflection of the seal’s diametral cross-section increases. Dynamic 54.3 50 squeeze levels typically should be limited to around 12% due to fric44.3 40.4 40 tion and related-wear considera35.5 30.3 tions. Static squeeze levels can be as 30 25.0 high as 30%. 19.8 19.4 It generally is recommended that 20 19.9 13.3 12.6 a minimum of 0.009-in. squeeze be induced on radial seal cross sections 10 12.6 7.2 9.1 due to compression set considera5.0 tions. Maximum radial squeeze 0.070 0.103 0.139 0.210 0.275 2 should be held to 30% because Seal cross-section — in greater squeeze causes assembly Fig. 9. Plots of compressive load vs. O-ring difficulties and elastomer deterioraseal cross section for three different seal ma- tion. Compressive sealing load is also directly related to the size of terial hardnesses.
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Static (RMS)
Cryogenic/low molecular gas
4-8 in.
6-12 in.
Low viscosity fluid and gases
6-12 in.
6-16 in.
seal’s cross-section, Figure 9.
Phosphonitrilic fluoroelastomer
10
Dynamic (RMS)
Fluid media
Gland surface finishes Two physical characteristics of the seal contact-band areas can affect how well the available sealing force is transmitted. These are: ● sealing surface finishes in the gland, and ● parting line projection and flash on the seal. The finish on machined surfaces that come into contact with the seal is a significant factor in achieving optimum seal performance. Finishes can be defined by different systems, which are often misunderstood and sometimes incorrectly specified in hydraulic design. The American Standard Association provides a set of terms and symbols to define basic surface characteristics, such as profile, roughness, waviness, flaws, and lays. Roughness is the most commonly specified characteristic and is usually expressed in units of min. Roughness provides a measure of the deviation of the surface irregularities from an average plane through the surface. In most cases, geometric average roughness or root mean square (RMS) is the preferred method. RMS measurement is sensitive to occasional peaks and valleys over a given sample length. As related to low-pressure sealing, the sealing element must penetrate these micro imperfections and irregularities in order to block the passage of the fluid media across the contact band area. It is generally accepted and recommended that dynamic interfaces should not exceed RMS values of 16 min. or 0.4 mm. Static interfaces should not exceed RMS values of 32 min. or 0.8 mm. Special fluid media would benefit from smoother finishes as listed in Table 2.
SEALING TECHNOLOGY
controlled. Control is especially critical in low-pressure applications and applications sealing gas-oil interfaces. Standards such as MIL-STD-413E and those in the Rubber Manufacturers Association (RMA) Handbooks provide guidelines on allowFree state Installed able flash criteria for manufacturers and users. Sealing perforFig. 10. Cross-sectional sketch of a combination gasket/Omance characteristics ring seal before and after installation. can be enhanced by eliminating the flash line completely from dynamic and static sealing interfaces. This practice is especially desirable in accumulator applications and those requiring low-viscosity fluid media, such as silicone oils. Manufacturers may offer an optional flash-free seal design for these stringent applications.
Mechanically bonded
Fig. 11. The basic element of a combination gasket/O-ring seal is a retainer with grooves in one or both surfaces into which an elastomer is molded.
Parting line projections and flash Just as there are irregularities in the form of roughness on the gland surface, there are irregularities or imperfections on the sealing element known as parting line projections and flash. A parting line projection is that continuous ridge of material along the line where the mold halves come together at the ID and/or OD of molded rubber seals, such as O- and T-rings. It results from worn or otherwise enlarged corner radii on the mold edges. Flash is a thinner, film-like material that extrudes from the parting line projection. It is caused by mold separation when material is introduced or inadequate trimming or buffing after molding. Because flash lines are inevitable in clam-shelltype, compression molding processes, the degree of flash must be
Gasket/O-ring-seal for static applications There are three primary static sealing methods in use today. The flat gasket is the oldest of the three. Where reusability is not required and where the possibility of some leakage can be tolerated, the flat gasket may be the best choice. The O-ring represents a marked improvement over the flat gasket for installations where lit-
tle or no leakage can be tolerated. The combination gasket/O-ring seal, Figure 10, represents a significant improvement over both the flat gasket and the O-Ring in a groove for near zero-leakage sealing in static applications. Advantages of the combination gasket/O-ring seal are: ● ease of installation ● sealing element(s) molded precisely in place ● limited area of seal exposed to fluid attack ● visibly inspectable after assembly ● no retorquing required ● high reliability, and ● no special machining of mating flange surface required (no grooves). The combination gasket/O-ring seal consists of a retainer plate with a groove in one or both element(s). This seal may be either chemically bonded to the groove, Figure 7, and/or mechanically locked in place by cross-holes in the groove, Figure 11. The combination gasket/O-ring seals are relatively more expensive than O-rings. FEA-assisted seal design Vitally important to any method of sealing is the ability of the seal to achieve the proper balance between developing enough elastomer stress to provide an adequate seal and not developing too much stress, which would prematurely degrade the seal. Depending on the type and requirements of the seal, this seal/stress rela-
Fluid pressure
Squeeze
Squeeze
Fig. 12. The illustration to the left is an FEA mesh model of a U-cup cross section, while at right is the deformed shape after installation.
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SEALING TECHNOLOGY
tionship will be different. The study of elastomer stress and its relationship to seal effectiveness has been dramatically enhanced with the advent of Finite Element Analysis (FEA). FEA is a numerical modeling technique that has been used quite successfully for seal applications. FEA can predict seal deformed shapes and stress distributions after installation, in operation and under various conditions. This information is very important in evaluating the following: stability, sealability, thermal deformation, swelling, and seal life. FEA is becoming a very powerful tool for seal design optimization.
The procedure for FEA-assisted seal design can be summarized as follows: ● seal shape sketch ● material selection ● material characterization testing (such as tensile stress strain curve, bulk modulus, thermal constants, friction constants, etc.) ● material model selection (MooneyRivlin, Ogden, etc.) ● mesh modeling, boundary condition definition ● numerical analysis ● post-processing (output), and ● to see if the seal shape needs to be modified.
Figure 12 shows an example of an FEA plot. FEA is also used for flow and mold analyses, which are desired for elastomer processing control. Seal materials The worldwide industries that design equipment incorporating hydraulic and pneumatic technology have changed considerably over the last 20 years — largely in response to the increased expectations of the end user. From the standpoint of sealing, these expectations now call for effectively leak-free systems, regardless of the application. Whereas two decades ago almost
Basic properties of elastomeric seal compounds Although elastomeric compounds used in aerospace seals are derived from relatively few base polymers (such as nitrile, fluoroelastomer, and ethylene propylene), each seal manufacturer usually develops special compounds of these base polymers to enhance or suppress different chemical or physical properties to fit specific requirements of an application. Proprietary formulations of these compounds are kept secret. Even the analysis of a finished elastomer seal presents an incomplete picture of the original elastomer compound because some ingredients are consumed in processing. Of all compound properties, the most critical are the changes that occur. Every property of every compound changes with age, temperature, fluid, pressure, squeeze, and other factors. Standardized tests have been developed to provide comparability in changes among compounds. Compounds with the least tendency to change properties are the easiest to work with; they produce a seal that is adaptable to more applications. The number of properties evaluated for an application depends on the severity of conditions. Factors are highly interdependent, but typi-
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cally include resilience and memory, abrasion resistance, coefficient of friction, and fluid compatibility. Let’s take a closer look at each of these. Resilience and memory are defined as a compound’s ability to return to original shape and dimensions after a deforming force is removed. Resilience implies a rapid return, while memory implies a slow return. In seals, resilience is important because it permits a dynamic seal to follow variations in the sealing surface. Although elastomer resilience is frequently measured on a Bashore resiliometer, field experience is required to relate ratings to seal performance. Additional attention is required for low-temperature applications. When temperature is too low, a compound loses its memory. Abrasion resistance — resistance to wear when in contact with a moving surface — is the product of other properties, including resilience, hardness, thermal stability, fluid compatibility, and tear/cut resistance. It also is influenced by the compound’s ability to hold a film of protective lubricant on its surface. Harder compounds are usually more resistant to wear, so dynamic seals of 85-durometer compounds are common. However, if the seals encounter high temperatures, it may be good practice to specify an even harder material to
compensate for the softening effect of heat. In low-temperature applications, a softer material might be preferred because elastomers tend to harden a temperatures drop. Coefficient of friction (usually only important in dynamic seals) is compound-specific and different for running and break-out. Usually break-out friction is higher. Breakout friction increases with time between cycles. Coefficient of friction is affected by temperature, lubrication, and surface finish. Aging and the influence of service fluids on the compounds may also affect hardness and, in return, both breakout and running friction. As far as fluid compatibility is concerned, a fluid is considered incompatible with a compound if the fluid causes enough property changes to reduce sealing function and/or shorten the working life of the compound. Dissimilar chemical structure is the key to fluid compatibility. For non-polar liquids — such as hydrocarbon fuels and oils — nitriles, fluorocarbons, or fluorosilicone polymers are normally used. For polar liquids, such as phosphate ester hydraulic fluids, ethylene propylene compounds are most satisfactory.
SEALING TECHNOLOGY
all leading OEM’s around the Material Applications Positive factors Precautions world had their own accept- Nitrile Fluid power cylinders Inexpensive; good Not tough enough to resistance to set withstand very smooth ability curves which aspiring surface finishes (,0.4 min. CLA) suppliers had either to meet or Better wear resistance Limited low-temperature beat, today their approval pro- Carboxylated nitrile than nitrile flexibility, compared with cedures simply state that zero standard nitrile leakage is the standard. Much EPDM Exposure to fireResistant to HFD fluids Not resistant to mineral of the credit for this situation resistant fluids and Skydrol oils, greases, other lies at the door of the Japahydrocarbons nese; not so much for any in- Fluoroelastomer High temperatures Resistant to most Relatively expensive and novative design but for their (to 4008F) hydraulic fluids difficult to process attention to detail, and their PTFE General sealing Low friction Not elastomeric, requires elevation of the market perenergization ception of quality. Part of Polyurethane General sealing elements Good wear resistance and First generation subject resistance to set — to hydrolysis effects of this, of course, demanded energization not required water above 1208F leak-free systems. Rubbing faces of seals. Elastomeric; good Poor resistance to set; Europe in the ‘70s responded Polyester Anti-extrusion elements resistance to wear and fluids requires energization to the export drives of the large Japanese off-highway equipment manufacturers with tough new is now so good that TPU seals are for instance. For port-passing appliquality standards, plus manufacturing, used in underground-mining cylin- cations, such as phasing cylinders, by design, and sourcing reviews. One re- ders that operate on high-water- exploiting the wear resistance and sult of these reviews was a move to- based, fire-resistant fluids. hardness of TPE, seals can be deward higher system pressures to inPneumatic cylinder designers also signed specifically to overcome probcrease machine output. Typical have benefitted from the advances in lems often associated with this type European off-highway equipment now TPU sealing. Calls for very low fric- of cylinder design. operates between 5000 and 8000 psi. tion and ultra-long service life have The key to success in today's inOther sectors followed this trend, and been accommodated by TPU seals dustry for the seal maker lies in comtoday we see 5000-psi and higher-pres- which offer 50% of nitrile's breakout bining the latest material technology sure hydraulic systems in many differ- friction and have lasted for 12 2 106 with innovative profiles to provide cycles in 2-in. bore, 10-in. stroke the customer with solutions which ent industries. To meet these challenges, leading cylinders with non-lubricated air. work. Modern thermoplastic poly-ester international seal manufacturers have modified existing materials and de- elastomers (TPE) have also im- Future trends veloped new ones. These materials proved. It is possible to chemically As environmental issues continue enable seals to be made today in pro- engineer TPEs to produce such desir- to influence almost all industries, the files and configurations unheard of able properties as outstanding wear hydraulics sector will be no excep20 years ago. Modern hydraulic and and fluid resistance. These character- tion. In Europe and the U.S., sopneumatic systems commonly use the istics have made them a first choice called environmentally friendly fluseal materials listed in the table at in many sealing applications — par- ids are being developed. Vegetable ticularly as piston seals where, with oils, such as rape and sunflower seed, right below. suitable energization, extremely effi- have been tried, but they can cause cient performance can be produced. problems for the system (forming TPU and TPE The greatest strides have been Many of these TPE seals compete resin above 1808F) and for the seals made in the thermoplastic with PTFE elements where the elas- and other components (forming acid polyurethanes (TPUs). The major tomeric nature of TPE makes them in any water present that can attack limitations of the first-generation more easy to install and also prevents elastomers). Other fluid contenders TPUs — high lip preload loss (partic- piston drift. An example is in truck- include polyglycols and synthetic esularly at elevated temperatures, say mounted crane outriggers, where the ters, but these also present problems above 1608F) and poor resistance to elastomer can bond into the adjacent — not the least of which is a cost up water and high humidity — have surface finish. TPEs with their supe- to ten times that of mineral oil. New materials and blends will be been overcome. Second-generation rior wear resistance and tensile required to combat the effects of TPUs are now available which take strength are ideal for this use. In Europe, TPEs have a growing these fluids while still providing the the system operating range up to 2508F without suffering serious loss importance in specialty sealing appli- sealing integrity users expect. Preof lip preload, and generally do not cations such as the mining and steel liminary work indicates that there is a require O-ring energization. Hydrol- industries. TPE's heat and fluid resis- long road ahead if this issue becomes ysis resistance in some formulations tance perform well in rolling mills, a reality. 1998/1999 Fluid Power Handbook & Directory
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SEALING TECHNOLOGY
Sealing pipe-thread fittings A review of the important performance properties of compounds of tetrafluoroethylene (TFE) resin and filler materials shows that the resin performs well in many applications without fillers. In fact, fillers can lessen TFE’s outstanding electrical and chemical properties. In mechanical applications, however, compounds of TFE and inorganic fillers offer improved wear-resistance, reduced initial deformation and creep, and increased stiffness and thermal conductivity. Hardness is increased, and the coefficient of thermal expansion is decreased. These compounds can therefore make it possible to gain the advantages of TFE in applications where the unfilled resin cannot be used. Many different fillers can be blended with TFE, but most application requirements have been met with five filler materials: glass fiber, carbon, graphite, bronze, and mo-lybdenum disulfide. The properties of any compound depend on filler type and concentration, and processing conditions. Compounds — such as plain TFE — are made into finished parts by molding, extrusion, or machining. One example of the application of TFE resin and fillers is O-rings made of TFE. They are used where resistance to solvents and other chemicals, or extremely high- or low-temperature resistance is required. These are applications where elastomeric materials are not suitable. An additional benefit of TFE O-rings, in certain applications, is the material’s low coefficient of friction and anti-stick properties. Typical applications are gaskets, rotary seals, piston seals, and valve seals. There are four methods used to seal pipe threads: Trapped dope. The use of drying or non-drying dopes is the old-
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est and least costly thread-sealing method. Made from ingredients ranging from crushed walnut shells in shellac to other fillers and oils, usually with some thinning volatiles, they are inherently weak, and will shrink when the volatiles evaporate. Yielding metal. The sealing interface is limited in area and unlimited in force so that yielding takes place. Metal flow fills misalignment and leak paths. These dryseal joints can be effective up to 98% of the time, but usually cannot be disassembled and reused without leaking. Trapped elastomer. Confined Orings can seal effectively, but also can suffer from sloppy assembly. Damaged threads or pinched rings also can contribute to leakage. Orings typically are used in high-pressure fluid power systems where the extra cost is more easily justified and freedom from contamination is especially desirable. Curing resins. Sometimes called machinery adhesives, these anaerobic acrylic materials develop strength by curing. They are very forgiving of tolerances, tool marks, and slight misalignment. They make tapered fittings as effective as O-rings at a fraction of the cost. They lock free-standing fittings — such as gages. They can also improve the 98% effectiveness of yielding-metal joints to 100%. The correct grade must be selected because of their wide range of strengths so that disassembly will not be hampered. Curing materials are so effective in sealing threads that they are often used on straight threads which enter or plug pressure vessels. In addition, the curing materials are effective even when tapered threads are lightly torqued. Lightly torqued threads (straight or tapered) do not leave high residual stresses in housings or valves that can distort valve bodies to the point of inoperation or long-term fatigue failure.
Design engineers must choose between these options to assure that equipment will function as planned. Specifications should not be left to assembly workers — as is often done. There probably is little argument that the most significant event in sealing fittings during the past 25 years was the appearance of anaerobic pipe sealant with TFE materials. Since the first appearance of these materials, many other companies have added anaerobic thread sealants to their lines. The new sealant technology offers a variety of benefits: Convenient curing. Being anaerobic, it cures in the absence of the air, remaining uncured until the parts are assembled. There is no evaporation, hardening beforehand, or other work-life problems. Lubricity. Containing TFE filler, the material eliminates galling or other component-assembly problems. These products prevent over-torquing to affect a seal. Fills threads. Due to high wetting ability, the material fills threads so well that leakage past nicks, scratches, and dents does not occur. Fitting movement. Systems being assembled with anaerobic sealant can be initially readjusted without breaking the seal in the threads. Vibration resistance. Anaerobic sealant does not permit a fitting to be loosened by vibration. Reusability. Fittings sealed with acrylic and latex-based materials can be disassembled and reused with sealant in the field without danger of leakage. Freedom from contamination. Unlike the tape most often replaced by the anaerobic material, sealant does not break up to contaminate lines and valves.
Shock absorbers
The purpose of energy-absorbing devices is to stop moving loads with minimum load rebound, with minimum shock to the loads, and minimum shock to surrounding structures.
A
without transmitting potentially damaging shocks to equipment. In its simplest form, a shock absorber consists of a double-walled cylinder (with space between the concentric inner and outer walls), a piston, some mechanical means to return the piston, and a mounting, Figure 2. The piston return is usually a spring, which can be mounted externally around the piston rod or internally on the inside of the cylinder body. A series of orifices are drilled in the inner cylinder wall, spaced at exponential intervals. The reason for this exponential spacing is derived from the basic equation for kinetic energy:
rubber snubber, a compression spring, and a dashpot all can absorb energy. The snubber and spring store energy and then release it after the load is removed. This often results in a rebound. The dashpot is a fluid-filled cylinder with an opening through which fluid may escape in a controlled manner. Any force acting against the piston in the cylinder encounters high resistance from the fluid at the beginning of the stroke, then much less as the piston retracts. These three methods of energy absorption accomplish the same objective: they smoothly stop a moving load. But they do not dissipate the load’s kinetic energy uniformly. The impact of a moving load against a resisting force produces peak forces that are transmitted to adjacent equipment and structures, or to the load itself. Figure 1 shows plots of force versus stroke for the same load moving at the same velocity striking a rubber snubber, a spring, a dashpot and a shock absorber. The kinetic energy to be absorbed is the same in each case, but it is dissipated at differing rates. A
Rubber cushion Force Dashpot Spring
Shock absorber Stroke
Fig. 1. Areas under each curve are equal, so energy being dissipated is the same for all four method of energy absorption.
linear rate of deceleration is the most efficient combination of force, space, and time that can be used to stop a moving object. The ideal rate is the almost square plot, where a constant force resists the load, until it is slowed to a stop. Linear deceleration of loads Shock absorbers convert the kinetic energy of a load into heat which is dissipated into the atmosphere. They stop moving loads with no rebound and
KE=1/2mv2. The cylinder is filled with fluid from which all air must be bled because air bubbles will cut the efficiency of the shock absorbers by causing spongy or erratic action. When a moving load contacts the piston rod, it moves the piston inward, forcing fluid through the orifices in the inner cylinder wall. The fluid is forced
Oil return passage
Return spring Knife-edge orifices
End cap
Hollow cylinder wall Piston-ring check valve
Fig. 2. Standard fixed-orifice shock absorber. Note the piston-ring check valve to permit fluid to flow around the piston during repositioning.
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SHOCK ABSORBERS
through the oil return passages, into and sound very smooth. Metering tube the space behind the piston head, FigObviously, hybrid/custom profiles Lock Inner tube Check valve ure 2. As the piston retracts, it blocks are designed to meet special applicathe orifices behind it, reducing the eftion needs. They may include various fective metering area, and maintaining feed controls or have a half-shock/halfa uniform deceleration force as the feed -control characteristic. load loses its energy. Fluid pressure is Sharp- or knife-edge orifices, Figconstant in such a shock absorber, proure 4, with short lengths will create a viding constant resistance to the load. metering path that is essentially unafPiston/rod Spring for return Knurled assembly The load slows to a stop as its ki- adjusting fected by changes in fluid viscosity Rod seal and wiper knob Orifices, adjustable netic energy approaches zero. There is and therefore more temperature-stano rebound because the shock abble. sorber has stored no energy. Instead, it Fig. 3. Adjustable shock absorber shows meterabsorbs it and converts it to heat. ing tube around the orifice area. This alters area Selecting a shock absorber To return the piston to its extended openings to control rate of deceleration. When choosing a shock absorber, the position, several events must happen. most important factor to consider is the First, the load must be removed. The ferent settings. Constant resistance to type of load to be stopped. Basic types spring then pushes the piston outward, load should be evident throughout the of loads encountered in shock absorber opening a check valve which permits stroke of the shock absorber — similar applications include pure inertial, freefluid to flow from behind the piston to to applying a constant stopping force falling, and rotating, as well as loads the space where the piston was prior to with car brakes. subject to additional propelling forces. extension. Smaller shock absorbers — Load weight and velocity are the with bore diameters under 0.75 in. — Orifice design next two most important factors in sizhave ball check valves to control this Orifices can be designed to create a ing a shock absorber. Additionally, pofluid flow. Larger models use piston- certain metering/deceleration profile. tential shock to equipment, number of ring check valves. These profiles are seen by plotting impacts per unit of time, and ambient stopping-force vs. shock-stroke. The conditions must be considered to propFixed vs. adjustable shocks three basic categories of deceleration erly select a shock absorber. ApplicaThere are two basic types of shock profiles are square, rectangular, and hy- tion conditions include extreme temabsorbers: fixed-orifice and adjustable. brid/custom. peratures, load acceleration, maximum The fixed-orifice type stops only one Square profiles minimize stopping propelling force applied to the load, weight, or in the self-compensating force and therefore provide the longest and time limitations imposed on the version, can stop a fairly narrow range equipment life. Fixed-orifice and prop- equipment. Time limitations would inof weights — over a 1:5 ratio, e.g. This erly tuned adjustable shock absorbers clude minimum and maximum cycle type usually is designed for a specific produce square profiles. times and the time required for the weight and cannot be changed to meet Triangular profiles are used when shock absorber to return to the exthe needs of other applications with dif- time is critical, such as when rotary tended position between strokes. ferent weights. turntables or linear slides must move Cycle rate is another important conAdjustable shock absorbers can ac- through their strokes as quickly as pos- sideration. If the shock absorber must commodate a wide range of weights — sible. This profile can be made to look handle many impacts within a given a ratio of 1:50 is possible. They are adtime, it may overheat, resulting in poor justed by moving a graduated dial on performance and premature failure. the outside of the shock absorber. This Rapid cycling can heat the fluid enough moves a ring across the orifices to conto reduce its ability to dissipate energy. trol the size of their exposed openings, As a safety feature, most manufacFigure 3. Controlling the amount of turers recommend that shock absorbers fluid forced through the orifices conbe sized to 50% to 80% of capacity. Betrols the deceleration rate. The dial ro- Fig. 4. Knife-edge orifice showing rela- cause the amount of impact the shock tates through 90º or 180º, and has a cal- tive size of opening compared to thick- absorber can accommodate is inversely ibration scale from 1 to 10. Adjustment ness of cylinder wall. Fluid flow is not af- proportional to the length of its stroke; generally is made on a cut-and-try basis fected by changes in viscosity with this doubling stroke length will cut the imby observing energy absorption at dif- type of orifice. pact of the load in half.
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T R A N S D U C E R T E C H N O L O G Y PA R T 1
Basic principles of transduction E
tioners. There are only three electrical properties that must be dealt with while considering transducers: resistance, inductance, and capacitance.
Changes in electrical properties Thousands of commercial transducers take advantage of changes in the electrical properties of an element to produce a usable output. The advantage of using electrical properties for sensing purposes is that the support equipment must have an electrical component, and the output is an electrical signal that can be fed to one of several electronic signal condi-
Resistance changes The resistance of a given conductor depends on its dimensions and its resistivity, a physical property of the conducting material: the resistance in Ohms equals the resistivity times the length of the conductor divided by the conductor’s uniform cross- sectional area. All conducting and semiconducting materials have a thermal-sensitivity property. That is, when the temperature of the material changes, its resistivity does too. The resistivity of metals increases with increased temperature, but the temperature coefficient of resistance in the resistive films used in many commercialgrade fixed resistors is negative. This, then, results in a decrease in resistance with a temperature increase for these films. Semiconductors also exhibit a negative temperature coefficient of resistance but the rate of resistance change with temperature is much greater than with film. When commercialized, these semiconductor devices are called thermistors, Figure 1(a). They undergo large increases in resistance for small temperature changes and thus can be used as temperature transducers with proper calibration. Semiconductors also are sensitive
very transducer uses a particular physical principle for its operation. A change of some variable causes a reaction in the transducer’s sensing element that can be detected. For example, internal pressure applied to a tube causes an observable, resultant deformation of the tube. When the tube is semicircular, a small deformation of the tube produces a large movement of the tube tip. This transduction principle drives the popular Bourdon tube pressure gage. Another transduction example is the turbine flow meter; fluid moving past a propeller turns the propeller. Following an explanation of the various operating principles of transducers, the discussion will cover the nature of output signals for the measurement and control network chain. The output signal dictates the nature of the necessary signal conditioning and thereby determines, or at least limits, the way interfacing must be done.
The following discussion explains some of the various physical phenomena used in commercial transducers. The insights gained will help the practitioner judge the suitability of a given transducer
to the amount of light energy directed at them. That is, simply exposing the semiconductor to photons causes its resistivity to change. As light intensity increases, the semiconductor conducts more readily because its resistance decreases. The change in resistance can be detected as a change in current or a change in voltage; the commercial version is a photo resistor, Figure 1(b). Coupled with a voltage source and a voltmeter or ammeter, the device becomes a usable photo detector. Thermistors are completely encased in opaque envelopes that elimi-
R RT
E
ºt
Vo
(a)
R Rp
E
ºt
Vo
(b)
Fig. 1. Thermistor Rt and photo resistor Rp can be used in voltage-divider circuits as a temperature transducer, (a), or light detector, (b).
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T R A N S D U C E R S PA R T 1
L A ∆L
∆L
Fig. 2. When a conductor is stretched, its length increases while the area of its cross-section decreases. The smaller area increases the resistance of the conductor.
(b) (a)
(e)
R1
Vo
E R2
(c)
(f)
R3
Rx (a)
(g)
RT1 E RC2
Vo
RC1
RT2
(b)
Fig. 3. Geometrically induced resistance changes are detected with a single activearm bridge circuit (a), or a four active-arm bridge circuit, (b).
nate possible photonic effects. On the other hand, The photo resistor cannot be insulated from temperature changes, so it becomes necessary to ensure that no significant temperature change takes place in the environment of the resistor. Lacking that, the temperature change can be measured and a mathematical correction applied to detector output. Geometric factors such as conductor length and cross-sectional area are used to make resistance changes in strain gages, Figure 2. Stretching a conductor of length L and area A makes the resistor longer while reducing its area and results in an increased resistance. Conversely,
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(d)
squeezing the conductor from the ends makes it shorter and bigger around, decreasing its resistance. These resistance changes are typically less than 0.2% of the total resistance of the element. To detect such small changes, a bridge circuit, Figure 3, is necessary. Strained resistor Rx, Figure 3(a), is in a circuit called a single active-arm bridge circuit; resistors R 1 , R 2 , and R 3 complete the bridge. When this strain gage is used, the three resistors that complete the bridge usually are unstressed strain gages called dummy gages for obvious reasons. When R x is unstressed, all four resistors are nominally the same and output voltage V o is nominally zero. But when R x is strained or stretched, the bridge becomes unbal-
Fig. 4. SR-4 strain gages are available in single-element, (a) and (c), dual-element (b), (e), and (f), or triple-element configurations, (d) and (g). Multipleelement gages measure strain along different axes of the strained member.
anced and produces a measurable output voltage. This voltage can be calibrated against the strain-inducing phenomenon and the result is a usable transducer. This is the principle that pressure transducers, force transducers (load cells), torque transducers, and drag-body flow meters use. Commercial transducers do not use the single active-gage bridge because the circuit lacks sensitivity. The alternative is the four active-arm bridge circuit, Figure 3(b). In this configuration with no strain in any gage, the four resistors nominally are identical in value and the bridge is in balance; there is no output voltage. The four gages must be arranged so they all experience the same temperature to automatically compensate for temperature sensitivity.
T R A N S D U C E R S PA R T 1
Furthermore, the gages must be arranged and mounted on the strained member so they experience different levels of strain. Ideally, resistors RT1 and RT 2 experienced tension while RC 1 and RC 2 experience compression. Now, when the measurand is applied, two gages undergo an in-
T
R1
RTC
Vo
R2
E R3
R4
Fig. 5. Temperature compensation resistor RTC has a negative temperature coefficient so that its resistance rises as gage temperature falls and vice versa. Thus, transducer sensitivity does not change with temperature. The bridge circuit inherently compensates for thermal zero shift when all gages are at the same temperature.
(a) (b)
Fig. 6. An on-off switch changes resistance from zero to infinity when going from closed to open.
Conductor
Induced current, l
i
Alternating flux
lp
Flux AC source
Fig. 7. Alternating flux creates an induced current in a nearby conductor. This results in a change in primary current that can be calibrated in terms of distance between the coil and the conductor.
crease in resistance while the others undergo a decrease in resistance. When arranged as shown in Figure 3(b), the bridge produces the maximum possible output voltage for a given strain and the circuit sensitivity is four times greater than in the single active-arm bridge circuit. The sensitivity of commercial transducers usually is given in millivolts of bridge output V o per volt of supply voltage E when the transducer is subjected to rated full-scale value of the measurand. Note the construction and configurations of several versions of the SR-4 strain gage, Figure 4. This gage is available in single-, double-, and triple-element configurations, usually with resistance values of 120 Ω or 300 Ω. Bridge supply voltage typically is limited to a maximum of about 12 V. Temperature variations affect a bridge circuit in two ways. First, there can be a zero shift because one gage is at a different temperature than the others. That is why it is desirable for all gages to be at the same temperature. If all the gages are of the same material and at the same temperature, the bridge circuit inherently rejects temperature variation as it affects zero or null. On the other hand, current through the resistors affects the sensitivity of the bridge: its gain or volts of output per unit of measurand input. Secondly, temperature changes affect total resistance, current, and sensitivity. To compensate for sensitivity changes, a small resistance with a negative temperature coefficient-of-resistance is connected in series with the bridge, Figure 5, so total resistance and sensitivity remain constant over the usable temperature range. The process of matching the R T C to the bridge is called temperature compensation. In spite of these efforts, the result is always less than perfect, so a statement about the sensitivity to temperature variations of gain (sometimes called span) and zero comes with commercial transducers. Lastly, a common device that makes use of resistance change to create a usable output is the simple on-off switch, Figure 6. Clearly, there are only two possible switch
states, because it can only be fully open (infinite resistance) or fully close (zero resistance). Inductance and inductive coupling changes Variations in magnetic properties also are used to make transducers. A conceptually simple device is the eddy-current detector that uses Lenz’s Law. This law says that a voltage will be induced whenever a conductor is placed in a time-varying or AC magnetic field. The AC magnetic field, Figure 7, is created by a coil excited by an AC source. In practical applications using this principle, the AC source is a local electronic oscillator running at several kHz. The source powers a coil which produces a time- varying magnetic field around it. If a conductor approaches that magnetic field, an eddy current is induced in that conductor and circulates as indicated, Figure 7. Because the conductor is imperfect, there is a small power loss and the conductor’s temperature will rise. The energy causing the power loss must come from the AC source so the current must increase slightly. Any increase in primary current I p, Figure 7, indicates the conductor’s presence. The eddy-current principle is used in metal detectors. Note that the conducting material need not be ferromagnetic; it only need be a conductor. Because of the distance between Alternating flux field Movable slug
lp
AC source
X
Fig. 8. As ferromagnetic member moves closer to coil, inductance increases so primary current goes down. Inductance can be calibrated to measure the distance between the moveable member and the coil.
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T R A N S D U C E R S PA R T 1
the coil and the conductor affects the amount of current in the primary winding, it is possible to calibrate the primary current-change to distance and the device then constitutes a usable position transducer. Commercial versions have been developed that can measure a distance of less than
Ferrous material
LA
LB
X
V0
RA
RB
Fig. 9. Variable-reluctance transducer measures amount of AC bridge unbalanced voltage Vo as function of the position of a ferromagnetic material moving between two coils.
Vo
Movable ferromagnetic core
Fig. 10. Changing slug position changes the amount of magnetic coupling between primary and secondary windings; output voltage depends on slug position. Slug A VoA Vo x VoB B
Fig. 11. LVDT uses dual secondary windings in series subtractively. Thus, output voltage is zero with the slug magnetically centered.
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0.100 in. They have been used to measure paint thickness as well as main-spool position of a servovalve during frequency-response testing of the valve. Such a transducer is called an eddy- current detector. A similar application of magnetism for sensing, Figure 8, uses the fact that when a ferromagnetic material is placed in an alternating magnetic field, net inductance of the primary winding increases. As this inductance increases, primary current I p decreases. This current can be calibrated against the distance between the coil and the ferromagnetic material and the device becomes a position transducer. These are called variable reluctance transducers, because the reluctance or resistance to magnetic flux is the physical quantity that varies. A variation of the variable-reluctance transducer has two coils whose magnetic fields aid one another, Figure 9. In the gap between them is a movable, detectable magnetic material that, when perfectly centered between the two coils, results in equal inductances, LA = LB. Note that these coils are connected as a bridge circuit with RA and RB forming the other two legs of the bridge. R A and R B are selected to be equal and when the two inductances are equal, the bridge is balanced and output voltage Vo is zero. Should the magnetic material move closer to one coil than the other, one inductance will increase while the other will decrease, unbalancing the bridge. This imbalance results in a non-zero output voltage that can be calibrated against the position of the magnetic material. This principle has been successfully applied to make a variable- reluctance differential-pressure transducer where the magnetic material is a diaphragm subjected to the pressure differential. Pressurized deflection of the diaphragm unbalances the bridge, resulting in usable output voltage. Calibration completes the process. Two basic schemes of inductance changes can convert transformers into transducers. The first inductance-change scheme uses fixed-position primary and secondary coils
and relies on a moveable ferromagnetic core to vary the magnetic coupling, Figure 10. As the transformer core or slug moves into and out of the alternating magnetic field, the amount of coupling changes. Consequently, secondary output voltage Vo varies as well. It is apparent that upon
VoA
Time
VoB Time
Vo
Vo = 0
(a)
Time
VoA
Time
VoB Time
Vo Time
(b) V oA Time
VoB Time
Vo Time
Fig. 12. Amplitude of output voltage indicates how far slug has moved from null. Output phase indicates which direction from center or null slug has moved; (a), slug is above center, (b), and below center, (c).
T R A N S D U C E R S PA R T 1
proper calibration, the device is a position transducer. Furthermore, there is nothing to stop a user from mounting the core on the end of a Bourdon tube, for example, to create a pressure transducer. The commercial version of the moving-core variable transformer, called a linear-variable-differentialtransformer (LVDT), uses two secondary windings, Figure 11. The secondary windings are connected so that with the slug positioned precisely at the magnetic center of the flux field, the two secondary voltages will be equal but of opposite phase for zero output voltage. With the slug magnetically centered, output voltage of the A winding is the same amplitude as output voltage from the B winding but of opposite phase. At every instant of time, the sum of the two voltages is zero. When this is the case, the center position is called the null point. Now if the slug moves up from center, there is greater magnetic coupling between the primary and the A secondary and less between the primary and the B secondary. This results in non-zero output V o with the same phase as the A- winding voltage. On the other hand, if the slug moves down, the B-winding output voltage is greater than the Awinding output — again creating non-zero total output voltage Vo with a phase the same as that of the Bwinding voltage. Finally, see that the amplitude of the output voltage is a measure of
0 Stator
Vo Vmax cos
Rotor
Brushes and slip rings
Fig. 13. Angular orientation of a resolver rotor is changed relative to magnetic axis of stator. RMS value of output voltage is related to the cosine of the shaft angle.
Fig. 14. Resolver with a resolver-to-digital converter can interface directly with parallel port of digital computer.
Machine slide
A
Resolver
Parallel output Resolver -to digital converter
Computer
how far the slug has removed from its center position, while the phase of the output indicates whether the slug has moved from center in one direction or the other. The output voltage signal undergoes a 180° phase change when the null is passed, Figure 12. The second inductance-change scheme uses either a moving primary or secondary coil to cause a change in the degree of magnetic coupling between the two coils, Figure 13. Here, the variable transformer has been made so one winding can move. The most common version of this is the family of devices called synchros and synchro resolvers. By varying the distance between primary and secondary windings, the amount of flux linking the two windings changes; consequently, output voltage varies with separation distance. In the synchro, while the distance remains fixed, the angular orientation of the rotor axis changes with respect to the magnetic axis of the stator. Synchros are used as rotary position transducers, or, with rack-and-pinion gearing, as linear position transducers. Because of the trigonometric relationships between the rotor and stator and with dual windings on both, synchros can solve trigonometric problems. Synchros find use as linear-position transducers in machine tools that require precise position measurement but this usually involves resolver to
digital data conversion. In these systems, the resolver is the basic transducer, but instead of processing the output with another resolver as in the classical implementation, resolver output is fed into a special-purpose AC analog-to-digital converter that converts resolver output into an equivalent binary number representing the angular position of the resolver shaft. The digital signal then can be fed directly into a computer or other digital receiving device. The schematic of this system, Figure 14, is configured for position control. Change in capacitance The geometry of a capacitor, Figure 15, and the properties of its dielectric affect its capacitance. A people detector can be built using this type of transducer by exciting a large, door-sized pair of capacitor plates
Conducting plates
r
Dielectric material
Conducting plate
Fig. 15. Simple parallel-plate capacitor has two conductors separated by dielectric material.
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with high-frequency AC voltage. If there is free air only between the plates, the capacitance will be a certain value. When a human walks between the plates, the dielectric material changes; because the relative permittivity of a human is higher than the permittivity of free air, the capacitance changes. Connected into an AC bridge-detection circuit, it is possible to detect the presence of a human within the doorway and with calibration, to determine the bulk volume of the human, thus distinguishing small people from large people. This transducer would not be very good at distinguishing a small cow from a person, however. Some proximity switches use this permittivity change principle and change capacitance with the presence of a finger on a key on a keyboard. Because the switch has no moving parts, its life is unlimited. A change in physical dimensions, that is a change in plate-to-plate separation or plate area also changes capacitance. One type of pressure transducer, for instance, has one capacitor plate fixed while the other is a diaphragm exposed to pressure fluid. The pressure deflects the diaphragm, reduces the separation distance and increases capacitance. The change in capacitance is calibrated against the pressure. Summary of inductive and capacitive transducers Note that all inductive and capacitive transduction principles require AC excitation; DC excitation does not work. This places a special requirement on signal conditioning equipment mentioned in the discussion about LVDTs. Recall that the amplitude of the AC output voltage reflected the distance that the slug had to move from center, and the phase of either 0° or 180° indicated the direction of movement. If it is necessary to convert AC output voltage to an analog DC voltage for a subsequent receiving device, simple rectification of the AC signal will lose track of the algebraic sign. In the case of the LVDT, we could not tell an upward displacement from
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a downward displacement of the slug, Figures 11 and 12. Recall the discussion about phase-sensitive demodulators. These demodulators rectify AC signals, create a proportional DC output, look at the phase of the AC input, determine if it is 0° or 180°, and attach an appropriate algebraic signal on the DC output. In all transducers that use AC, the generic name for the signal that is generated is a suppressed-carrier, double sideband amplitude modulated signal. The frequency of the AC source, which may be several kHz or MHz, is called the carrier frequency, and the information of interest changes the amplitude of the AC voltage. That is, the information is carried in the amplitude of the AC signal; hence oscillator frequency is called the carrier frequency. The importance of the suppressed carrier can be seen by deriving the equations for the instantaneous output signal: they show that there are only so-called sum-and-difference frequencies. The carrier frequency itself does not appear and is said to be suppressed. When coming across such terminology, it is helpful to know the origin of the words. In the vernacular of the day, the bottomline information to be understood includes: ● whether AC is used by choice or necessity, the resulting output signal is suppressed-carrier, double-sideband, amplitude- modulated ● the magnitude of the measurand is reflected in the amplitude of the modulated AC output voltage, but the algebraic sign is carried in the phase of the output voltage, and the carrier voltage undergoes a 180° phase shift as the transducer passes through null or zero ● conversion of the suppressed-carrier, double-sideband, amplitudemodulated signal into a fully-re versible (sign sensitive) analog DC signal requires use of a phase-sensitive modulator ● when using inductive- or capacitive-transduction principles, AC excitation is imperative, and therefore the existence of suppressed-carrier amplitude-modulation is assured. When using resistive transduction
principles, the use of AC excitation is optional. If it is used, the resulting output voltage will be of the suppressed-carrier double-sideband amplitude-modulated type, and ● the reason for using AC excitation on resistive transducers has to do with the nature of high-gain amplifiers. With AC excitation, the amplifier’s signal conditioner can easily be designed to reject low-frequency drift anywhere in the signal-processing chain. This eliminates the need to frequently re-zero the electronic equipment and yet maintains the high gain that often is necessary to process the very low level signals generated by, for example, strain-gage transducers. It is true that many commercial strain-gage conditioning modules in use today actually excite the transducer with AC. Instructions are given about how connections are to be made, and many times the user is completely unaware of the use of AC until there is a problem that requires the use of an oscilloscope for troubleshooting. Do not be surprised, then, to find a 5- or 10-kHz AC voltage where a DC voltage was expected to be. As long as the equipment works properly, use of AC may be completely invisible to the user. Commercial transducer packaging at once creates a solution and another problem. The solution lies in the fact that many transducers can be purchased with onboard signal conditioning while others can be purchased without signal conditioning. The problem exists because it creates more options and confusion in the minds of non-expert users. It will be helpful if the transducer user is prepared to ask these questions when selecting a transducer: 1. What is the basic transduction principle? 2. What is the nature of on-board signal conditioning? 3. What is nature of the signal that is output by the package? To be sure, there are other questions that must be asked and answered, but these three will help organize the whole scene when trying to make critical decisions about selection of the components needed to form the control network chain.
T R A N S D U C E R T E C H N O L O G Y PA R T 2
Electrical generation and energy conversion T
here are two components that use these physical phenomena: the photo-voltaic cell, also called a photocell or solar cell, and the photo transistor. Strictly speaking, the photo transistor is not a true energy converter as is the photo cell, but is included because the mathematical model of the transistor has a current generator. The photo-voltaic cell is a semiconductor device that converts light energy directly into electrical energy without an intervening thermodynamic process. The methods that accomplish this are contained in the physics of semiconducting material and no detailed discussion is included here. The user simply needs to know that the process exists. The most important aspect of the phenomenon is that no external power supLight energy
Vo
Fig. 16. Photo-voltaic or solar cell converts light energy directly into electrical energy.
ply is required to obtain an electrical output response to light input. That is, if a light shines directly onto a photovoltaic surface, Figure 16, the cell will generate a voltage. If a load resistance is connected between the output leads, there will be a current and, of course, power will be delivered to the load resistor. All the power comes only from the light that impinges on the cell. The photo transistor, Figure 17, looks like a conventional NPN transistor except that no electrical connection is made to the base material. Instead, the base material is exposed to ambient light through a lens in the case. Incident light falling on the base serves the same function as the base current in a normal transistor. That is, when the incident light increases in intensity, the reversebiased collector-to-emitter sandwich conducts in the reverse direction. A basic test circuit, Figure 18, indicates that with light energy to the base totally blocked, the collector is cut off and output voltage Vo is nearly equal to supply voltage Vcc. When incident light intensity increases, the collector conducts, voltage drops across collector resistor Rc which, in turn, causes output voltage Vo to decrease. The amount of current, and therefore change in output voltage, is directly proportional to the
Any of several phenomena are used convert energy in one form directly to energy in another form. They have found use as transducers and we now address their principles of operation.
Fig. 17. A photo transistor uses incident light impinging on the base material to control collector-to-emitter current. RC Light
C
IC VCC
VO E
Ie
Fig. 18. Simple photo transistor test circuit has external resistor Rc and supply voltage Vcc. Incident light on base causes collector to conduct and output voltage Vo to fall.
intensity of the light impinging on the base as suggested in the transfer characteristic, Figure 19. An important distinction between the photo cell and photo transistor is that the latter requires an external DC power supply to function whereas the photo
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VO
Light intensity
Fig. 19. Photo transistor current and output voltage are affected proportionally by light intensity in the common emitter test circuit, Figure 18. Light-receiving photo transistor
Light source LED
Windows Rotating disk
Fig. 20. Photo-optical shaft encoder uses rotating disc with small windows to alternately turn on and turn off photo-transistor. This results in a pulse train, the frequency of which is proportional to shaft speed and whose total pulse count is proportional to incremental shaft angle.
cell does not. That means, strictly speaking, that the photo transistor is not really an energy converter in the same sense that the photo cell is. It has been lumped into the energy converter category more because the model uses a current generator to explain its function in an operational circuit. A photo transistor looks more like the photo-sensitive resistor discussed earlier. Comparisons — Certainly, the photo transistor and the photo-sensitive resistor have similarities. They both: ● require an external power supply ● conduct more heavily with stronger incident light, and ● have two electrical leads and a lens to focus incident light. Two significant differences are not at all apparent without detailed study: 1. Photo resistors conduct in either direction with incident light, while the transistor must be connected so the collector has the correct-polarity voltage consistent with transistor type; that is,
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an NPN transistor requires a +Vcc and a PNP transistor requires a 1Vcc. 2. Photo transistors switch from nonconducting to conducting much faster than photo-sensitive resistors. A transistor can switch from off to on and vice-versa in a few microseconds while the photo-sensitive resistor requires from tens to a few-hundred milliseconds to fully change conducting levels. Clearly, high-speed applications use a transistor instead of a photo-sensitive resistor. The most common use of photo voltaic cells is in solar energy conversion as an alternative energy source. Some uses also are found in the socalled electric eyes that detect the presence or absence of an opaque or refracting body. Provided that the processes are relatively slow, the photo cell or the photo- sensitive resistor work equally well. For electrohydraulic purposes, the most common use of photo-sensitive devices is in the family of encoders that detect shaft and slide position. As the windowed disc of an incremental encoder turns, Figure 20, it alternately passes and blocks light to the photo transistor, causing it to alternately cut off and conduct. By connecting the encoder into a circuit similar to that in Figure 18, the output is a train of pulses that go from zero to Vcc as the windows pass. Note that if the number of pulses per revolution of the encoder is 60, output frequency in Hz is exactly equal to shaft speed in rpm, a popular way to measure shaft speed using a digital frequency meter. With the pulse train coupled to an electronic totalizing counter, the counter records the total number of windows that have passed and is therefore directly proportional to the total angle that the shaft has turned. Be aware that in this simplistic scenario, the counter is unable to distinguish the direction of shaft rotation. To make the counter add, for example, CW pulses but subtract CCW-generated pulses, a second output called a quadrature signal must be generated. It is possible to use quadrature output to tell the counter to count up or count down so that the current value in the totalizing counter is an accurate reflection of the net shaft rotation. Installing another LED light source and another photo transistor can gener-
Main pulse train
TimeCCW
TimeCCW
Time-CW Quadrature pulse train
Time-CW
Fig. 21. The main and quadrature pulse trains are displaced 90 electrical degrees. The logic for direction of rotation can be seen by imagining time increasing to right for CW rotation but increasing to left for CCW rotation.
ate the quadrature pulse. The LEDphoto transistor sets are arranged so that when the first transistor is at the midpoint of its on-state, the second LED-transistor set is positioned to transition from on to off. The quadrature pulse, then, is displaced 90 electrical degrees or a quarter of an electrical cycle from the first pulse and is the reason the quadrature pulse is so named. Direction of rotation can be deduced using this logic: ● if the quadrature pulse changes from off to on when the main pulse is on, the shaft is turning CW, and ● if the quadrature pulse changes from on to off when the main pulse is on, the shaft is turning CCW. Encoders that produce a quadrature output often are also equipped with circuits that perform this logic. In addition, these encoders have a third signal wire that carriers direction-of-rotation information. That is, the logic is used to set a single bit which is, say, on for CW and off for CCW rotation. The three signal wires now are: ● main pulse train ● quadrature pulse train, and ● direction of rotation bit, Figure 21. Note that in the normal course of events when the quadrature pulse makes its transition from off to on, the main pulse is high or on, satisfying the first logical condition — the shaft is rotating CW. To visualize how reversing the direction of shaft rotation affects the signal, simply imagine that time increases to the left instead of to the right. Then,
T R A N S D U C E R S PA R T 2
012345678910 12 14 16 18 20 22 24 26 28 30 1 2 4 8 16 (a)
} }
Bits
Bits
(b) 012345678910 12 14 16 18 20 22 24 26 28 30 32 34 26 38
1 2 4 8 10 20
Fig. 22. Schematic codes: natural binary code (a) has 32 positions, five bits; natural Gray code (b) illustrates same thing; BCD — binary code (c) has 40 positions, six bits; and BCD — Gray- excess code (d) also has 40 positions, six bits.
(c)
(d)
when the main pulse is high, the quadrature pulse goes from high to low, satisfying the second logical condition. Output of the logic circuit is used to set the CW/CCW bit appropriately. An up/down counter totalizes the net number of pulses that have accumulated so the counter carries an indication of actual shaft position. A disadvantage of incremental encoders is that because of a malfunction or operation in an electronically noisy environment, they may pick up an extra count or lose a count in the totalizing counter. Then, indicated position does not agree with actual position and measurement error results. The error will remain until the encoder is forced to a home position and the counter is forcibly filled with the exact digital value that corresponds to that position. Absolute encoders circumvent that problem by generating a parallel-bit combination that corresponds to the actual position of the shaft. There are four codes commonly in use, Figure 22. The MA -metal A Junction 2
Junction 1
MB-metal B (a)
I≠0 E1
T1
T2
E1
(MA/MB , T1 )
(MA/MB , T ) (b)
Fig. 23. Two dissimilar metals in contact will generate a voltage. When connected in a loop without a temperature difference, the voltages cancel and there is no current.
first two codes are binary and have five tracks capable of resolving a revolution of the shaft into 32 parts (25). The last two are binary-coded decimal (BCD) codes with six tracks capable of resolving one-fortieth of a revolution. Most practical encoders have more tracks than these because with five bits, the encoder cannot resolve less than 1/32 nd of a revolution — about 11°. The principles remain the same. The problem with the natural binary code occurs at the transitions from one position to the next when more than one track has to change from high to low or vice-versa at exactly the same instant. Thus, if a digital receiving device interrogates the encoder for its current position and happens to read at a transition point, and if only one of two bits scheduled to change has already done so because of, say, manufacturing imperfections, an erroneous reading will result. The error will not be simply one increment of the most rapidly changing (LSB) track — it could be as much as on-half a shaft revolution, an intolerable situation. Furthermore, given that the encoder is interrogated millions of times in a day, be assured that a transition point reading error will occur frequently. Multiple-read scans with a vote for the most likely winner can be employed in the software provided the programmer has the necessary programming skills. The Gray code was invented to circumvent ambiguity problems with the natural binary code. The genius of this code is that the natural binary code can be rearranged so that within the interval of two LSBs, Figure 22(b), there is no
segment-to-segment transition that requires more than one bit change at a time. With this scheme, the reading error when interrogating at the transition point never exceeds the angle allocated to the LSB. The BCD codes merely interpret the position in a BCD manner rather than in pure binary. The problem with the straight BCD is the same as with the natural binary, that is there are transitions where more than one bit must change at a time. Readingerror possibilities are the same as well. The more reliable code, Figure 22(d), is the BCD — Gray excess code. Note that is has only one bit changing at a time, eliminating the transition point reading ambiguity. Thermo-electric generation Thermocouples convert heat energy directly into electrical energy. Their simplicity gives them a particularly high value; thermocouple is formed merely by putting two dissimilar metals into intimate contact with one another. Any time that two dissimilar metals are in contact, a voltage is generated at the junction that is affected by the internal energies of the electrons and the temperature of the junction. If both ends of the conductors are connected, Figure 23, there is a voltage generated at each of the junctions. If both junctions are at the same temperature, there voltages are equal and they cancel one another; there is no current. On the other hand, if one junction is at a different temperature, say junction 1 is at a higher temperature than junction 2, Figure 24, then E1 will be greater than E2 and a current will result. The power dissipated in wire resistance R wire comes directly from the energy source that causes the temperature difference, The voltage difference, that is the net voltage around the loop is E1 — E2. It is a function of metal 1, metal 2, and the temperature difference, T1 — T2. The important consideration is that voltage is dependent on the temperature difference between the two junctions. Therefore to find the temperature of junction 2 from a knowledge of the voltage, it is necessary to know the temperature of junction 1. The traditional way that junction 1 temperature is determined is to use a reference bath of known temperature.
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MA -metal A Junction 2
Junction 1 T1 > T2
T1
T2
MB-metal B (a)
I≠0 E1
E1
E2
E2
Rwire (b)
Fig. 24. If the two junctions are at different temperatures, one voltage exceeds the other and there is a current.
f
Quartz crystal
+
Cp
f
Fig. 25. Subjecting a quartz crystal to an external force causes an electric charge to be transferred from the crystal to external capacitance Cp, establishing a voltage difference across the capacitor.
For most measurements, this is an ice and distilled-water bath with ice and water existing simultaneously. Accurate measurements require that the bath be properly maintained throughout the measurement process. Tables for commercial thermocouples list the thermally induced voltage for each thermocouple junction as a function of temperature difference and the materials of the two wires. Therefore, the voltage is a measure of how far the junction temperature is above the reference junction, which is at 32° F. Maintaining the bath is a nuisance that is unnecessary with modern commercial thermocouples; they contain simulated-reference junctions and eliminate the ice bath. This author recommends that these simulated junctions be purchased by the electrohydraulic practitioner any time that thermocouples are to be used for temperature measurement. In fact, any thermocouple interface device probably will come equipped with the junctions but it would not hurt to ask just to make sure.
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Piezoelectric generation The prefix piezo comes from the Greek word piezein meaning pressure, or to press. Piezoelectricity then, is electricity that results from some sort of squeeze. The phenomenon occurs with quartz and other dielectric crystals. Simply squeezing or distorting the crystal material results in a transfer of piezoelectric charge. The crystal can be thought of as a syringe filled with electrical charges; squeezing the syringe pushes out the electrons. When a capacitor is connected to the output leads, the capacitor charges to a voltage difference commensurate with the amount of force that has done the squeezing. The voltage also is affected by the size of the external capacitor. The piezoelectric effect is used commercially to make accelerometers and certain pressure transducers noted for their extremely fast response. Bandwidths can be as high as several hundred kHz. The load on the transducer output terminals must be purely capacitive because the crystal transfers a fixed electrical charge, as opposed to generating a steady flow of charges. The capacitive load holds the charge and converts it to a voltage, Figure 25. To the extent that there is a leakage-resistance path within the capacitor, output voltage eventually decays even with constant input force to the crystal. Thus, the quartz peizoelectric transducer is not regarded as a good steady-state performer. However, it is superb for highfrequency dynamic measurements because of its quick response. These transducers are ideally suited to look at fluid-borne noise, pump ripple, and other such high-frequency events. If there is interest in the DC pressure level, it also is advisable to use one of the more conventional pressure transducers whose performance has been opFig. 26. Hall effect explains that an electron moving in a magnetic field will be pushed up in the conductor because of force f . Electrons pushed to the top generate transverse Hall voltage VH which will be positive as shown.
timized to measure this. Both dynamic and steady state performance can be monitored reliably. Hall effect In 1879, E. H. Hall showed that a current-carrying conductor immersed in a magnetic field, Figure 26, produces electron motion that interacts with magnetic flux field B in such a way that there is a force acting on the electrons that pushes them to the top of the conductor for the conditions shown. The charge separation results in a voltage difference from the top to the bottom of the conductor with the bottom being more positive for the conditions shown. This voltage difference is the Hall voltage and is directly proportional to conventional current i and magnetic-field density B. This is an important electronic relationship because, unlike the Faraday voltage that requires a timechanging magnetic flux to induce voltage, the Hall-effect voltage is generated with a stationary magnetic field and a stationary conductor. Hall effect transducers have been made commercially in a number of ways. Some, without going into great detail, are: ● non-contacting switches ● magnetic-field strength meters ● clamp-on ammeters for direct current ● position sensors that generate a voltage based on the relative position of a magnet compared to the Hall-effect sensor, and ● shaft encoders that function down to zero shaft speed. Faraday-induced voltages In 1831, Michael Faraday discovered that the induced electromagnetic voltage was directly proportional to the time rate of change of flux. Faraday’s EMF Law has produced such diverse devices as transformers, generators,
Conventional current, i
Conventional current, i Conductor
Electron
f + + + + + + + + + + + +
Magnetic flux field, B
T R A N S D U C E R S PA R T 2
Output winding
vo
vo
Rotation
Space quadrature AC magnetic field magnetic field component Magnetic axis of induced current
Aluminum drag cup (a)
Local oscillator
(b)
Fig. 27. The drag-cup tachometer is an AC output device.
tachometers, alternators, and the like. For transduction purposes, the two uses that are important are tachometers and alternators. In its basic form, a DC tachometer is a special-purpose DC generator that has a steady magnetic field (usually created by a permanent magnet) with a commutated armature immersed within the field to measure shaft speed. As the armature turns, an AC voltage is induced in the armature winding but brushes in contact with commutator segments convert the internal AC into an external DC voltage in a way analogous to the port plate and pistons in a piston pump. The reciprocating pistons produce an alternating flow but the port plate assures that the external plumbing sees only a unidirectional, albeit pulsating, flow. Similarly, the commutator and brushes also produce a ripple voltage in a DC tachometer. These pulsations limit the minimum speed that can be effectively transduced or controlled. The more precise the tachometer, the greater the number of commutator segments. Most commercial tachometers have only five commutator segments, adequate for high-speed applications but problematic when used below about 100 rpm. One of the major advantages of DC tachometers is their ability to generate high output voltages over their intended speed ranges. It is not unusual to find one that generates 10 volts per 1000 rpm. In most applications, this voltage is sufficiently high to be useful without any amplification. Another characteris-
tic is that they continue to generate a voltage down to zero speed and with some limitations and perhaps some filtering, output voltage is still usable although pulsations can be a problem. Additionally, DC tachometers are very linear because the amplitude of the output voltage is directly proportional to shaft speed. They are popular rotary speed transducers. A drag-cup tachometer is an AC device that lacks the linearity of its DC counterpart and has no commutator and brushes, Figure 27(a). A local oscillator generates an AC voltage — usually 400 Hz but sometimes higher. The stator has two windings physically displaced 90°. One winding is connected to the local oscillator and becomes the power input. Current in the powered winding induces a circulating current (eddy current) in the aluminum drag cup. With the rotor stationary, the orientation of the magnetic field of the induced current is directly in line with and opposes the causative magnetic field. As a result, there is no flux linkage with the output winding and output voltage is zero.
Local oscillator
Unipolar DC
Drag Cup rotor
Fig. 28. For unidirectional rotation, a simple diode rectifier is sufficient to convert AC into usable DC.
But when the drag cup turns, Figure 27(b), an eddy current is induced in the drag cup because of the alternating magnetic field of the powered coil. Due to the cup’s motion, however, a speed current also is induced as the conductor (drag cup) passes through the magnetic field. The combination of the induced current and the speed current causes the axis of the combined magnetic field to shift so that it has a horizontal and vertical component. The vertical component is in line with the output coil’s magnetic axis, hence there is a voltage induced therein whose amplitude is in direct proportion to the rotating speed of the drag cup, but whose frequency is controlled only by the frequency of the local oscillator. It is important to know that drag-cup AC output voltage has: ● a frequency equal to that of the local oscillator ● an amplitude directly proportional to and nearly linear with rotational speed, and ● output voltage, like that of the LVDT and synchro, is a suppressedcarrier, double-sideband amplitudemodulated signal. Therefore, if the tachometer measured rotation only in one direction and it is necessary to have a DC rather than an AC voltage, simple rectification is sufficient, Figure 28; but if bidirectional rotation is to be measured and algebraic sign needed, a phase-sensitive demodulator is necessary, Figure 29. There are systems found in aircraft and aboard ships in which conversion of a 400-Hz output into a DC voltage is unnecessary. It is possible that these systems could be encountered in the industrial environment. More likely than not, though, the conversion to DC will be made in industrial application because the servo or proportional amplifier requires DC input and feedback signals. In aircraft, the output actuator is often a 400-Hz servomotor that can use an amplified version of the 400-Hz signal without further modification. The AC servomotor is, at once, the actuator and demodulator. Note that the AC device is converted into a DC device with some basic electronic signal conditioning circuits. When the drag cup tachometer, its associated phase-sensitive demodulator,
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Reference signal Phasesensitive demodulator
Local oscillator
V0
Bidirectional DC Drag Cup rotor
Fig. 29. A phase-sensitive demodulator is required for bidirectional rotation and output algebraic sign that follows direction of shaft rotation. When all these parts are put into a single package, the result is a brushless DC tachometer.
v
N +
Time
S (a)
(b)
and the local oscillator are all packaged within the tachometer envelope, the device is called a brushless DC tachometer. All the AC running around inside is invisible. The advantage of the brushless DC tachometer is that its output is nearly devoid of ripple and the output signal is usable down to zero speed, although it is more expensive than a simple DC tachometer. Additionally, the brushless DC tachometer is slightly more non-linear that a brush-type version. Finally, when speed range is great and reliability is important, the brushless DC tachometer can be a wise choice. Alternators and AC generators are machines that generate an AC voltage. The amplitude of this voltage is directly proportional to the alternator’s shaft speed, and the frequency of this voltage is an exact integer multiple of the shaft speed. Alternators also are called AC generators, but this is technically incorrect; generators generate DC, and alternators generate AC. The kinds of alternators include: ● rotating permanent magnets with AC voltage generated in stationary wind-
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+ E Slip rings and brushes
+ Vo
+
Fig. 31. Drawing of automotive-type rotating electromagnet alternator.
+ Vo Slip rings and brushes
+ E X
+
Fig. 32. Drawing of stationary magnetic field alternator.
N
+
Fig. 30. Drawing illustrates rotating permanent-magnet alternator, (a), and output wave, (b). This wave shape is also the same for the next three types of alternators.
ings, Figure 30 ● rotating electromagnets, with AC voltage generated in stationary windings, Figure 31. This configuration requires that electrical power be supplied to the member that turns, and so slip rings and brushes carry electrical power from the stationary, exterior world to the rotating interior world. This machine is used in automotive electrical systems; the machine is not used as a transducer because the brushes and slip rings add cost with no increase in capability ● permanent or electromagnetic stationary magnetic fields with AC voltage generated in rotating coils, Figure 32. This device also requires slip rings and brushes but is not used industrially because it has no advantages over other configurations, and ● stationary magnetic fields and stationary pickup coils. This variable-reluctance alternator generates voltage by changing the reluctance of the rotating member, Figure 33. The most common commercial machine that uses this principle is the magnetic pickup, popular as a shaft speed transducer. For transduction purposes, the magnetic field is always a permanent magnet, and for construction simplicity as well as for maximum sensitivity, the stationary coil is always wound directly around the permanent magnet. The amplitude of alternator output voltage and its frequency vary in direct proportion to the speed of the input shaft, but frequency is the variable of interest when measuring data. This is true because there is an exact integer relationship between shaft speed and frequency; no errors arise in the transduction process. When measuring shaft speed, all errors occur because of the resolution of the read-out device. The transducer itself cannot be calibrated or adjusted. Only the read-out device needs such attention; when a digital frequency meter is use for the read out, it displays the shaft speed in rpm when the frequency-meter gate time is set to one sec. Thus the accuracy of speed measurement is affected only by the accuracy and resolution of the frequency meter. Although frequency generation and measurement is a popular way to determine shaft speed, all other methods us-
S
Vo
Fig. 33. Variable-reluctance alternator or magnetic pick-up is a popular speed transducer.
ing frequency to measure shaft speed introduce a time lag. As soon as actual speed is interpreted by the reader, be it human or mechanical, the value read becomes out of date. The final interpretation is the speed as averaged over the time interval. If speed is relatively constant, the interpreted value will be current. But if the speed changes, the displayed value is always out of date. This time lag complicates the control problem when using feedback to regulate and control speed. Note, then, that in spite of other flaws, tachometric methods do not produce the time lag that frequency methods do. Mechanically variable impedance devices Potentiometers and variable autotransformers, strictly speaking, are not generating devices in the sense that they convert energy. Instead, they accept a mechanical position input signal and develop an output voltage that de-
T R A N S D U C E R S PA R T 2
pends on the input position. Potentiometers can be excited with either alternating or direct current, but autotransformers always must be used with AC. A potentiometer, or pot, is a threeterminal variable resistance, Figure 34. The three terminals are labeled CW, CCW, and wiper that sometimes also is called the slider. Note that there is a fixed resistance between the CW and CCW terminals but the resistance between the wiper and either of the other two terminals varies as the wiper is moved. The u notation in Figure 35 for the mechanical input implies that the potentiometer is a rotary device, the mechanical input for the vast majority of pots. When using a pot to transduce the position of a cylinder, linear input is the obvious choice. Pot linearity refers to the constancy of the rate at which resistance changes with respect to changes in the input member. Most pots encountered in the fluid power industry are linear pots which means that nominally the slope of the resistance characteristic is constant over the entire range of input. Of course, the resistance must be measured between the wiper and one of the fixed terminals. Note that if the pot of Figure 34 is started at its fully CCW position, the resistance between the CCW terminal and the wiper nominally begins at 0 Ω and increases toward total pot resistance, Rp. At the same time, resistance from the CW terminal begins at Rp and decreases toward 0 Ω, as indicated in the plot of resistance/shaft angle characteristics, Figure 35. The slopes of the two curves of a pot designed to be electrically linear will be the same and will be nominally constant over the whole anccwo Rp
{
Wiper
cwo
Fig. 34. A potentiometer is a three-terminal resistor in which one terminal is in intimate sliding electrical contact with the resistive element.
R
CW to wiper CCW to wiper
Rp
Fully CCW
Fully CCW
Fig. 35. Resistance-shaft angle characteristics of linear, rotational-input potentiometer.
gular range of the shaft. Pots designed for linear input motions also are designed to have linear electrical characteristics. Such pots are said to be rectilinear, which means they have linear electrical properties and linear input motion. When speaking of linear pots, the implication usually is that the notation applies to the electrical-resistance curves. Therefore, in the parlance of the industry, linear pot usually refers to its electrical characteristics. To communicate the nature of the motion, more information must be offered: rectilinear pot or linear-motion pot, for example, to indicate the nature of the mechanical input movement. Furthermore, when technicians say linear pot, they probably are talking about the electrical properties and thinking of a pot with rotary mechanical input. These words cause no small source of confusion. It is wise to request clarification as to what, specifically, is meant by linear potentiometer. Here, it is only important to know that there are two common forms of input motion for pots, they will most likely have linear resistance characteristics, and they will be called linear pots. Where the nature of the mechanical input is crucial to the point at hand, some clarifying words will remove ambiguity. Mechanical input u, Figure 34, causes the slider to travel the length of the resistive element. The resistive element could be a very fine wire (most often copper) or a conductive plastic film. This film is impregnated with tiny conducting particles so that the particles are in electrical contact with one another. The plastic thus becomes an electrical conductor. The plastic material provides a bonding agent for the conduct-
ing particles as well as a smooth, lowfriction surface over which the slider can move. Pot linearity. Even though we are speaking only of linear pots, we emphasize that when resistance characteristics are examined with sufficient precision, true linearity never is achieved — only approached. When potentiometers are designed and manufactured with controlled linearity, they are usually called instrument-grade pots, clearly implying that they can be used for instrumentation functions. All such transducers are accompanied by technical data sheets that state the expected worst- case nonlinearity. Note that non-linearity actually is reported but the industry always calls it linearity. R
Rmax
Actual curve (test data)
+ R R
Best mathematical straight line
0 End resistance (a)
0 max
R
Rmax
Actual curve (test data)
+ R R
Best mathematical straight line
0 End resistance
0 max
( b)
Fig. 36. Potentiometer independent nonlinearity, (a), is reported as the deviation from best straight-line independent of end resistance and measured with a mechanical input range from mechanical zero to maximum mechanical position. Terminal non-linearity, (b), is deviation for the best straight line that passes through origin. Terminal non-linearity is always greater than independent non-linearity, but only the latter is critical.
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VS
RC1
CW
RT1
(command)
RE1 V0
}
WC WT
CCW
RT2
RE2
RP1-RE1-RE2 (transducer)
CW
CCW
Fig. 37. End resistances RE1 and RE2 of the transducer pot can be compensated in a pot-pot bridge circuit by deliberately introducing adjustable trimmer resistors RT 1 and RT2.
The two methods of measuring the non-linearity of a potentiometer are the terminal non-linearity and the independent non- linearity, Figure 36. The difference between the two is accounted for by the end resistance of the pot. That is, the pot cannot be constructed so there is no resistance between the wiper arm and the respective end terminal when the arm is located at its mechanical end limit. This is not merely the resistance of the connecting wire, it is a mechanical consideration: the resistive element must always be made slightly longer than the allowed mechanical travel. The potentiometer cannot be manufactured so the physical end of the resistive element coincides exactly with the mechanical travel of the wiper. Furthermore, when the pot is used to measure cylinder position, for instance, the mechanical travel of the pot must be greater than the mechanical travel of the cylinder. To be otherwise would invite
the cylinder to be stopped by the mechanical barrier in the pot, a duty for which it is not designed. The result is that even if the pot could be manufactured with zero end resistance, application considerations demand there be end resistance in the system. The common mechanical method used to compensate for end resistance is to place the potentiometer in a pot-pot bridge circuit. The scheme, Figure 37, graphically indicates how end resistances in the transducer side of the bridge are placed because the wiper of the transducer pot cannot mechanically reach the true electrical end of the resistive element. In reality, end resistances RE1 and RE2 are physically part of the transducer pot’s resistance element but are often drawn as discrete resistors for analytical clarity. On the other hand, trimmer resistors RT1 and RT2 are discrete, adjustable resistors that have been deliberately introduced into the command-pot side of the bridge. By way of explanation, consider that supply voltage VS is applied to the circuit. For each setting of the command pot, a corresponding setting of the transducer-pot wiper results in zero output voltage, V o , end resistances notwithstanding. If it were not for the end resistances, it would be possible to position the wiper of the transducer pot so that it electrically touches either the CW or CCW terminal. If the command pot likewise had no end resistances, it too could be positioned at its electrical limits. Thus, fully CCW on the command pot, for example, would require the transducer
PS
Error
A
i
R T1
VS
RC R T2
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Transducer pot with end resistance (feedback)
4-way valve
Fig. 38. A pot-pot servomechanism uses command and transducer pots in a bridge configuration. An error voltage causes the valve to shift, porting pressure fluid to the actuator. The actuator moves the feedback wiper until the error goes to zero.
pot to be positioned fully CCW to set output voltage Vo to zero. Similarly, a completely CW setting of the command pot would require a fully CW setting of the transducer pot to acquire zero output voltage. Now consider the actual situation where the end resistances are not zero. To compensate for them, the transducer pot must be moved to its mechanical limit in the CCW direction. That limit is a function of the potentiometer and the mechanical constraints of the load and the fluid power actuator. At the same time, the command pot must be moved to its CCW limit. If output voltage then is measured, it will not be zero because of end resistance on the CCW end. Next, trimmer resistor RT2 is adjusted until output voltage is zero. To set conditions for the other extreme, the transducer pot is moved to its CW mechanical limit while the command pot also is moved to its CW limit. Output voltage will not be zero, and trimmer resistor R T1 is then adjusted until the output voltage is again zero. Because the two adjustments are not totally independent of one another, it becomes necessary to check and recheck each end of the travel until they both produce zero output after repeated trials. Then, the electrical and mechanical limits of both pots have been forced to coincide with one another and the trimmers can be locked into their adjusted positions. The circuit in Figure 37 has practical application in a pot-pot servomechanism, Figure 38. It can be shown that output voltage Vo, the error in servomechanism parlance, is sent to a servo amplifier. Amplifier output moves the spool of a servo or proportional valve to port fluid to an actuator. Output of the actuator is mechanically connected to the transducer pot wiper shaft, so that as the actuator moves, the wiper shaft of the transducer pot moves too. The transducer pot also is called the feedback pot or the feedback transducer. When the feedback pot has moved to a position where output voltage from the bridge is zero, the valve spool centers, the actuator stops, and the system is in an equilibrium position commensurate with the command pot setting.
T R A N S D U C E R T E C H N O L O G Y PA R T 3
Physical principles of transduction T
his transducer converts electrical energy into heat energy and then into light energy. The process is called incandescence and was first made practical by Thomas Edison. The most common device that uses this principle is the household incandescent light bulb. Of the total power that enters the conductor, less than 1% is converted to light energy; most becomes heat, so as an electrical-to-light energy converter, the incandescent bulb’s inefficiency makes it less than ideal. Instead, certain solid state electronic devices that avoid heat generation are favored in modern electronic equipment. There are certainly a lot of light bulbs still in use today, but use of the incandescence principle is unpopular in electronic control system designs. Gas ionization When certain gases are subjected to a high potential difference, the gas ionizes, becomes a conductor, and emits visible light. Two of the more popular gases used for this type of transduction are neon and xenon. Gas ionization provides a more efficient electric-tolight conversion because the temperatures are lower than in the incandescent process.
The investigation of transduction continues with an examination of light generation. When electricity flows in a conductor, some energy is converted into heat. This heat energy raises the temperature of the conductor and, if the power is sufficiently high, conductor temperature can increase to a point where the conductor glows. Its temperature is affected directly by the amount of power it dissipates. The intensity and wavelength of the resulting light is directly affected by its temperature.
LEDs Light-emitting diodes (LEDs) are gallium-arsenide PN semiconductor junctions that emit visible light. Their unique characteristic is that they directly convert electrical energy into light energy without an intervening thermal process. The action within the junction is the direct opposite of that which changes the resistance of a junction when light energy impinges on it. Current through the junction in the forward direction of an LED causes the electrons to jump to higher energy bands within the atom. When they subsequently return to their original bands, the energy so expended is emitted in the form of visible light. LEDs are popular as indicators in electronic equipment because they operate at low temperature, are very small, and are efficient, because there is no intermediate thermal energy conversion process. Electrical resistance and current Electrical current in a resistive conductor is converted from electrical to thermal form. The amount of power, a measure of heat energy, is P=I 2 R, where P is the power in watts, I is the current in amperes, and R is the resistance in ohms. In the field of instru-
mentation, the principle is put to work to create a true RMS voltmeter. In a true RMS voltmeter, voltage is applied to a calibrated and thermally insulated resistor. The voltage and its attendant electrical current result in a power dissipation that elevates the temperature of the resistor. Then, the temperature rise of the resistor is measured and related to the unknown voltage. Of course, RMS, a measure of equivalent heating, is the basis for the correlation. The advantage of the instrument, in addition to being a true RMS- measuring device, is that it yields the RMS value of any wave shape even though the instrument can be calibrated with DC. Thermoelectric cooling — In 1834, French watchmaker Jean Peltier was duplicating Thomas Seebeck’s 1821 thermocouple experiments, when Peltier discovered that putting an electrical current into a thermocouple junction resulted in a temperature difference. Seebeck’s thermocouple effect caused a current in a two-metal loop when the extreme ends of the loop were at different temperatures; he discovered the direct electric-to-thermal energy conversion process. Peltier, on the other hand, discovered that the process is entirely re-
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x
Pin Fig. 39. When internal pressure in a Bourdon tube pressure gage straightens the tube, the tip moves in direct proportion to the amount of pressure.
versible; by inputting electrical energy, the junction can be cooled. Commercial use of the Peltier effect is limited to small, non-moving-part refrigerators that operate from a DC source. If the aim is to create a refrigerator, it is important that the amount of input current not be excessive. That is, there is a current level for the thermocouple where the amount of cooling is greater than the heating because of electrical current in the junction resistance. If the current is excessive, heating overpowers the cooling.
fL
Fig. 41. The proving ring is a metal ring that deforms on application of external load source fL. Once deflection x has been calibrated against known load forces, other x deflections are transduceable reflections on unknown loads. To O
Ti
Fig. 42. Applied torque causes the shaft to twist, resulting in an angular difference between the input and output ends of the shaft. The amount of twist can be calibrated against known torques.
, , ,,,,
Mechanical reaction All solid materials are elastic, a phenomenon indicated by their modulus of elasticity. When these materials are subjected to external stress, they deform. This well-known physical principle has been put to work in transducers. The most common implementation in fluid power is the Bourdon tube pressure gage, Figure 39. The instrument is made by bending a piece of tubing into semicircular form. When pressure is applied at the input port, the internal pressure causes the tube to straighten, deflecting the tip. Calibrating tip movement to known pressures, the device becomes a useful pressure transducer. Note that input is fluid pressure while output is mechanical displacement.
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Other types of transducers use the material modulus of elasticity principle as well. To measure pressure, diaphragms, bellows, and simple hollow tubes are used. Pressurized fluid, introduced into the fluid cavity of a simplified diaphragm pressure transducer, Figure 40, acts on the diaphragm. The diaphragm deflects because of the force generated by pressure fluid and the deflection is imparted to a cantilever beam through a small push pin. Deflection of the cantilever can be calibrated against known pressures to create a useful pressure transducer. The commercial version of this concept converts cantilever deflection to an electrical output by mounting strain gages on the beam. Note in Figure 40 that the top of the raised cantilever beam is in compression while the bottom of the beam is in tension. With two strain gages bonded to the top and two bonded to the bottom of the cantilever, the gages can be interconnected electrically to form a bridge circuit. The two tensioning gages are diagonally opposite one another in the bridge while the two compressing gages occupy the complimentary diagonals. Thus, all four gages are active and all are affected by input pressure, Pin. This construction offers maximum sensitivity. Force transducers are called load cells. One type has a proving ring that carries and transducers the unknown force, Figure 41. Applied external load force fL causes the ring to become elliptical. Cross-center distance x is a function of the amount of force applied and can be calibrated; x may be converted to an electrical signal with an LVDT or
Cantilever beam
Push pin
Diaphram
Fluid medium
Pin
Fig. 40. In one version of a diaphragm pressure transducer, the pressure-deflected diaphragm imparts motion through a push-pin to a cantilever beam. Deflection of the cantilever is calibrated against inlet pressure.
with strain gages mounted on the inside and outside surfaces of the ring and on the axis that is transverse to the direction of applied force fL. Torque transducers, also called torque shafts, are used to measure the torque transmitted in a rotating shaft. The metal of the shaft twists in the presence of the transmitted torque, causing an angular displacement between the input and output ends of the shaft, Figure 42. During the calibration process, the amount of angular twist or wind-up, is measured for each known torque and can then be used in a measurement situation to infer the amount of an unknown torque. In commercial versions of torque tubes, the angular deflection has been transduced using strain gages or LVDTs. Thermal expansion — When materials undergo a change in temperature,
,,,, ,,,,,,, ,
T R A N S D U C E R S PA R T 3
X
Material 1
Material 2
(a)
(b)
Fig. 43. Two materials with different temperature coefficients of expansion form the bi-metal thermometer assembly. During manufacture, (a), the assembly is straight but when exposed to elevated temperature, (b), greater expansion of material 1 causes the assembly to bend. Tip movement x is a measure of temperature change. Sealed, evacuated capillary
Elevated temperature
Room temperature
Liquid (a)
(b)
Fig. 44. When the thermally expandable liquid in the liquid-in- glass thermometer, (a), is exposed to an elevated temperature, the liquid rises in the sealed-andevacuated capillary tube giving an indication of temperature change, (b).
a change in the physical dimensions of the body accompanies that change. It is usually an expansion with a temperature increase, and occurs in liquids, gases, and solids. This principle has been used in temperature transducers; such bi-metal thermometers consist of two dissimilar metals bonded together, Figure 43. Their major dissimilarity is their respective temperature coefficients of expansion. When the bonded pair is exposed to a change in temperature, one expands more than the other to bend the assembly. The degree of bending can be calibrated for use as a
temperature indicator. The advantage of bonding two materials is that the amount of bending is greater than the simple expansion or contraction of either material alone. Thus the effect of the temperature change is apparent more readily in the bending of the assembly than it would be with only one material. Mounting strain gages on either side of the bimetal strip and then forming them into a four active-gage bridge circuit converts mechanical instrument output to an electrical signal. An LVDT or other position-sensing transducer might be connected to the tip. Liquid expansion is used in the familiar household liquid-in-glass thermometers. Alcohol-based solutions and mercury are liquids of choice because they have very low freezing temperatures; this gives the thermometer a large temperature range. The expansion liquid is contained in a relatively large bulb at the bottom of the transducer. The bulb is attached to a thin capillary tube that has been evacuated and sealed. The amount of liquid used in the manufacture of the instrument exceeds the volume of the bulb by a slight amount at room temperature so the liquid reaches up into the capillary at that temperature. When the bulb is exposed to an elevated temperature, the expanding liquid rises farther in the evacuated capillary tube, and with calibration, becomes a useful temperature transducer, Figure 44. Although it is difficult to convert this output to an electrical signal, it has been done by immersing a resistive element into the capillary. Then the liquid shorts out a portion of the resistor so that resistance changes with temperature. Gases also expand in the face of rising temperature and this principle has been put to work to make temperature transducers. Construction consists of filling an evacuated bulb with a gas (usually nitrogen because of its low freezing temperature), then connecting the bulb to a pressure transducer through a capillary tube. As the temperature of the gas changes, the internal pressure changes, reflecting the temperature of the bulb. Electromechanical energy conversion In electromechanical energy conversion, electrical energy is first stored in
an electromagnetic or electrostatic field. Then, movement of an output member extracts some of the field energy to do useful mechanical work. The electromagnetic field surrounds a current-carrying conductor and is commonly called a magnetic field. In the electromechanical energy conversion process, this is the most commercially predominant method of energy conversion. Note also that the energy contained in the magnetic field is directly proportional to the amount of magnetic flux in the field, and that the amount of flux can be increased by forming the current-carrying conductor into a coil so that the flux produced by one wrap of the coil adds to that of the next wrap, and to the next, and so on. The flux can be further increased by providing it with a path that has low reluctance (resistance) with the most common path material being iron. The resulting device is an iron-core inductor, but somewhat complex in that at least one part of its iron core is able to move. Electrostatic fields for the most part are invisible to non- technical observers and are mostly misunderstood even when their effects can be seen. These fields exist because a positive electrical charge is separated from a negative charge — a charged capacitor. The electrostatic field exists in the dielectric between the capacitor plates and is substantially confined to that region. One plate is positively charged and the other is negatively charged, creating the necessary and sufficient condition for an electrostatic field. This kind of field also exists after walking across a carpet when the humidity is low. The arc and shock that occur when a metal object is touched is palpable evidence of the energy that is stored in the electrostatic field. The shock arises because of the current that results from the electrical charges being redistributed between the charged body and the neutral metal. The heat of the arc dissipates the energy that was stored because of charge separation. Passing a comb through your hair also picks up electric charges and the comb is able to attract small pieces of paper. Lifting the paper against gravity constitutes an electromechanical energy conversion because work is done on the paper. It is possible to harness
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the energy of the electrostatic field to make workable motors but they have not met with commercial success because the ability to: • store sufficient energy requires a large number of capacitor plates, and • generate sufficient force and torque requires extremely high voltages. When making the calculations needed to build machines of a practical speed, power, and torque capacity, the machine either becomes very large, the voltage becomes excessive, or both. Electromagnetic energy conversion, therefore, has become much more popular than electrostatic energy conversion and is certainly more important to practitioners of electrohydraulics. Quartz crystals generate an electrical
charge when subjected to an external force. The process is reversible. That is, when subjected to an external voltage, the crystal will change its shape or generate a force if its motion is constrained. To apply this phenomenon, the quartz is formed into helical coil that produces a useable motion of the tip and provides a force on the order of a few pounds. Quartz motors have been used as valve operators, but are not commercially popular. Of the electromagnetic energy converters, there are many in regular use in today’s industrial environment. Motor types include induction, synchronous, DC, stepper, brushless DC, and torque. Linear machinery includes AC, DC, and proportional solenoids,
linear-force and linear-induction motors. All have found use in fluid power applications, but the few that are significant in the electronic control of hydraulic systems are torque motors, proportional solenoids, and linear force motors. These three are the links between the electronic world and the hydraulic world. Torque motors, with only 1° or 2° of angular rotation, are popular on servovalves and also can be found on some proportional valves. Proportional solenoids got their name from their application to that class of continuously variable, electrically modulated control valves commonly called proportional valves. Linear force motors are used on both servo and proportional valves.
Mechanical reaction to fluid motion
F
rom an instrumentation and measurement point of view, mechanical reactions to fluid motion can indicate the rate at which a volume of fluid passes the measurement point. The generic name for the instruments that do this is flowmeter. A wide variety of physical principles has been put to work to create commercially viable flowmeters. Each has its strengths and weaknesses, its advocates and detractors. In spite of, or perhaps because of, the great number of different kinds of flowmeters, precise flow measurement remains a challenge. The state-of-the-art in flow measurement accuracy (actually, the correct term is error) is about ±0.05%. That is, no one in the entire world can measure flow more accurately than that and even if they could, there is, as yet, no way to prove it. Flow measurement is done routinely in the laboratory but is not widely incorporated in the industrial control of fluid power machinery. Cost is one deterrent and physical size is another. Of course, insertion of a flowmeter into a system requires that the plumbing be disassembled, the flowmeter plumbed in, and the plumbing reassembled. However, as electronic controls become more widespread and commonplace, there will undoubtedly be increased use of flow transducers in fluid power machinery. Each of the wide vari-
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ety of flow-transduction principles generates its own kind of signal, has its own response, and requires its own kind of signal- conditioning equipment. Positive displacement Positive displacement is the basic means of energy conversion in fluid power. The term describes the process in which motion of a rigid member sweeps or positively expels fluid when making its motion, such as a piston that slides within a close-fitting cylindrical bore. Conversely, when fluid is forced into a cylinder, the piston moves positively. This is the basis of a positivedisplacement motor. In fact, positivedisplacement hydraulic motors can be successfully used as flow transducers after appropriate calibration. The positive-displacement process can be achieved using several mechanical schemes, namely pistons within bores, meshing gears, vanes, nutating discs, abutments or any combination of these. All have been used in fluid pumps, motors, and flowmeters. A positive-displacement flowmeter is actually a hydraulic motor designed to have low internal leakage and low starting torque relative to its torque capability. The nutating disc is another design principle. It accomplishes its positive displacement by virtue of a disc
that nutates (like the nodding of the head), alternately accepting influent on one side and expelling it on the other. The nutation or wobbling is imparted to a shaft that turns so the total angular excursion of the output shaft is a reflection of the total volume of fluid that has passed during some metering time interval. Output shaft speed, on average, is a measure of the volumetric rate of fluid transfer (flow) between the inlet and outlet ports. A tachometric transducer at the output shaft produces an output voltage or frequency proportional to flow. When frequency data are fed into an electronic counter, the counter will hold a digital value indicative of the total amount of fluid that has passed through the meter since the electronic counter was last reset. This meter-counter combination is commonly called a totalizing meter or an integrating flowmeter. The most common use of nutating-disc flowmeters is in home water-metering systems. In most older homes, the totalizing counter is a geneva mechanism that mechanically displays the total gallons of water that has passed through the meter. Spur-gear sets also have been used to implement the positive- displacement function. An oval gear set, Figure 45, is the internal mechanism in some flowmeters. The meshing gears, along with the
T R A N S D U C E R S PA R T 3
Q
Fig. 45. The gear shape of a positive-displacement, oval-gear flowmeter provides a large displacement in a small package. PA A
RLAC
Case drain
Outside motor Inside motor RLPP
RLBC B
PB
Fig. 46. When PA is greater than PB, there are resistances to leakages through paths RLAC (A port-to-case) and RLPP (port to port) that do not pass through the positive metering elements. These result in measurement error.
small clearances at the gear tips and along their faces, produce the positive displacement. The reason for the oval-gear shape is that displacement per gear-shaft revolution is larger than it would be with a circular gear. In fact, the oval-gear set can be viewed as two gears, each with only two effective teeth. The actual gear teeth, which are relatively small, merely serve to keep the two gears synchronous. A drawback of this instrument arises from the fact that the rate of volume transfer (instantaneous volume change per degree) varies considerably with gear position. This results in a non-uniform instantaneous shaft speed even when the inlet flow is perfectly constant. If one is interested only in total volume metered over a substantial time interval or the average flow, the pulsating speed is in-
consequential. On the other hand, if you need to know the instantaneous or dynamic flow, the instrument is deficient. A simple spur-gear hydraulic motor is a better dynamic flowmeter provided that it has a tolerably small internal leakage and low running pressure. Positive-displacement flowmeters should always be used in the low- pressure lines of a hydraulic system. One reason is that most of them have a low casepressure capability. High-pressure hydraulic motors are not necessarily limited by their case pressures, but they still must be used in the low-pressure lines. This is because there is always an internal component of leakage that finds its way from the high-pressure inlet port to the motor case without passing through the positive-displacement elements within the motor. Figure 46 shows a simplified schematic of a typical hydraulic motor that is constructed on an ideal positivedisplacement section plus some leakage paths. If the differential pressure across the meter (motor), PA to PB, is small, then the leakage component through port-toport leakage path RLPP, (leakage resistance, port to port) is similarly small. Of course, the running torque of the motor controls this differential pressure. Thus, the best positive-displacement flowmeter will have low friction, low running torque, negligible load on the output shaft and because of this, a low differential pressure drop. This keeps the port-to-port bypass leakage small. However, if the differential pressure is low, it is still possible that PA and PB could be high. In that instance, there can be a substantial leakage component through the RLAC (leakage resistance, A port to case) path, assuming that the case drain pressure was low. The component of leakage through RLBC, (leakage resistance, B port to case) has been metered so it does not matter. The case drain port can be maintained at a high pressure simply by blocking the port, but that requires a high-pressure case and shaft seal. That type seal could induce undesired friction on the shaft and require a higher differential pressure. It is better to place the flowmeter in the low-pressure lines, because any flow that enters the flowmeter but does not have to pass through the displacement elements constitutes a measurement error in the amount of the bypass flow.
Turbines and propellers When moving fluid engages the pitched blades of a turbine flowmeter, Figure 47, a torque is created that turns the turbine. The turbine blades usually are made of a magnetic material to facilitate transduction. It is common industrial practice to include a magnetic pickup coil that generates a voltage pulse each time a turbine blade passes and alters the reluctance of the magnetic path. Output voltage frequency varies in an integer relationship to the rotational speed of the turbine. The speed of the turbine is directly related to the volumetric flow through the meter but that relationship must be determined by calibration. The calibration curve for a turbine flowmeter, Figure 48, is typically presented as the K factor, a function of independent variable ƒ/n, on semilog graph paper. This Universal Viscosity Curve is typical of most turbines. The range on the horizontal axis is about two decades (100:1), stretching from about 20 to about 2000. The horizontal-axis limits correspond to K-factor limits of about 6580 pulses per gallon to a peak of about 6800 pulses per gallon. That is, as the user of this instrument changes the ƒ/n ratio in the application from 20:2000, the turbine will deliver 6700 pulses for each gallon that passes through it, within ±110 pulses. We would say then that the meter has a constant K factor of 6700, within ±110 counts (±1.5%) over the 100:1 range on the horizontal axis. On the other hand, if use of the instrument is confined to the range of 10:100 horizontally, then the K factor is constant within perhaps ±0.1%. On some turbines, the range on the ƒ/n axis may be as high as 1000:1, or three decades. The user must determine the proper interpretation when using these graphs because the simplistic conclusion is that the flowmeter has a useful flow measuring range of 100:1 for the instrument whose curve is given in Figure 48. This is not the case, which best can be understood by considering the test procedure. The test is done at several different flows, and at two, three, or perhaps as many as four different viscosities. That is, a test is run at some high viscosity and a 10:1 flow range. This describes the curve say, from 7 ƒ/n to about 70 ƒ/n. Then the viscosity is reduced (by using a different fluid) and the test is run again with a 10:1 flow range that will be used to describe
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T R A N S D U C E R S PA R T 3
the curve from 70 ƒ/n to 700 ƒ/n, and so on. The correct interpretation of the curve is that it is a measure of the sensitivity of the instrument to changes in viscosity, not the useful flow rangeability. The flowmeter industry uses the term turn-down ratio. This refers to the ratio of the maximum flow that can be measured to the useful minimum value of measurable flow. When making tests at a single viscosity, the turn-down ratio of a typical turbine flowmeter is about 10:1 in the best instruments, and may be as low as 5:1 in some others. When operated outside that range, the constant K factor is not constant at all, and can lead to measurement error if not corrected. This limitation is not at all apparent from the graph of Figure 48. Turbine flowmeters are popular instruments in fluid power systems and laboratories because they offer advantages over other types. First, they can be used in high-pressure lines up to the pressure rating of the envelope; second, they are much smaller than many of the other choices; and third, they are inexpensively adapted to electrical output with the magnetic pickup coil. To the extent that the integrity of the bearings is maintained, a turbine flowmeter will maintain its calibration very well. Turbine blade shape must also be maintained, therefore their use in contaminated fluids can lead to serious data errors because of blade and bearing deterioration. The major disadvantage of turbine flowmeters is that the output data is always frequency. Electronically, it is impossible to convert from frequency to either digital or analog values without a time delay. Therefore, even though the turbine may be very fast to respond to changes in flow, the electronic signal conditioning always adds delays to the readout value of flow. The time delay in a feedback control system always makes the control design problem more difficult. Flow-to-pressure conversion In flow-to-pressure conversion, the kinetic energy contained in the moving fluid is converted to potential energy. That potential energy results in a pressure difference. The measurement problem then becomes one of measuring pressure, but of course the pressure must be calibrated against known flows through instrument — a flowmeter.
A/186
Turbine wheel
Pickup Coil
Fig. 47. When moving fluid impinges on the turbine blades of this flowmeter, it creates a torque that turns the turbine. A magnetic pickup generates a voltage each time a turbine blade passes. Average K value
6700
K factor 6700 6600 6400
+- limits usually15 but < 22.5 in.-Hg., and N = 3 for vacuum > 22.5 and up to 26 in.-Hg. One further complication: pump capacity in the equation is not the open capacity (capacity at atmospheric pressure) usually cataloged by manufacturers. Instead, it represents the average capacity of the pump as system pressure drops to the final vacuum level. This value is not readily available but can be approximated from pump performance curves, furnished by manufacturers. These curves plot pump capacity at various vacuum levels. To mesh these curves with the equation, simply substitute values in the equation using pump capacity readings from the curve at various vacuum levels at 5-in.-Hg increments, up to the desired level. Then total these times. Finally, note that this pumpdown time is based on all system components operating at optimum levels. A 25% additional time allowance is recommended to compensate for leakage.
inches-Hg for either continuous or intermittent duty cycles, and can be obtained from pump manufacturers. Because the maximum theoretical vacuum at sea level is 29.92 in.-Hg, actual pump capabilities are based on and compared to this theoretical value. Most pumps don’t come near this maximum value because of high internal leakage. Depending on pump design, the vacuum limit ranges from 28 to 29.5 in.-Hg or about 93% or 98% of the maximum theoretical value. For some pump types, the maximum vacuum rating will be based on this practical upper limit. For others, where heat dissipation is a problem, the maximum vacuum rating might also take into account allowable temperature rise. Pump ratings must be adjusted for operation under local conditions. An adjusted maximum pump rating can be calculated from this formula: Ra = Rc (Pa) 4 Ps where: Ra is adjusted pump rating Rc is catalog pump rating P a is ambient atmospheric pressure, and Ps is standard atmospheric pressure. The formula can be simplified by rounding standard atmospheric pressure to 30 in.-Hg. For example: Assume that a pump has a maximum vacuum rating of 24 in.-Hg, but ambient atmospheric pressure is only 22 in.-Hg. It is obvious that the pump cannot achieve its maximum rating. The formula indicates how much vacuum it can produce: Pa= (24)(22)430 = 17.6 in.-Hg. Air flow requirements The second major step is establishing air flow requirements of a vacuum system that must be matched by pump performance. To do this: 1. Determine the amount of free or atmospheric air to be removed from all work devices connected to the system. Then select a pump size that will deliver that flow of air at the required vacuum level. The required air flow is determined by adding the volumes of free air that must be removed from each work device in the system during each work cycle. These values are
VACUUM TECHNOLOGY
found from calculations based on handbook formulas, catalog specifications, performance curves, or actual tests. 2. Multiply the total volume of air that must be removed by the number of work cycles per minute. This gives the required air flow (in cfm) that the pump must be able to remove. Note that these calculations should be based on peak system requirements. The optimum pump size for a given application is determined by comparing the rate at which air must be removed from the system with the capacities of various commercial pumps available. Flow capacities are usually given in cfm of air. But as vacuum increases, flow decreases. Pump manufacturers’ performance curves or tables show cfm of free air pumped while operating at rated speed against inlet conditions ranging from 0 in.-Hg to maximum vacuum capacity. The rated capacity of any pump will be highest at 0 in.-Hg (no vacuum) and will drop rapidly as the vacuum level is increased. Compared to those of air compressors, power requirements of vacuum pumps are relatively low. The horsepower should meet the peak power requirements of the pump. In general, that should be sufficient to ensure satisfactory operation under all rated operating conditions. Of course, the vacuum pump must also have adequate energy to overcome friction and inertia during start-up. As a rule of thumb, about 1 hp is needed for each 20 cfm of air pumped. The power requirements for compressed air-driven vacuum generators is even lower. Many generators can be cycled on and off depending upon when vacuum is needed. Vacuum pumps are available as either motor-mounted units or units for use with a separate drive. Choosing a motormounted unit with an integral motor greatly simplifies drive selection. With these units, the pump is literally built
What about high altitudes? Atmospheric pressure determines the maximum vacuum force that can be achieved. And standard atmospheric pressure at sea level is 29.92 in.-Hg. But what happens at locations a mile above sea level? What is the atmospheric pressure there? When working with vacuum, that’s important to know. Why? Because the maximum vacuum force that can be generated in locations above sea level will be less than 29.92-in.-Hg. The force will be limited by the ambient atmospheric pressure. Vacuum pumps have maximum vacuum ratings based on sea level conditions. They must also be re-rated for operation in other locations. The first step is to calculate local atmospheric pressure; a fairly easy around the motor. Advantages include assurance of proper motor sizing, compactness, and cost savings. There are cases when a separate drive should be considered. If no available motor-mounted unit meets other selection criteria, of course, a separate drive will be necessary. A separate drive might also be more economical if an existing power source — a gasoline engine or electric motor — will always operate when the vacuum pump is needed. Drive efficiency must be considered as well. Different techniques have been developed to evaluate the energy efficiency of vacuum pumps. Some have been incorporated into standard industry test codes, and many pump manufacturers catalog their test results, including brake horsepower (actual hp) and flow vs. vacuum level. In addition to actual brake horsepower for various vacuum levels, the catalog power requirement generally will list a drive speed along with the horsepower requirement.
procedure. The greatest determining factor in atmospheric pressure is altitude; the higher the altitude above sea level, the lower the pressure. Use this rule of thumb: for every 1000 ft. of altitude above sea level, atmospheric pressure drops by 1 in.-Hg. Using rounded-off figures, for Denver’s elevation of 5280 ft, the atmospheric pressure is about 25 in.-Hg. To adjust a pump rating, think of that rating as a percentage of atmospheric pressure at sea level. If a pump is rated for 25 in.-Hg, it can achieve 83.4% (25 ÷ 29.92) of a sea level perfect vacuum. In Denver, that same pump can achieve 83.4% of 25 in.-Hg — the local atmospheric pressure — or a vacuum of 20.85 in.-Hg. The required pump speed to develop the various rated capacities also will be given. Effects of temperature Temperature considerations are very important in vacuum pump selection. High external or internal heat can greatly affect pump performance and service life. Internal pump temperature is important to consider, because as vacuum level increases, the pump itself absorbs more of the heat generated. Heavy-duty pumps with cooling systems are often required for high vacuum applications. But light-duty pumps can operate at maximum vacuum for short periods of time if there is an adequate cool-off period between cycles. The pump experiences a total temperature rise as a result of all the heat sources acting on it — internally generated heat plus heat for internal leakage, compression, friction, and the added load of external ambient temperature.
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Basic circuits A
number of circuits are used frequently in fluid power systems to perform useful functions. For example, metering circuits offer precise control of actuator speed without a lot of complicated electronics, decompression circuits reduce pressure surges within a hydraulic system by controlling the release of stored fluid energy, and pump-unloading and regenerative circuits make a system more energy efficient. Other circuits are designed for safety, sequencing of operations, and for controlling force, torque, and position.
Still other circuits enhance the application of specific components, such as pumps, motors, accumulators, filters, and airline lubricators. The circuits appearing on the following pages are provided as a resource of general ideas. They may be used as: ● an educational resource to aid understanding of circuits already in use ● a starting point for new designs, and ● as a modification to enhance operation of existing equipment. They certainly do not have to be implemented as shown. In fact, many of the circuits use purely mechanical com-
ponents, so incorporating them into new or retrofit applications may involve integrating electronic feedback and control into the circuit as a modern alternative to mechanical control. However, many existing and new applications still gain the greatest benefit from mechanical control — especially those applications where electricity could pose a threat to health and safety. However, whether using mechanical control or electronic, perhaps the greatest benefit may be gained by customizing one of these circuits to serve the specific requirements of an application.
Basic circuits contents Accumulator circuits . . . . . . . . . . . . A/197 Air lubrication circuits . . . . . . . . . . . A/199 Cylinder locking circuits . . . . . . . . . . A/200 Deceleration circuits . . . . . . . . . . . . . A/202 Decompression circuits . . . . . . . . . . A/204 Electrical control circuits . . . . . . . . . A/206 Filter circuits . . . . . . . . . . . . . . . . . . . A/208 Hydraulic filter circuits . . . . . . . . . . . A/208 Hydraulic motor circuits . . . . . . . . . A/209 Hydraulic speed-control circuits . . . A/222 Intensifier circuits . . . . . . . . . . . . . . A/211 Locking circuits . . . . . . . . . . . . . . . . A/200 A/196
Meter-in & meter-out circuits . . . . . A/222 Motor braking circuit . . . . . . . . . . . . A/209 Parallel motor circuit . . . . . . . . . . . . A/210 Pneumatic speed-control circuits . . A/212 Pressure-control circuits . . . . . . . . . A/215 Pump-unloading circuits . . . . . . . . . A/217 Regenerative circuits . . . . . . . . . . . . A/219 Safety circuits . . . . . . . . . . . . . . . . . . A/220 Sequencing circuits . . . . . . . . . . . . . A/221 Series motor circuit . . . . . . . . . . . . . A/210 Speed-control circuits . . . . . . . . . . . A/222 Synchronizing circuits . . . . . . . . . . . A/225
Accumulator circuits 1 gal 10 gal
c
a
b
Accumulators store fluid energy and are used to reduce pump capacity requirements, speed operation, reduce pressure surges, and as standby power sources. Fluid pressure is developed by weights, spring tension, or compressed gas.
§ Traverse and clamp This arrangement of a large and small accumulator acts similar to a hi-lo circuit for rapid traverse and clamp. Fluid from the large accumulator combines with pump output to extend the cylinder rapidly. Fully extending the cylinder trips the limit switch to actuate solenoid (c). The small accumulator then maintains high clamping pressure on the cylinder for a timed period, during which the pump recharges the large accumulator. Any fluid lost by the small accumulator will also be replaced during this time.
LS
Press ram Starting motor
High-pressure relief valve Lowpressure relief valve Pressure switch Hand pump Diesel engine
M Main pump
¶ Standby power source In this diesel engine starting circuit, maximum power is required only for a short period, and time between operations is long. Power for starting the engine is stored in the accumulators. During operation of the engine, the main pump charges the accumulators to the pressure setting of the unloading valve. The pump is then unloaded for the remainder of engine running time. For starting, the manual valve is opened, connecting the combined output from the accumulators to drive the fluid motor. The hand pump serves as a means of recharging the accumulators in case of leakage over a long period of inactivity.
¶ Reduced pump capacity Frequently, pumps can be downsized if the circuit uses an accumulator. The accumulator volume adds to that of both pumps to speed downward travel of the press ram. When the ram meets sufficient resistance, the pressure switch is actuated, shifting the solenoid valve. This directs fluid from the large pump to recharge the accumulator, while the small pump continues to supply high-pressure fluid to the ram. When the manually operated, 4-way valve is shifted for the return stroke, pressure is relieved, the solenoid valve is de-energized, and both pumps and the accumulator deliver fluid for rapid return.
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BASIC CIRCUITS Vise jaws
M M
Unloading valve
¶ Surge reduction Operating the 4-way, closed-center valve in this circuit can cause the formation of shock pressures several times the value of the maximum pressure setting on the relief valve. Because the relief valve cannot act fast enough to drain off fluid, the high pressures can be dangerous to personnel and equipment. The accumulator in this circuit absorbs the surge pressures generated when the valve is placed in the neutral position.
¶ Clamping Holding pressure, leakage compensation, and power savings are obtained by using the accumulator in this vise circuit. While the vise jaws are in the clamp position, pressure is held by the accumulator, and pump output is unloaded at low pressure. The accumulator compensates for any leakage past the piston seals in the cylinder. When clamping pressure drops below the setting of the unloading valve, the valve closes, and the pump recharges the accumulator.
Unloading valve Pressure loaded rolls
M
¶ Increased speed Using a pilot-operated check valve allows adding fluid from the accumulator to pump output at the proper time within a cycle. Operating the manual valve directs fluid to retract the cylinder, exerting a pulling force. When pressure increases, the check valve opens, connecting the accumulator to the cylinder for fast action. Releasing the manual valve allows the pump to recharge the accumulator to the pressure setting of the unloading valve.
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M ¶ Safety device These mill rolls are loaded by hydraulic pressure. Using an accumulator allows running the pump unloaded most of the time, which saves power. The accumulator also protects the rolls from damage if a large piece of foreign matter enters the mill by absorbing fluid displaced when the roll rises. This fluid returns to the circuit when the foreign matter has passed through.
Air lubrication circuits
Valves, cylinders, and air motors can be lubricated by injecting oil into the air stream powering them. Properly locating the lubricator in the circuit is important to ensure proper lubrication.
§ Oil injection In the upper circuit, oil is injected into the head end of the cylinder each time the control valve cycles. In the bottom circuit, oil is injected into the cylinder through a small tube inside the air line.
Injector
Ä Lubricating short-stroke cylinders In this circuit, the lubricator bypasses the control valve and lubricates the cylinder while its rod retracts. When the rod extends, exhaust air from the head end lubricates the control valve. ¶ Downstream lubrication The upper drawing shows a conventional circuit: the lubricator mounts ahead of the control valve. In the lower drawing, air flow through the directional valve lubricates the cylinder. The valve is lubricated when the cylinder exhausts.
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Cylinder-locking circuits © Cylinders interlocked on press This open-center hydraulic system consists of a variable displacement, pressure compensated hydraulic pump and a 4-way, 3-position, solenoid operated control valve, which is interlocked hydraulically and electrically with a punch cylinder. If the punch is operated with the shear in mid-cycle, the shear hesitates until the punch operation is completed. A regenerative hydraulic circuit is provided by the check valve A. Oil—at pressures to 500 psi—flows directly from the head to the cap end of the cylinder without returning to tank. At 500 psi, relief valve B opens, allowing rod end oil to return to reservoir. Valves C and D operate in a similar manner.
M
A/200
Punch cylinder
500 psi
Shear cylinder
500 psi
M
§ Relief valve holds load A pilot-controlled relief valve can be used to control a single acting ram which is returned by a spring or by gravity. When the pressure is increased to the point where it balances the load and friction forces, the ram stands stationary. A slight increase in pressure setting will cause the ram to move upward. The check valve allows the ram to move upward. The check valve allows the ram to be held for a length of time while the pump is unloaded. The manual valve is then operated to lower the ram. This should be a needle valve for smooth operation.
BASIC CIRCUITS
§ Lock telescoping cylinder Loaded dump trucks had to be raised (requiring 100 gpm of pressurized fluid) and held up with no drifting. Then the load would be dumped and the empty trucks lowered (150 gpm to reservoir). This would have required a directional control valve large enough to bypass 100 gpm from the pump while simultaneously returning 150 gpm to the reservoir. It would have been an expensive, special valve. The practical solution used two vented relief valves. Because one relief valve (valve A) was needed anyway; it is used to unload the pump during idle time and during load holding. Relief valve B provides the outlet for fluid flowing from the cylinder to the tank during load lowering. As a byproduct, valve B provides shock protection to the cylinder.
Telescoping cylinder
1200 psi
A
1200 psi
B
M
© Differential cylinders Although the cylinder can usually be held against its end stops with fluid pressure, the check valve is needed for positive locking. Without the check valve, an external load could overcome the locking force and move the cylinder by flowing fluid through the relief valve. In addition, a drop in pump pressure could cause unloading.
M
M
§ Leakage prevention When a 3-position valve is used to lock a cylinder, pilot-operated check valves connected in the cylinder lines prevent leakage through the manual spool valve. In these poppet-type valves, the seal becomes tighter with increasing pressure. The check valves should have light enough springs to allow them to open and allow reverse flow at a pressure determined by the load and by cylinder resistance. Pilot-operated check valves ahead of each port, with pilot pressure supplied from the opposite port prevent drifting in double-acting cylinders.
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Deceleration circuits Oil
Water in
To water drain Air
Load
Air in
¶ Air-oil system stops load quickly, smoothly — This simple, reliable, and inexpensive system uses a tandem cylinder arrangement where pneumatic cylinder A (4-in. bore, 18-in. stroke) is coupled to hydraulic cylinder B, which has a 11/2-in. bore and a 19-in. stroke. The air cylinder provides power; the oil cylinder controls speed. As the air cylinder moves down, pulling cylinder B with it, oil is forced out of the rod end of cylinder B to the cap end through cam-operated, deceleration valve E. Flow is unrestricted until the cam gradually closes valve E, stopping the load smoothly. Flow from the rod end of the oil cylinder returns to the cap end through a water-cooled heat exchanger, and a pressurized reservoir supplies make-up fluid. Pressurizing the reservoir also prevents sucking air into the oil cylinder through the rod packing on the upstroke. The heat exchanger has a thermostatically controlled water valve.
A/202
BASIC CIRCUITS
M
§ Deceleration — Cylinders moving at high speed often must be decelerated to a controlled, smooth stop. Connecting a cam-operated, 2-way valve in the output line of a cylinder allows slowly closing off oil flow from the cylinder. In this circuit, with a double-rod cylinder, cam-operated valves are installed in both lines to the cylinder to provide deceleration in both directions. Cushions built into the end caps of the cylinder would perform the same function. However, when a cylinder must be stopped before reaching the end of its stroke, external deceleration valving is required. An electronic control method would replace the cam-operated valves with solenoid-operated valves. Limit or proximity switches would then sense when the cylinder stroked to a position requiring deceleration to commence. The machine’s controller, upon receiving the signal from the switch, would relay a signal to close the appropriate valve.
© Valve dampens spring force This pneumatic actuator is suitable for use in potentially explosive environments. Opening air switch B momentarily allows pilot air to pilot-operated valve A. System air then shifts the valve, and the cylinder extends. When extended fully, the cylinder trips the cam of a bleed valve, and the pilot line exhausts when the bleed valve opens. This air pressure loss closes pilot valve A, retracting the spring-returned cylinder and its load. Air exhausting through the pilot valve dampens the spring force and prevents the load from coming to an abrupt halt.
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Decompression circuits
Sudden release of pressure in a large volume of fluid can cause pressure surges in the system, as can the release from tensioned machine members. Controlled decompression gives a slow release of stored energy.
Platen
Counterbalance valve
Prefill valve Timer
M
¶ Pump decompression Small, auxiliary cylinders, called kicker cylinders, increase closing and opening speeds on large presses. The large cylinders are filled directly from the reservoir through a prefill valve. The 2-way pump delivers any volume from zero to maximum flow to the kicker cylinders. Kicker speed is reduced by the pump as the platen reaches the workpiece. Pressure or position shifts the prefill valve to direct fluid to all cylinders for full-force operation. Upon completion of work, a timer initiates metered, variable decompression of the large cylinder through the pump. Pressure to the kicker cylinders is maintained while decompressing the large cylinders. Upon completion of decompression, the prefill valve opens, initiating metered variable decompression of the kicker cylinders. The press then opens by gravity at a preset, pump-metered speed.
A/204
BASIC CIRCUITS
© Pressure-drop decompression To decompress the system and lower the ram, valve A’s solenoid is de-energized. Stored energy is released at an adjustable rate through flow control valve D. Backpressure created by valve D holds valve B shifted, even though solenoid valve A is de-energized. Pressure continues to decrease at a rate controlled by flow control valve D. Relief valve C closes when line pressure drops to 250 psi. Once valve C is closed, pilot flow to valve B stops. Its spring shifts valve B, connecting the ram directly to tank. The pressure setting of valve C determines the reversal pressure of valve B. Flow from the high-pressure pump passes through the decompression valves until valve B shifts. Now all output from the highpressure, fixed-displacement pump flows to tank through valve B and this pump is fully unloaded. Relief valve C is set higher than the low-pressure, variable-displacement pump to allow the pump to deadhead rather than discharge its full output through the system. This also allows the use of small decompression valving. Ä Pilot check valve decompression Pilot pressure first opens the pilot-operated check valve, allowing fluid to return to tank at controlled rate. This valve then opens fully, and the ram quickly retracts. Poppet areas and line pressure relationships in the pilot-operated check valve determine the amount of decompression. If the valve has a decompression ratio of 3 to 1 and an opening ratio of 15 to 1, the area of the pilot ram is 36 times that of the decompression poppet. This means that pilot pressure need be only 1/36 of main line pressure to enable the pilot ram to unseat the decompression poppet. After the system has been decompressed, the area ratio drops to 1.5 to 1, requiring a pilot pressure of only 2/3 to open the valve fully.
250 psi
3000 psi 200 psi
M
4-in. bore x 10-in. stroke
5000 psi 750 psi
7:1
High: 715 psi Low: 200 psi
M
M ¶ Control valve decompression In this circuit, a slow-shifting valve spool controls the rate at which system pressure decreases and fluid returns to tank. This is done by regulating and controlling the volume of pilot oil allowed to flow to the main valve. The slow shifting main spool meters fluid passing through it, and controls the rate of oil flow from the cylinder to tank. This method is particularly useful for low-pressure circuits where system compliance is low.
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Electrical control circuits
Limit switches sense cylinder positions, pressure switches sense system pressures, times control cylinder dwell, and relays actuate hydraulic and air solenoid valves.
© Two pressures The combination of presHigh pressure sure, position, and time ele1-CR 2-PS ments is illustrated in this cirCycle start cuit to control one source at 1-CR two pressures. With relay 1CR actuated by the cycle start Low pressure 1-CR switch, limit switch 1-LS 1-PS 1-LS 2-CR closes, starting the timer when the cylinder contacts the work. Pressure rises to open pressure switch 1-PS, 2-CR de-actuating relay 2-CR 1-TR which de-energizes a solenoid 1-CR valve to release pressure on 1-TR the cylinder. When the pressures drops, pressure switch 1-PS closes actuating relay 1CR to advance the cylinder. This cycle of increasing and decreasing pressures continues until the timer times out and its contacts in parallel with pressure switch 1-PS close. Pressure then rises to the setting of pressure switch 2-PS. The opening of 2-PS opens the interlock for 1-CR, and the cylinder returns to its initial position for the next cycle.
Ä Sequencing Cylinder A must advance and dwell while cylinder B rises to position 4. Both cylinders then retract together. Pushing the start switch energizes relay 1-CR and solenoid b, advancing cylinder A. Limit switch 1-LS is then actuated to energize relay 2-CR and solenoid d to extend cylinder B and de-energize solenoid b to stop cylinder A. When limit switch 2-LS is actuated, relay 3-CR is energized to open the circuit to relay 2-CR and deenergize solenoid d. Solenoid b is now energized, so pressure builds up on cylinder A until the pressure switch opens to de-energize relay 1-CR and energize solenoids c and a, retracting both cylinders to their original positions.
1-CR 1-LS 2-LS
Stop
Start
1-CR
2-CR
PS
Position 2
1-CR
Position 1 Position 4 b
c
Position 3
Sol. X
1-LS 2-CR 2-CR
Sol. Z
2-LS a
d
3-CR 1-CR
M A/206
Sol. Y
Sol. W
BASIC CIRCUITS
Stop Reverse
Start
Reverse
Forward Rev.
Forward Rev.
Reverse Forward
Reverse
For.
Reverse
For.
Forward
Forward
(a)
(b) Forward Start
Stop Stop
For.
Forward
Off Forward
Reverse For.
Reverse Forward Rev.
Reverse
Reverse
Rev.
Reverse
Forward
(c)
(d)
Cam
1-LS Cycle start
¶ Actuating two relays The circuits show four methods to actuate two relays controlling cylinder motion through solenoid valves. In (a) doublethrow pushbutton switches are connected in the actuating circuit of the opposite relay to prevent actuating both relays at once. In circuit (b), wiring to a pushbutton station is reduced, but a stop switch must be pushed before pushing the forward or reverse switch. Simultaneous actuation of both pushbuttons will not energize both relays. In circuit (c), the forward and reverse switches open contacts in the holding circuit of the opposite relay to de-energize that relay. Actuating both switches at the same time will energize both relays. Circuit (d) uses a 3-position selector switch; only one start switch is required. If a 2-position selector switch were used, it would have to be the break-beforemake type. Use of a single pushbutton switch or energizing either relay may be a safety hazard.
2-LS
1-LS 1-CR
1-CR
2-CR
1-CR
1-PS
Low-pressure pump control 2-CR High-pressure pump control
1-CR
2-PS 3-CR
3-CR
Stop
Left
1-LS
2-CR
Sol. A
1-CR 2-LS
Right
1-CR 1-CR
§ Two-pump control This circuit provides control of a hi-lo circuit. Limit switch 1LS is held closed in the machine starting position. The cycle start switch energizes relay 1-CR, which in turn energizes both the low and high pressure control systems. Pressure switch 1-PS opens at the low pressure setting, de-energizing relay 2-CR and relay 1-CR. The high pressure system has been interlocked through its own relay contact 3-CR. At the high pressure, 2-PS opens, de-energizing relay 3-CR. The pressure control system is now inactive.
Sol. B
§ Reciprocation Two limit switches provide automatic reciprocation when either the right or left pushbutton is depressed. The cam moves 2-CR 1-LS with the cylinder motion and actuates the limit switches at either end of its travel. The limit switches are double-pole units with one normally open and one normally closed contact. Pushing the left switch energizes solenoid A and relay 1-CR to move the cylinder and cam to the left. At the end of travel, limit switch 1-LS is actuated, de-energizing solenoid A and energizing solenoid B and relay 2-CR. Each relay locks in when actuated and remains energized to prevent actuating the alternate solenoid. Pressing the stop switch de-energizes the system at any point, except the extreme positions. 2-CR
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Hydraulic filter circuits
High-performance circuits require clean fluids. Filters may be installed in pressure lines, return-to-tank lines, or bleedoff lines.
¶ Return line bypass filter A backpressure check valve forces fluid through an orifice to the filter. The only limitation here is that the backpressure must not interfere with circuit operation.
¶ Suction line strainer or filter To prevent pump cavitation, components in this system must have low pressure drop and high capacity. Filter elements should be submerged in reservoirs so no part of the filter surface is exposed to air.
¶ Pressure line bleed-off filter An orifice between the pump discharge and filter assembly maintains a constant flow through the filter. This circuit generally is used on high-flow circuits because flow through the filter is lost as effective pump output.
§ Discharge line filter An advantage of this circuit is that it filters oil immediately as it returns from the work station. Return line filters can tolerate a higher pressure drop than those in a suction line.
¶ Pressure line filter This circuit has the advantage of full-flow filtration but the disadvantage of high pressure drop. Often it can be used instead of a low-pressure return line filter to prevent building up backpressure.
¶ Independent circuit Sometimes called a kidney loop, this circuit filters full flow from a separate filter pump. These circuits often incorporate a heat exchanger and multiple filters.
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Hydraulic motor circuits
Torque is produced hydraulically with a fluid motor. Pressure regulation allows torque control. Flow regulation permits speed control. Use of a variable displacement motor produces a constant horsepower drive.
Brake valve Brake
M
M
¶ Constant torque & constant power Driving a fixed-displacement fluid motor at constant pressure produces a constant torque drive, left-hand drawing. Used with a variable-volume pump to vary flow, the horsepower output of the motor varies with speed. If the load becomes excessive, pressure rises to actuate the pressure switch, de-energizing the solenoid valve — pump unloads, fluid motor stops.
In the right-hand drawing, a fixed-displacement pump supplying a variable-displacement motor at constant pressure produces a constant-horsepower drive. The motor produces lowest speed and highest torque when displacement is maximum. Highest speed and lowest torque are produced when the displacement is minimum.
Ä Braking When pump delivery is cut off from a motor, it continues rotating because of its inertia and that of the connected load. The motor then acts as a pump, and a source of fluid must be available to prevent it from cavitating. In (a) the manual valve allows coasting to a stop and braked stop, as well as the normal driving condition of the motor. With the valve spool up, the pump output drives the fluid motor. With the spool centered, pump output and both sides of the fluid motor are connected to tank, so that the motor coasts to a stop. With the valve spool
down, the pump is unloaded and the motor, acting as a pump, forces fluid through the relief valve, which brakes it to a stop. Circuit (b) shows a brake valve that is a modified sequence valve. It supplies braking force as well as control of a negative work load. Under normal conditions, system pressure holds the brake valve open for free discharge from the motor. A negative load reduces pressure at the motor inlet, and the brake valve closes to throttle motor discharge and create a backpressure. To stop the motor, the 4-way valve is shifted to neutral.
X
M (a)
M
Brake valve
(b)
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BASIC CIRCUITS
Vent Line
¶ Series connection Connecting two fluid motors in series minimizes pump size and eliminates the need for a flow divider. Line sizes are also smaller than in a comparable parallel circuit, and piping is usually simpler, with only one pressure line and one return line required. Maximum torque at each motor is adjustable with the relief valves. The speed of motor A is controlled by the bleed-off flow control valve. Direction of motor B is controlled by the 4-way valve, which has an integral 2-way valve, which vents the relief valve when the motor is stopped. Total system pressure is then available at motor A.
¶ Parallel connection Pump pressure can be lower in a parallel circuit because in a series circuit the pressure at the pump must be the sum of the pressure drops across the motors. However, where motor pressures vary widely, there is a loss of efficiency in supplying the motors requiring lower pressures. This circuit is most efficient where the load on each motor is the same. Raising the pressure on one motor renders the others less efficient and may disrupt the speed relationship. In parallel circuits, the only way to increase torque of the highest-pressure motor is to increase system pressure.
Ä Replenishing When a fluid motor and pump are connected in a closed circuit, make-up fluid to compensate for the leakage must be supplied through replenishing valves. These valves also supply fluid to the motor during braking. In the left-hand circuit, a nonreversing variable volume pump is used and control of fluid motor direction is by the 4-way valve. The fixed displacement
pump provides supercharge pressure. The network of check and relief valves provides for replenishing and braking in either direction. Braking pressure in each direction can be set independently on the two brake valves. In the right-hand circuit, a reversing pump is used. Although the replenishing network is simpler, brake pressure must be the same in both directions.
Brake valve Brake valve Brake valve
Supercharge relief
M A/210
M
Intensifier circuits
Intensifiers boost system pressure in one section of a circuit. They are suited for applications where pressure must be held more than a few seconds and provide a method of obtaining high force with air systems. The use of intensifiers has the advantage of keeping high pressure fluid in short lines near the work cylinder. This reduces maintenance of equipment and eliminates high-pressure pumps and valves.
Tandem cylinder
§ Force multiplication The two sections of this tandem cylinder are connected in parallel to provide fast approach to the work and high clamping force without use of high pressure. Shifting the 4-way valve directs fluid to the small cylinder for rapid closing. When the work is contacted, pressure builds up to open the sequence valve and apply pressure to the large cylinder. Clamping force is equal to the pressure of the relief valve setting times the combined areas of both cylinders. A replenishing line to the cap end of the large cylinder prevents pulling air in behind this cylinder during rapid extension. This circuit allows using a smaller capacity pump than would be needed in a circuit using a large cylinder operating at the same speed. Intensifier 3:1
M ©Reciprocating booster The booster in this applications contains an integral 4-way valve to reciprocate the piston and supply larger volumes of high-pressure fluid than a single-acting booster. When the required pressure is reached, the booster stalls but will resume reciprocation automatically when additional high-pressure fluid is required. In this circuit, the ram is lowered with fluid directly from the pump. When the load is contacted, the pressure rises to open the sequence valve. The booster is supplied through a pressure reducing valve and triples pressure for high-force squeeze.
Press ram
M Oil tank
Air Oil
Oil tank
Air-operated hydraulic booster
Air Oil
Work cylinders
§ Hydraulic intensification Normal shop air pressure can be multiplied by using an air-oil booster. When the manual valve is shifted, air pressure forces oil from the upper tank to extend the work cylinders. When the work cylinders contact the load, pressure rises, opening the sequence valve and applying pressure to the booster. The booster piston extends to cover the oil port and then develops high pressure for holding the cylinders. Because forces are equal on the air and oil ends of the booster, pressures are inversely proportional to the rod and piston areas.
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Pneumatic speed-control circuits © External cushions Control of external cushioning may be done with cam-operated valves or solenoid valves actuated by limit switches. In this circuit, three solenoid valves provide cushioning at both ends of the cylinder’s strokes. The cycle ends with the cylinder extended. Pressing the start push-button energizes 1-CR and solenoids that operate valves A and B. Air flows to the head end of the cylinder, while the cap end exhausts unrestricted through valve B. The cylinder retracts fast. When the cam on the piston rod actuates limit switch 1-LS, relay 2-CR is energized. This de-energizes the solenoid to close valve B. Air from the end cap of the cylinder flows through valve D, gradually decelerating the piston. When the cam actuates 3-LS, no circuit changes occur. Near the end of the stroke, the cam actuates 2-LS, deenergizing relays 1-CR and 2-CR and energizing relay 2-CR. Valve A’s solenoid de-energizes, and the solenoid is energized to shift valve C. The cylinder reverses its movements and extends fast. When the cam actuates limit switch 3-LS, relay 2-CR de-energizes. Valve C closes, and air remaining in the head end of the cylinder is metered through valve E. The cylinder gradually decelerates through the rest of its stroke. Limit switch 1-LS actuates only when the piston retracts.
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Cushioning, absorbing shock, and controlling speed of air cylinders are complicated by the compressibility of air, but this can be overcome to provide timing, deceleration, dwell, and other functions.
Fast Stop
Fast
Fast
ct
Retra
2-LS
Exten
d
3-LS
Stop
1-LS
Start 1-CR 2-LS 1-CR 2-CR 1-LS 2-CR 3-CR 3-LS
2-LS 3-CR Sol. 1
1-CR 2-CR 3-CR
Sol. 2 Sol. 3
BASIC CIRCUITS
1-LS
Advance Sol. 1
Speed selector
1-LS
Sol. 2
Sol. 2
Sol. 1
¶ Meter-in control The cylinder extends fast by bypassing the flow-control valve. Limit switch 1-LS is closed when the cylinder retracts. This limit switch energizes solenoid 2. When the advance pushbutton is depressed, solenoid 1 is energized. This valve shifts, which routes air to the cylinder and bypasses the flow-control valve. The cylinder advances rapidly until the cam rides off the limit switch, de-energizing solenoid 2. This valve shifts, directing air to the cylinder through the flow-control valve, which regulates speed. The cylinder retracts rapidly when the pushbutton is released. Air exhausts through the check valve until the cam closes limit switch 1-LS, shifting valve 2. Exhaust air then bypasses the flow-control valve.
§ Cylinder starts fast
Air supply
With valve A in the position shown, the cylinder retracts at a speed controlled by flow-control valve B. With the cylinder fully retracted, valve C opens. When valve A shifts to advance the cylinder, air is applied to the pilot of valve D. The valve opens, exhausting air from the head end of the cylinder to atmosphere. The cylinder advances rapidly until the cam rides off valve C, closing the atmosphere exhaust. The remainder of the cylinder’s stroke is regulated by the setting of valve E.
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BASIC CIRCUITS
§ Shear cushioning circuit Valve operating
Shear
Shear return
Clamp
At the beginning of the cycle, valve A, in the position shown, supplies air through the shuttle valves into the head ends of the shear-return cylinders. The valve-operating cylinder is retracted, holding valve D in the position shown and venting the single-acting shear cylinders. The clamp cylinders are retracted. To begin a shear cycle, valve B is shifted, which extends the clamp cylinders. Valve A is then shifted, extending the valve-operating cylinder and venting the shear-return cylinders through the shuttle valves. The valve-operating cylinder extends, and cam-operated valve D spring-shifts. Supply air is then directed to the shear cylinders, starting the shear stroke. After the cut, but before the shear cylinders reach the end of their strokes, cam valve C is shifted, retracting the valve-operating cylinder. Supply air also flows to the shear-return cylinders cushions and gradually decelerates the shear stroke after the cut. Valve A is released, and supply air through the shuttle valves retracts the shear cylinders. Valve B is shifted to release the clamp cylinders.
§ Cylinder dwell Time delay in this circuit is regulated by the flow-control valves and begins when the manual valve shifts. Momentarily actuating valve A shifts valve B to extend the cylinder. Air continues to supply the pilot of valve C through valve B and the flow-control valves. When valve C shifts, the pilot line is exhausted, allowing valve B to shift and retract the cylinder.
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Pressure-control circuits
Low-pressure relief valve
Controlling pressure allows regulation of output force of a cylinder or torque of a motor or rotary actuator. Most hydraulic systems operate at pressures ranging from 150 to 5000 psi; compressed air systems usually are maintained at 80 to 125 psi.
§ Two relief valves Using two relief valves in this circuit gives two working pressures. On the up stroke of the cylinder, the low-pressure relief valve limits system pressure. During the down stroke, the high-pressure valve limits maximum press tonnage for doing work. Using the low-pressure relief on the up stroke saves power by supporting the cylinder with low-pressure fluid.
M High-pressure relief valve
Pilot relief valve
¶ Low-pressure retraction Energizing the solenoid extends the cylinder at a maximum pressure corresponding to the main system relief valve setting. De-energizing the solenoid valve retracts the cylinder and holds it retracted at the reduced pressure setting of the pilot relief valve. The check valve prevents the pilot relief valve from operating during cylinder extension.
¶ Pilot pressure When open- or tandem-center valves are used in circuits requiring pilot pressure to shift the valves, there must be a means of maintaining pressure when the valves are in neutral. One method is to install a backpressure check valve in the tank line. The check valve maintains a backpressure of, say, 50 psi.
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BASIC CIRCUITS
Weld
Pressurereducing valve set at 400 psi
Clamp
§ Reduced pressure In a system with only one pump, reduced pressure for one branch of the circuit can be obtained with a pressure-reducing valve. This circuit is typical for a welder, which requires high clamping force to be set by the relief valve and reduced force on the welding gun to be set by the pressurereducing valve. Placing the check valve in parallel with the pressure-reducing valve allows free return flow when the weld cylinder retracts.
M Relief valve set at 800 psi
500 psi
1000 psi
Injection 400 psi
2000 psi Relief valve set at 1500 psi
Air supply
To work cylinders
M Mold close
¶Remote control Regulating pump pressure from a remote station can be accomplished by using small, pilot relief valves connected to the system’s main pilot-operated relief valve. With the 3-way solenoid valve de-energized, system pressure is limited to 1500 psi in this circuit. Energizing the 3-way solenoid valve permits venting the relief valve to either 1000 psi or 500 psi, depending on the position of the 4-way valve, which is determined by the pilot signal it receives.
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¶ Two pressures Pilot-operated relief valves provide two pressures for the mold-close cylinder and the injection cylinder of this plastic molding machine circuit. With the manual valve in neutral, the air-operated valves are actuated to extend the mold-close cylinder at a maximum pressure of 2000 psi. Operating the manual valve to extend the injection cylinder vents the 400-psi relief to tank and vents the 2000-psi relief at 400 psi. High-pressure fluid is held by the check valve.
Pump unloading circuits © Hi-lo circuit Many systems require a high volume at low pressure for rapid traverse of a vise or tool, and then low volume, high pressure for clamping or feeding. This can be accomplished by a hi-lo circuit using two pumps. During rapid traverse, both pumps supply the system. When pressure rises during clamping or feed, the large-volume main pump unloads, and the small pump maintains pressure. Output flow of the small pump is low enough to prevent appreciable heating of the oil. Instead of pilot operation, the unloading valve can be solenoid controlled and actuated by a pressure switch.
Load
Load
M
(a)
When no flow is required in a hydraulic system, pump output can be returned to tank at low pressure instead of dumping high-pressure fluid over a relief valve.
High-pressure relief valve
Unloading valve
Pressure Maintaining pump
Main pump
M
§ Pilot-operated relief Using a pilot-operated relief valve gives fast action at the preset pressure with very low unloading pressure. The valve’s main spool is held closed by a spring and balanced pressures. Opening the pilot connection unbalances the spool and allows it to open quickly. In these circuits, the pilot is opened to tank through a 2-way valve operated either by a solenoid or by pilot pressure. The solenoid is actuated by a pressure switch in circuit (a). The pilotoperated valve in circuit (b) is controlled through a sequence valve set for maximum system pressure. The accumulator and check valve maintain system pressure during unloading.
(b)
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BASIC CIRCUITS
Dump Valve
To system
M Relief valve
¶ Manual control A manually operated 2-way valve can be connected to unload the pump. This places responsibility on the operator to unload the system when no pump flow is required. In this circuit, the entire system is connected to tank through the manual valve. ¶ Pressure-compensated pump A pressure-compensated, variablevolume pump is controlled by system pressure. As pressure increases, displacement of the pump decreases so that pump output at the preset pressure is only sufficient to make up for leakage. Used with a closed-center valve, the pump is stroked to minimum (zero) displacement when the valve is centered.
M
¶ Pressure unloading two-way pump A 2-way, variable-displacement pump with automatic pressure unloading control and servo gives variable speed in either direction. When the cylinder bottoms, system pressure at one pump port reduces pump delivery and power input to maintain a preset force. The manual lever is centered for zero pump delivery or is moved through 100° to provide any volume up to maximum at either of the pump ports.
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M
¶ Open-center system In the open-center system, pump output is connected through a directional control valve to tank. When this circuit is used with low flows and the directional valve spool has tapered lands, the circuit provides a simple, efficient system. When several cylinders are used, the valve can be connected in series — that is, the tank port of one valve is connected to the pressure port of the next.
Many applications require highforce actuation for only a portion of a cylinder’s stroke. Regenerative circuits provide reducedforce actuation with the advantage of faster speed, which reduces over-all cycle times.
Regenerative circuits Work
1
X
Work
Ä Circuit 4 When electrical control is desired, a limit or proximity switch can be used to activate and de-activate a regenerative circuit. In the circuit shown, energizing solenoids a and b extends the cylinder in a differential circuit. The limit switch de-energizes solenoid b, directing cylinder discharge fluid to tank.
2 b
Work
a
c
3 4 ¶ Circuit 1 Rapid rod extension can be achieved by returning the flow of oil from a cylinder’s head end back into its cap end. With no load, pressure in both head and cap ends is equal, so when the load is encountered, available working force depends on the differential area. Circuit 2 Combining the rapid extension of Circuit 1 with full force in response to an applied load takes full advantage of a regenerative circuit. The circuit shown produces a rapid approach stroke of the piston rod. When the rod encounters resistance (workpiece load), pressure rises on the cap end to open the sequence valve and allow oil
from the head end to flow to tank through the 4-way valve. Once this occurs, full effective force on the workpiece becomes available. Circuit 3 This circuit provides a costly means to accomplish the same end as Circuit 2. Instead of a sequence and built-in check, an orifice and check are used. There is some backpressure remaining in the cylinder’s head end because of the orifice resistance during final squeeze. But with the cylinder extended, no fluid flows across the orifice, so total available force acts on the cap end. Whether or not this circuit is appropriate for a given application depends on working force requirements. 1998/1999 Fluid Power Handbook & Directory
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Safety circuits
Circuits in this section are designed to help prevent accidental operation, protect against overloads, and ensure safe operation.
Ä Two-hand operation For reasons of safety, it is often important on presses to have both of an operator’s hands engaged in pushing controls to keep them out of the way of the descending press ram. In the circuit, both manual valves must be activated before the piston rod will extend. Releasing either valve while the rod is extending stops the piston; both valves must be released before the piston rod retracts. Tying down one of the valves for “convenience” prevents retracting the rod for a new cycle. Air supply
¶ Overload protection Connecting a sequence valve and auxiliary 2-way, pilotoperated valve in the circuit above protects against pressures caused by overloads. The manual valve is operated to activate the air-piloted valves connected in the cylinder lines. If the extending piston rod encounters an obstruction — causing pressure in the head-end line in increase — the sequence valve will open to activate the auxiliary 2-way valve. This event exhausts the pilot line, retracting the rod. Under normal operation, the cam-operated valve is activated at the end of the cylinder stroke to exhaust the pilot line.
Start compressor
Compressor motor
Stop
Air supply
M M
Start
Stop PS
1CR LS
1CR
1CR Solenoid
LS
¶ Electrical safety circuit Accidental operation of a machine is prevented with a pressure switch in the electrical circuit above. The pressure switch closes when air achieves the required pressure, which permits energizing relay 1-CR with the Start switch after the compressor motor relay has been actuated. The solenoid is actuated to extend the piston rod. At the end of the stroke, the
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limit switch is deactivated, which de-energizes the solenoid valve and causes the rod to retract. The pressure switch prevents the solenoid valve from being energized when the air supply is off. Energizing the solenoid valve under this condition would allow the machine to operate without having been actuated once air pressure was restored.
Sequencing circuits 1
2
4 1 Clamp
3 2
M
Drill
Several basic methods are used to sequence operation of fluid power circuits: electrical, mechanical, and pressure-operated control systems. In some automatic sequencing systems these methods are combined. Many of the mechanical means of actuating valves have been replaced by electronic control.
§ Backpressure check valves Cylinders may be sequenced by restricting flow to one cylinder. One method of restricting flow is with backpressure check valves. They prevent flow until a set pressure is reached. In this circuit, cylinder 1 extends and retracts ahead of cylinder 2.
Work table cylinder
Cam-operated pilot valve
Clamp cylinder
Sequence valve
¶ Clamp and reciprocation A common requirement of machine tools is clamping of a work piece followed by automatic reciprocation of the work table. Shifting the manual valve in this circuit directs fluid to extend the clamp cylinder. When this bottoms, pressure builds to open the sequence valve. Directing fluid through the pilot-operated valve to the reciprocating cylinder. When this cylinder reaches the forward or reverse end of its stroke, it actuates the cam-operated valve to reverse the pilot-operated valve. Reciprocation continues automatically until the manual valve is reversed to release the clamp and remove pressure from the reciprocating cylinder, stopping reciprocation.
§ Sequence valves Several cylinders can be connected to move in sequence on forward and return strokes. In this circuit, a clamp must close before a drill descends. On the return stroke, the drill must pull out of the work before the clamp opens. The sequence valves are arranged to cause pressure buildup when one cylinder completes its stroke, the valve opens to allow flow to the other cylinder. 1998/1999 Fluid Power Handbook & Directory
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BASIC CIRCUITS
3
4 1
a
b
2
c Momentarily energized solenoid valves
d
§ Electrical control Limit switches momentarily actuated by the cylinders control the solenoid valves to sequence this circuit. Solenoid a is energized by a pushbutton to initiate movement 1. At the completion of movement 1, limit switch E is actuated to energize solenoid c, initiating movement 2. At the end of movement 2, limit switch F is actuated to energize solenoid b, initiating movement 3. At the end of this movement, limit switch G is actuated to energize solenoid d, initiating movement 4. The sequence valves prevent a pressure drop in either cylinder while the other operates.
Speed-control circuits
¶ Meter-in circuit Speed control during a work stroke can be accomplished by regulating flow to the cylinder. The check valve allows free reverse flow when the cylinder retracts. It normally gives finer speed control than a meter-out circuit.
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¶ Meter-out circuit Regulating flow from the cylinder is another way to control speed. This circuit maintains a constant backpressure during rod extension and prevents lunging if the load drops quickly or reverses.
One of the great advantages of fluid power is the ability to control speed of a cylinder, motor, or rotary actuator. Speed is primarily a function of oil flow and size of the actuator.
¶ Bleed-off circuit Flow to the cylinder is regulated by metering part of the pump flow to tank. This circuit is more efficient than meter-in and meter-out circuits because pump output pressure is only high enough to overcome work resistance. However, it does not compensate for pump slip.
BASIC CIRCUITS
3
4 1
a
b
2
c Momentarily energized solenoid valves
d
§ Electrical control Limit switches momentarily actuated by the cylinders control the solenoid valves to sequence this circuit. Solenoid a is energized by a pushbutton to initiate movement 1. At the completion of movement 1, limit switch E is actuated to energize solenoid c, initiating movement 2. At the end of movement 2, limit switch F is actuated to energize solenoid b, initiating movement 3. At the end of this movement, limit switch G is actuated to energize solenoid d, initiating movement 4. The sequence valves prevent a pressure drop in either cylinder while the other operates.
Speed-control circuits
¶ Meter-in circuit Speed control during a work stroke can be accomplished by regulating flow to the cylinder. The check valve allows free reverse flow when the cylinder retracts. It normally gives finer speed control than a meter-out circuit.
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¶ Meter-out circuit Regulating flow from the cylinder is another way to control speed. This circuit maintains a constant backpressure during rod extension and prevents lunging if the load drops quickly or reverses.
One of the great advantages of fluid power is the ability to control speed of a cylinder, motor, or rotary actuator. Speed is primarily a function of oil flow and size of the actuator.
¶ Bleed-off circuit Flow to the cylinder is regulated by metering part of the pump flow to tank. This circuit is more efficient than meter-in and meter-out circuits because pump output pressure is only high enough to overcome work resistance. However, it does not compensate for pump slip.
BASIC CIRCUITS
§ Auxiliary cylinders and prefill On a press requiring a large cylinder to generate the necessary force, platen lowering speed can be increased by using smaller auxiliary cylinders, called kicker cylinders. The manual valve routes pump output to two kicker cylinders while the main cylinder prefills by drawing fluid in from the reservoir. When the kicker cylinders bottom, the sequence valve opens and pressure is applied to the main cylinder. Shifting the manual valve raises the platen and directs pilot pressure to open the prefill valve; fluid from the main cylinder flows back to the reservoir.
Tank Prefill valve
Counterbalance valve Press platen
M
© Variable feed Many machines require intermittent fast and slow feed during the cycle. This can be accomplished by having a cam-operated 2-way valve in parallel with a meter-out flow control valve. Rapid forward movement takes place any time the 2-way valve is open. Closing off the valve thereby slowing down cylinder speed. Properly positioning the cams obtains the required speeds in sequence. The check valve in parallel with the flow control permits free return flow, allowing the cylinder rod to return rapidly.
M
M
§ Variable-volume pump Pump flow can be controlled by various means such as manual, electric motor, hydraulic, or mechanical. How closely flow output actually matches command depends, in part, on slip, which increases with load. With a pressure-compensated, variable-volume pump, output flow decreases with the increasing pressure. This type of pump can be used for traverse and clamp operations. An external relief valve is usually unnecessary when a pressure-compensated pump is used. For details on the different types of pumps, their operation, and how they vary flow, refer to the pumps section of this Handbook.
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BASIC CIRCUITS
From pump
© Drill-ream circuit Four different cylinder speeds — for example, rapid traverse, coarse drill, fine drill, and ream — are possible with a drill-ream circuit. Solenoid valves select flow control valves to meter fluid when the mechanically actuated 2-way valve closes.
Drill Ream Coarse feed
Fine feed
Drill-ream
Low speed Unload all pumps
Medium speed
High speed
Air supply
Oil output
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§ Pneumatic control Pneumatically loading and unloading three hydraulic pumps can give speed ranges of low, medium, and high. The 2section, low-speed valve loads pump A and unloads pumps B and C. (Valve E loads pump A, and valve F loads pumps B and C.) The hydraulic unloading valves are pilot-operated through valves J, K, and L. Valve E vents air from valve F, valve F from K and L. The check valves prevent interaction. Medium-speed valves G and H operate similarly. For high speed, valve I loads all pumps by venting the right-hand pilots of valves J, K, and L. To stop oil flow completely, valve D vents the left-hand pilots of valves J, K, and L.
Synchronizing circuits
Theoretically, two cylinders will move together when flow to each is the same. Variations in leakage, friction, and cylinder size cause them to get out of step. These circuits are designed to compensate for these variations.
§ Two pumps Under certain conditions two variable volume pumps can be used to move cylinders in unison. Leakage and slip are compensated for by adjusting strokes of the pumps. This circuit readily permits changing cylinder speed as well as synchronizing movement.
M
1
2
© Replenishing One of the considerations in synchronizing cylinders is leakage replacement. Under normal pressure, leakage can be practically zero over one stroke. Accumulated error is the main concern. A replenishing circuit, which replaces leakage after each cylinder stroke, eliminates this trend. In the circuit, the cylinders are connected in series and controlled by the 4-way manual valve. The cylinders actuate limit switches, which control valves A and B. On the return stroke, if cylinder 1 bottoms first, valve A is actuated to open valve C, permitting excess fluid from cylinder 2 to flow to tank. If cylinder 2’s piston returns first, valve B is actuated to direct fluid to retract cylinder 1.
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BASIC CIRCUITS
©Synchronizing rotary actuators Valve A is a conventional 4-way, double-solenoid, 2-position valve. Valve B is modified porting as indicated. In the normal or reset position, valves A and B have been shifted by momentarily energizing the left-hand solenoid of each valve. With the valves in this position, actuator C is returned by pressure oil through valve B. Drain from the opposite end of this actuator returns through valve A. Actuator D is returned by oil flow through valve A. Drain from its opposite end returns through valve B. For a synchronized index stroke, the right-hand solenoid on each valve is momentarily energized. Flow from valve A connects to the driving end of actuator C, which discharges oil to the driving end of actuator D. Discharge from actuator D connects to tank through valve A. To ensure that both actuators fully return before an index stroke can start, two limit switches are installed at the actuators. A third limit switch is installed at the 90˚ point on actuator D, to ensure a full stroke by both units.
From pump
Control valve
Control valve
M
M ¶Fluid motor flow divider An effective flow divider can be made up of two fluid motors of the same size coupled together. Both motors must rotate at the same speed and, therefore, deliver equal volumes of fluid. Variations in load or friction do not greatly affect synchronization, but motor slip is a factor.
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© Metering Independent adjustment of the meter-out flow control valves synchronizes cylinder motion. Leakage and friction will cause inaccuracies. Contaminated fluid can cause flow control valves to malfunction. Valves therefore should be pressure compensated if loads are unequal.
§ Rack and pinion Mechanically tying two cylinders together by installing a rack on each piston rod and fastening the pinions to a single shaft works well when the linkage is rigid and the mesh is proper. A chain and sprocket arrangement can be used if synchronized motion is required in only one direction.
M
Fluid power graphics T
he symbols contained on this and the following pages are presented with permission of the International Organization for Standardization (ISO). ISO is a worldwide federation of national standards bodies. The work of preparing International Standards normally is performed through ISO technical committees. International organizations also take part in the work in liaison with ISO. ISO collaborates closely with the International Electrotechnical Commission (IEC) on all matters of electrotechnical standardization. An updated standard, ISO 1219-1, is being prepared by Technical Committee ISO/TC 131, Fluid power systems, SubCommittee SC 1, Terminology, classification, and symbols. When published,
ISO 1219-1will contain a substantial amount of revisions and will cancel and replace ISO 1219: 1976. ISO 1219 establishes principles for the use of symbols and specifies basic symbols and rules for devising functional symbols. It also includes examples of functional symbols. Use of ISO 1219 symbols does not preclude the use of other symbols commonly used for pipework in other technical fields. The symbols are neither drawn to scale nor oriented in any particular direction. The following standards contain provisions which, through reference in this text, constitute provisions of this part of ISO 1219: ● ISO 3511-1: 1977, Process measurement control functions and instrumenta-
DESCRIPTION
5.1.1
Line
5.1.1.1
— continuous
5.1.1.2
— long dashes
5.1.1.3
— short dashes
5.1.1.4
— double
5.1.1.5
— long chain thin (optional use)
5.1.2
Circle, semi-circle
APPLICATION
2)
E
BASIC SYMBOLS
tion — Symbolic representation — Part 1: basic requirements ● ISO 128: 1982, Technical drawings — General principles of presentation, and ● ISO 5598: 1985, Fluid power systems and components — Vocabulary. For definitions of terms used, consult ISO 5598. The symbols for fluid power components shall be constructed from the basic symbols and functional elements contained within this part of ISO 1219. Rules designed to enable users to devise complete or composite functional symbols are given for each clause. These rules permit two or more users working independently on a common specification to produce the same final symbol. A number of complete functional symbols are given as examples.
SYMBOL
LL
L > 10E
L
}
flow lines
L < 5E D
5.1
In fluid power systems, power is transmitted and controlled through a fluid under pressure within a circuit. Graphic symbols serve as an aid to functional identification in diagrams of fluid power systems. They also can be used on hardware for the same purpose.
mechanical connections (shafts, levers, pistonrods) D < 5E
Enclosure for several components assembled in one unit
5.1.2.1
As a rule, energy conversion units (pump, compressor, motor)
5.1.2.2
Measuring instruments
5.1.2.3
Non-return link, roller, etc
5.1.2.4
Mechanical link, roller, etc.
5.1.2.5
Semi-rotary actuator
2) L= Length of dash, E = Thickness of line, D = Space between lines
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
APPLICATION
5.1.3
Square, rectangle
As a rule, control valves (valve) except for nonreturn valves
5.1.4
Diamond
Conditioning apparatus (filter, separator, lubricator, heat exchanger)
5.1.5
Miscellaneous symbols
3) Flow line connection
5.1.5.1 d
5.1.5.2
d ≈ 5E
Spring
5.1.5.3
Restriction:
5.1.5.3.1
— affected by viscosity
5.1.5.3.2
— unaffected by viscosity
5.2
FUNCTIONAL SYMBOLS
5.2.1
Triangle:
The direction of flow and the nature of the fluid
5.2.1.1
— solid
Hydraulic flow
5.2.1.2
— in outline only
Pneumatic flow or exhaust to atmosphere
5.2.2
Arrow
Indication of:
5.2.2.1
— direction
5.2.2.2
— direction of rotation
5.2.2.3
— path and direction of flow through valves. For regulating apparatus as in 7.4 both representations, with or without a tail to the end of the arrow, are used without distinction As a general rule the line perpendicular to the head of the arrow indicates that when the arrow moves, the interior path always remains connected to the corresponding exterior path
5.2.3
228
Sloping arrow
Indication of the possibility of a regulation or a progressive variability
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 6.1
PUMPS AND COMPRESSORS
6.1.1
Fixed capacity hydraulic pump:
6.1.1.1
— with one direction of flow
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL To convert mechanical energy into hydraulic or pneumatic energy. OUTLET
,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, INLET
6.1.1.2
— with two directions of flow
6.1.2
Variable displacement hydraulic pump:
6.1.2.1
— with one direction of flow
,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,, BALL CHECK ,,,,,,,,,,,,,,,,,,,,,,,,
The symbol is a combination of 6.1.1.1 and 5.2.3 (sloping arrow)
U
6.1.2.2
— with two directions of flow
The symbol is a combination of 6.1.1.2 and 5.2.3 (sloping arrow)
,,,,,, ,,,,, ,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,, ,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,, ,,,,,,,,,,, ,,,,,, ,,,,,,,,,,, ,,,,,,,,,,, ,,,,,,,,,,, ,,,,,,,,,,,
,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
6.1.3
Fixed capacity compressor (always one direction of flow)
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 6.2
MOTORS
6.2.1
Fixed capacity hydraulic motor:
6.2.1.1
— with one direction of flow
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL To convert hydraulic or pneumatic energy into rotary mechanical energy
SPRING
6.2.1.2
— with two directions of flow
6.2.2
Variable displacement hydraulic motor:
6.2.2.1
— with one direction of flow
The symbol is a combination of 6.2.1.1 and 5.2.3 (sloping arrow)
6.2.2.2
— with two directions of flow
The symbol is a combination of 6.2.1.2 and 5.2.3 (sloping arrow)
6.2.3
Fixed displacement pneumatic motor:
6.2.3.1
— with one direction of flow
6.2.3.2
— with two directions of flow
6.2.4
Variable displacement hydraulic motor:
6.2.4.1
— with one direction of flow
The symbol is a combination of 6.2.3.1 and 5.2.3 (sloping arrow)
6.2.4.2
— with two directions of flow
The symbol is a combination of 6.2.3.2 and 5.2.3 (sloping arrow)
6.2.5
Oscillating motor:
6.2.5.1
— hydraulic
6.2.5.2
— pneumatic
230
,,,,, ,,,,,,,,, ,,, ,,,,,,,,, ,,,,, ,,, ,,,,,,,,, ,,, ,,,,,,,,, ,,,, ,,, ,,, ,,,, ,,,,,,,,, ,,,, ,,,, ,,,,,,,,, ,,, , ,,,,,, ,, ,,,,,,,,, ,,,,,,,,, ,,,,,, ,,,,,,,,, ,
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
SYMBOL
Unit with two functions, either as pump or as rotary motor
6.3
PUMP/MOTOR UNITS
6.3.1
Fixed displacement pump/motor unit:
6.3.1.1
— with reversal of the direction of flow
Functioning as pump or motor according to direction of flow
6.3.1.2
— with one single direction of flow
Functioning as pump or motor without change of direction of flow
6.3.1.3
— with two directions of flow
Functioning as pump or motor with either direction of flow
6.3.2
Variable displacement pump/motor unit:
6.3.2.1
— with reversal of the direction of flow
The symbol is a combination of 6.3.1.1 and 5.2.3 (sloping arrow)
6.3.2.2
— with one single direction of flow
The symbol is a combination of 6.3.1.2 and 5.2.3 (sloping arrow)
6.3.2.3
— with two directions of flow
The symbol is a combination of 6.3.1.3 and 5.2.3 (sloping arrow)
6.4
VARIABLE SPEED DRIVE UNITS
Torque converter. Pump and/or motor are variable capacity. Remote drives, see 12.2
6.5
CYLINDERS
Equipment to convert hydraulic or pneumatic energy into linear energy
6.5.1
Single acting cylinder:
Detailed
Simplified
Cylinder in which the fluid pressure always acts in one and the same direction (on the extension stroke)
6.5.1.1
— returned by an unspecified force
General symbol when the method of return is not specified
6.5.1.2
— returned by spring
Combination of the general symbols 6.5.1.1 and 5.1.5.2 (spring)
, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, , ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, , ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, , ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 6.5.2
6.5.2.1
Double acting cylinder:
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL Cylinder in which pressure fluid operates alternately in both directions (extend and retract strokes) ,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, , , ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, , ,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, , ,,,,, ,,,,,, , , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,, ,,,,,
— with single piston rod
6.5.2.2
— with double-ended piston rod
6.5.3
Differential cylinder
6.5.4
Cylinder with cushion:
6.5.4.1
— with single fixed cushion
6.5.4.2
— with double fixed cushion
6.5.4.3
— with single adjustable cushion
6.5.4.4
— with double adjustable cushion
6.5.5
Telescopic cylinder:
6.5.5.1
— single acting
6.5.5.2
— double acting
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SYMBOL
,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, , , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, , ,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,
The action is dependent on the difference between the effective areas on each side of the piston
Cylinder incorporating fixed cushion acting in one direction only Cylinder with fixed cushion acting in both directions The symbol is a combination of 6.5.4.1 and 5.2.3 (sloping arrow) The symbol is combination of 6.5.4.2 and 5.2.3 (sloping arrow)
The fluid pressure always acts in one and the same direction (on the extend stroke) The fluid pressure operates alternately in both directions (extend and retract strokes)
, ,, ,, ,, ,,, , ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 6.6
PRESSURE INTENSIFIERS:
6.6.1
— for one type of fluid
6.6.2
— for two types of fluid
6.7
AIR-OIL ACTUATOR
7
CONTROL VALVES
7.1
METHOD OF REPRESENTATION OF VALVES (EXCEPT 7.3 AND 7.6)
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
SYMBOL Detailed
Simplified
Equipment transforming a pressure x into a higher pressure y
x
y
x
y
E.g. a pneumatic pressure x is transformed into a higher pneumatic pressure y
x
y
x
y
E.g. a pneumatic pressure x is transformed into a higher hydraulic pressure y
,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Equipment transforming a pneumatic pressure into a substantially equal hydraulic pressure or vice versa
Made up of one or more squares 5.1.3 and arrows In circuit diagrams hydraulic and pneumatic units are normally shown in the unoperated condition
7.1.1
One single square
Indicates unit for controlling flow or pressure, having in operation and infinite number of possible positions between its end positions so as to vary the conditions of flow across one or more of its ports, thus ensuring the chosen pressure and/or flow with regard to the operating conditions of the circuit
7.1.2
Two or more squares
Indicate a directional control valve having as many distinct positions as there are squares. The pipe connections are normally represented as representing the unoperated condition (see 7.1). The operating positions are deduced by imagining the boxes to be displaced so that the pipe connections correspond with the ports of the box in question
1998/1999 Fluid Power Handbook & Directory
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
7.1.3
Simplified symbol for valves in cases of multiple repetition
7.2
DIRECTIONAL CONTROL VALVES
Units providing for the opening (fully or restricted) or the closing of one or more paths (represented by several squares)
7.2.1
Flow paths:
Square containing interior lines
7.2.1.1
— one flow path
7.2.1.2
— two closed ports
7.2.1.3
— two flow paths
7.2.1.4
— two flow paths and one closed port
7.2.1.5
— two flow paths with cross connection
7.2.1.6
— one flow path in a bypass position, two closed ports
7.2.2
Non-throttling directional control valve
3
The number refers to a note on the diagram in which the symbol for the valve is given in full
The unit provides distinct circuit conditions each depicted by a square
7.2.2.1
Basic symbol for 2-position directional control valve
7.2.2.2
Basic symbol for 3-position directional control valve
7.2.2.3
A transitory but significant condition between two distinct positions is optionally represented by a square with dashed ends
A/234
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 7.2.2.4
Designation: The first figure in the designation shows the number of ports (excluding pilot ports) and the second figure the number of distinct positions
7.2.2.5
Directional control valve 2/2:
7.2.2.5.1
— with manual control
7.2.2.5.2
— controlled by pressure operating against a spring (e.g., on air unloading valve)
7.2.2.6
Directional control valve 3/2:
7.2.2.6.1
— controlled by pressure in both directions
7.2.2.6.2
— controlled by solenoid with return spring
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
Directional control valve with 2 ports and 2 distinct positions
Directional control valve with 3 ports and 2 distinct positions
Indicating an intermediate condition (see 7.2.2.3)
P
,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,
A
P
,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, T
7.2.2.7
Directional control valve 4/2:
7.2.2.7.1
— controlled by pressure in both directions by means of pilot valve (with single solenoid and spring return)
Detailed
Directional control valve with 4 ports and 2 distinct positions
Simplified
7.2.2.8
Directional control valve 5/2:
7.2.2.8.1
— controlled by pressure in both directions
A P B ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,, T
Directional control valve with 5 ports and 2 distinct positions
1998/1999 Fluid Power Handbook & Directory
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 7.2.3
Throttling directional control
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL The unit has 2 extreme positions and an infinite number of intermediate conditions with varying degrees of throttling All the symbols have parallel lines along the length of the boxes. For valves with mechanical feedback see 9.3
7.2.3.1
Showing the extreme positions
7.2.3.2
Showing the extreme positions and a central (neutral) position
7.2.3.3
— with 2 ports (one throttling orifice)
For example: Tracer valve plunger operated against a return spring
7.2.3.4
— with 3 ports (two throttling orifices)
For example: Directional control valve controlled by pressure against a return spring
7.2.3.5
— with 4 ports (four throttling orifices)
For example: Tracer valve, plunger operated against a return spring
7.2.4
Electro-hydraulic servo valve: Electro-pneumatic servo valve:
A unit which accepts an analog electrical signal and provides a similar analog fluid power output
— single-stage
— with direct operation
7.2.4.1
Torque motor
Spool
,,,, ,,,, ,,,, ,,,, ,,,, ,,,, ,,,, ,,,,
,,, ,,, ,,, ,,, ,,, ,,, ,,, ,,,
,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,, ,,,,,
T B
,,,, ,,,, ,,,, ,,,, ,,,, ,,,, ,,,, ,,,,
P
,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,, ,, ,,,,, ,, ,,,,,
A T
Torque motor armature
7.2.4.2
— two-stage with mechanical feedback
— with indirect pilot operation
7.2.4.3
— two-stage with hydraulic feedback
— with indirect pilot operation
A/236
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL Valves which allow free flow in one direction only
7.3
NON-RETURN VALVES, SHUTTLE VALVE, RAPID EXHAUST VALVE
7.3.1
Non-return valve
7.3.1.1
— free
Opens if the inlet pressure is higher than the outlet pressure
7.3.1.2
— spring loaded
Opens if the inlet pressure is greater than the outlet pressure plus the spring pressure
7.3.1.3
— pilot controlled
As 7.3.1.1 but by pilot control it is possible to prevent
7.3.1.3.1
7.3.1.3.2
P ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
— a pilot signal closes the valve
— a pilot signal opens the valve
Outlet
7.3.1.4
— with restriction
Unit allowing free flow in one direction but restricted flow in the other
7.3.2
Shuttle valve
The inlet port connected to the higher pressure is automatically connected to the outlet port while the other inlet port is closed
7.3.3
Rapid exhaust valve
When the inlet port is unloaded the outlet port is freely exhausted
7.4
PRESSURE CONTROL VALVES
Units ensuring the control of pressure. Represented by one single square as in 7.1.1 with one arrow (the tail to the arrow may be placed at the end of the arrow). For interior controlling conditions see 9.2.4.3
7.4.1
Pressure control valve:
General symbols
7.4.1.1
— 1 throttling orifice normally closed
Pilot
Drain
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
7.4.1.2
— 1 throttling orifice normally open
7.4.1.3
— 2 throttling orifices, normally closed
7.4.2
Pressure relief valve (safety valve):
Inlet pressure is controlled by opening the exhaust port to the reservoir or to atmosphere against an opposing force (for example a spring)
7.4.2.1
— with remote pilot control
The pressure at the inlet port is limited as in 7.4.2 or to that corresponding to the setting of a pilot control
7.4.3
Proportional pressure relief
Inlet pressure is limited to a value proportional to the pilot pressure (see 9.2.4.1.3)
7.4.4
Sequence valve
When the inlet pressure overcomes the opposing force of the spring, the valve opens permitting flow from the outlet port
,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, System ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
T
,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Outlet P
7.4.5
Pressure regulator or reducing valve (reducer of pressure):
7.4.5.1
— without relief port
7.4.5.2
— without relief port with remote control
7.4.5.3
— with relief port
A/238
A unit which, with a pressure variable inlet pressure, gives substantially constant output pressure provided that the inlet pressure remains higher than the required outlet pressure As in 7.4.5.1, but the outlet pressure is dependent on the control pressure
T
,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Outlet P
T
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
7.4.5.4.
— with relief port, with remote control
As in 7.4.5.3, but the outlet pressure is dependent on the control pressure
7.4.6
Differential pressure regulator
The outlet pressure is reduced by a fixed amount with respect to the inlet pressure
7.4.7
Proportional pressure regulator
The outlet pressure is reduced by a fixed ratio with respect to the inlet pressure (see 9.2.4.1.3)
7.5
FLOW CONTROL VALVES
Units ensuring control of flow excepting 7.5.3 positions and method of representation as 7.4
7.5.1
Throttle valve:
Simplified symbol (does not indicate the control method or the state of the valve)
,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,, ,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Inlet
7.5.1.1
— with manual control
7.5.1.2
— with mechanical control against a return spring (braking valve)
7.5.2
Flow control valve:
7.5.2.1
— with fixed output
7.5.2.2
— with fixed output and relief port to reservoir
Detailed symbol (indicates the control method of the state of the valve)
Detailed
Simplified
Variations in inlet pressure do not affect the rate of flow
Outlet
Control orifice ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
Outlet
Inlet
Fixed orifice As 7.5.2.1 but with relief for excess flow
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 7.5.2.3
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL As 7.5.2.1 but with arrow 5.2.3 added to the symbol of restriction
— with variable output
Control chamber
Inlet
7.5.2.4
As 7.5.2.3 but with relief for excess flow
— with variable output and relief port to reservoir
Outlet Tank
Vent connection
7.5.3
Flow dividing valve
The flow is divided into two flows in a fixed ratio substantially independent of pressure variations
7.6
SHUT-OFF VALVE
Simplified symbol
8
ENERGY TRANSMISSION AND CONDITIONING
8.1
SOURCES OF ENERGY
8.1.1
Pressure source
Simplified general symbol
8.1.1.1
Hydraulic pressure source
Symbols to be used when the nature of the source should be indicated
8.1.1.2
Pneumatic pressure source
8.1.2
Electric motor
M 8.1.3
Heat engine
8.2
FLOW LINES AND CONNECTIONS
8.2.1
Flow line:
8.2.1.1
— working line, return line and feed line
8.2.1.2
— pilot control line
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M
Symbol 113 in IEC Publication 117.2
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 8.2.1.3
— drain or bleed line
8.2.1.4
— flexible pipe
8.2.1.5
— electric line
8.2.2
Pipeline junction
8.2.3
Crossed Pipelines
8.2.4
Air bleed
8.2.5
Exhaust port:
8.2.5.1
— plain with no provision for connection
8.2.5.2
— threaded for connection
8.2.6
Power take-off:
8.2.6.1
— plugged
8.2.6.2
— with take-off line
8.2.7
Quick-acting coupling:
8.2.7.1
— connected, without mechanically opened non-return valve
8.2.7.2
— connected, with mechanically opened non-return valves
8.2.7.3
— uncoupled, with open end
8.2.7.4
— uncoupled, closed by free non-return valve (see 7.3.1.1)
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
SYMBOL
Flexible hose, usually connecting moving parts
not connected
On equipment or lines, for energy take-off or measurement
,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,
,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,
,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,, ,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,,
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 8.2.8
Rotary connection:
8.2.8.1
— one way
8.2.8.2
— three way
8.2.9
Silencer
8.3
RESERVOIRS
8.3.1
Reservoir open to atmosphere:
8.3.1.1
— with inlet pipe above fluid level
8.3.1.2
— with inlet pipe below fluid level
8.3.1.3
— with a header line
8.3.2
Pressurized reservoir
8.4
ACCUMULATORS
8.5
FILTERS, WATER TRAPS, LUBRICATORS AND MISCELLANEOUS APPARATUS
8.5.1
Filter or strainer
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL Line junction allowing angular movement in service
The fluid is maintained under pressure by a spring, weight or compressed gas (air, nitrogen, etc.)
Air or gas
Bowl
8.5.2
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Water trap
Filer element
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION 8.5.2.1
— with manual control drain
8.5.2.2
— automatically drained
8.5.3
Filter with water trap:
8.5.3.1
— with manual control
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
Combination of 8.5.1 and 8.5.2.1 Float
8.5.3.2
— automatically drained
Combination of 8.5.1 and 8.5.2.2
8.5.4
Air dryer
A unit drying air (for example, by chemical means) Desiccant
8.5.5
Lubricator
8.5.6
Conditioning unit
8.5.6.1
— Detailed symbol
8.5.6.2
— Simplified symbol
8.6
HEAT EXCHANGERS
Small quantities of oil are added to the air passing through the unit, in order to lubricate equipment receiving the air
,,,, ,,,, ,,,, ,,,, ,,,, ,,,,
,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,, ,,,,,,
Inlet air
Lubricated air
Consisting of filter, pressure regulator, pressure gage and lubricator
Apparatus for heating or cooling the circulating fluid
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
8.6.1
Temperature controller
The fluid temperature is maintained between two predetermined values. The arrows indicate that heat may be either introduced or dissipated
8.6.2
Cooler
The arrows in the diamond indicate the extraction of heat
8.6.2.1
— without representation of the flow lines of the coolant
8.6.2.2
— indicating the flow lines of the coolant
8.6.3
Heater
9.
CONTROL MECHANISMS
9.1
Mechanical components
9.1.1
Rotating shaft:
9.1.1.1
— in one direction
9.1.1.2
— in either direction
9.1.2
Detent
9.1.3
Locking device
9.1.4
Over-center device
9.1.5
Pivoting devices:
9.1.5.1
— simple
The arrows in the diamond indicate the introduction of heat
The arrow indicates rotation
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A device for maintaining a given position * The symbol for unlocking control is inserted in the square
Prevents the mechanism from stopping in a dead center position
ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL
9.1.5.2
— with traversing lever
9.1.5.3
— with fixed fulcrum
9.2
CONTROL METHODS
The symbols representing control methods are incorporated in the symbol of the controlled apparatus, to which they should be adjacent. For apparatus with several squares the actuation of the control makes effective the square adjacent to it.
9.2.1
Muscular control:
General symbol (without indication of control type)
9.2.1.1
— by pushbutton
9.2.1.2
— by lever
9.2.1.3
— by pedal
9.2.2
Mechanical control:
9.2.2.1
— by plunger or tracer
9.2.2.2
— by spring
9.2.2.3
— by roller
9.2.2.4
— by roller, operating in one direction only
9.2.3
Electrical control:
9.2.3.1
— by solenoid
9.2.3.1.1
9.2.3.1.2
— with one winding
— with two windings operating in opposite directions
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ISO FLUID POWER GRAPHIC SYMBOLS DESCRIPTION
SYMBOL
USE OF THE EQUIPMENT OR EXPLANATION OF THE SYMBOL — with two windings operating in a variable way progressively, operating in opposite direction
9.2.3.1.3
M
9.2.3.2
— by electric motor
9.2.4
Control by application or release of pressure
9.2.4.1
Direct acting control:
9.2.4.1.1
— by application of pressure
9.2.4.1.2
— by release of pressure
9.2.4.1.3
— by different control areas
In the symbol the larger rectangle represents the larger control area, i.e., the priority phase
9.2.4.2
Indirect control, pilot actuated:
General symbol for pilot directional control valve
9.2.4.2.1
— by application of pressure
9.2.4.2.2
— by release of pressure
9.2.4.3
Interior control paths
9.2.5
Combined control:
9.2.5.1
— by solenoid and pilot directional valve
The pilot directional valve is actuated by the solenoid
— by solenoid or pilot directional valve
Either may actuate the control independently
The control paths are inside the unit
9.2.5.2
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ISO FLUID POWER GRAPHIC SYMBOLS
LOGIC ELEMENT
AND
LOGIC ELEMENT FUNCTION
OUTPUT IF ALL CONTROL INPUT SIGNALS ARE ON
BOOLEAN ALGEBRA SYMBOL
ISO/R 1219-1970 GRAPHIC SYMBOLS FLUID POWER
OR
(A) • (B)
NOT
OUTPUT IF ANY ONE OF THE CONTROL INPUTS IS ON
OUTPUT IF SINGLE CONTROL INPUT SIGNAL IS OFF
(A) + (B)
A
B
Supply
ANSI B93.38-1976 MOVING PARTS FLUID CONTROLS ATTACHED
A
RV1
A
S 1
RV1 B
A
RV2
RV3
B Supply
RV4
RV4
N
B
B
RV5 Supply
N
Supply
RV5
A
A A
S 2
RV3
Supply
A
Supply
A
RV2
B
B
B
Supply
A A
(A) + (B)
A
B ANSI B93.38-1976 MOVING PARTS FLUID CONTROLS DETACHED
OUTPUT IF ALL CONTROL INPUT SIGNALS ARE OFF
(A) • (B)
A
B
NOR
NO OUTPUT IF ALL CONTROL INPUT SIGNALS ARE ON
(A)
A
A
NAND
B
Supply
N B
Supply
MIL-STD-806B LOGIC SYMBOL
NEMA LOGIC SYMBOL
D
ELECTRICAL RELAY LOGIC SYMBOL D
D
D D
D
D D
D
D
ELECTRICAL SWITCH LOGIC SYMBOL
FLUIDIC DEVICE TURBULENCE AMPLIFIER PROPOSED N.F.P.A.–A.S.A. SYMBOL
NOR LOGIC EQUIVALENT OF PROPOSED N.F.P.A.–A.S.A. SYM.
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ISO FLUID POWER GRAPHIC SYMBOLS
FLIP FLOP
A SIGNAL TO ONE INPUT TURNS A CORRESPONDING OUTPUT ON AND THE OTHER OUTPUT OFF
MEMORY (OFF RETURN)
DIFFERENTIATIOR (SINGLE SHOT)
ON DELAY TIMER (TIMING IN)
OFF DELAY TIMER (TIMING OUT)
MOMENTARY INPUT SIGNAL(S) PRODUCES AN OUTPUT UNTIL RESET(R)
PRODUCES A SHORT OUTPUT PULSE WHEN INPUT SIGNAL IS ON
PRODUCES AN OUTPUT FOLLOWING A DEFINITE DELAY AFTER INPUT IS PRESENT
REMOVES AN OUTPUT FOLLOWING A DEFINATE DELAY AFTER INPUT IS REMOVED
LOGIC ELEMENT
LOGIC ELEMENT FUNCTION
BOOLEAN ALGEBRA SYMBOL S
S
Supply
R
R
S
A
R
RV6
B
RV6 RV6
Supply
ISO/R 1219-1970 GRAPHIC SYMBOLS FLUID POWER Supply
S
Supply
S 3
RV7
Supply
RV8
S FF
R
DEL
1
ANSI B93.38-1976 MOVING PARTS FLUID CONTROLS ATTACHED
Supply
T
FF
DEL
Supply
Supply
Supply
S
RV9
MEM
R
RV10
ANSI B93.38-1976 MOVING PARTS FLUID CONTROLS DETACHED
RV10
RV9
RV7
R
S
RV8
MIL-STD-806B LOGIC SYMBOL
SS 5 MS
C 0
NEMA LOGIC SYMBOL
D
D
D D D
TD
TD
D D
TD
ELECTRICAL RELAY LOGIC SYMBOL
TD
ELECTRICAL SWITCH LOGIC SYMBOL
FLUIDIC DEVICE TURBULENCE AMPLIFIER PROPOSED N.F.P.A.–A.S.A. SYMBOL
NOR LOGIC EQUIVALENT OF PROPOSED N.F.P.A.–A.S.A. SYM.
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Fluid power glossary Absorption — The physical mechanism by which one substance attracts and takes up another substance (liquid, gas, or vapor) into its interior. Accelerator — A substance which hastens the vulcanization of an elastomer, causing it to take place in a shorter time or at a lower temperature. Accumulator — A container in which fluid is stored under pressure as a source of fluid power. Accumulator, hydropneumatic bladder — A hydropneumatic accumulator in which the liquid and gas are separated by an elastic bag or bladder. Actuator, pneumatic/hydraulic — A device in which power is transferred from one pressurized medium (pneumatic) to another (hydraulic) without intensification. Additive — A chemical added to a fluid to impart new properties or to enhance those which already exist. Adsorption — The physical mechanism by which one substance attracts another substance (either solid, liquid, gas, or vapor) to its surface and through molecular forces causes the incident substance to adhere to that surface. Aftercooler — A device which cools a gas after it has been compressed. Afterfilter — A filter which follows the compressed air dryer and usually for the protection of downstream equipment from desiccant dust. Air bleeder — A device for removal of air. Air breather — A device permitting air movement between atmosphere and the component in which it is installed. Air motor — A device which converts pneumatic fluid power into mechanical torque and motion. It usually provides rotary mechanical motion. Air, compressed (pressure) — Air at any pressure greater than atmospheric pressure. Air, dried — Air with moisture content lower than the maximum allowable for a given application. Air, free — Air at ambient temperature, pressure, relative humidity, and density. Air, saturated — Air at 100% relative humidity, with a dew point equal to temperature. Air, standard — Air at a temperature of 68.8° F, a pressure of 14.70 pounds per square inch absolute, and a relative humidity of 36% (0.0750 pounds per cubic foot). In gas industries the temperature of “standard air” is usually given as 60.8° F.
Air — A gas mixture consisting of nitrogen, oxygen, argon, carbon dioxide, hydrogen, small quantities of neon, helium and other gases. Amplification, power — The ratio between the output signal variations and the corresponding input (control) power variation (for analog devices only). Amplification, pressure — Ratio between outlet pressure and inlet (control) pressure. Amplification — The ratio between the output signal variations and the control signal variations (for analog devices only). Analog — Of or pertaining to the general class of fluidic devices or circuits whose output varies as a continuous function of its input. AND Device — A control device which has its output in the logical 1 state if and only if all the control signals assume the logical 1 state. Aniline point — The lowest temperature at which a liquid is completely miscible with an equal volume of freshly distilled aniline (ASTM Designation D611-64). Aniline point — The lowest temperature at which equals volumes of pure, fresh aniline and an oil will completely dissolve in one another is the aniline point of the oil. Bernoulli’s Law — If no work is done on or by a flowing frictionless liquid, its energy due to pressure and velocity remains constant at all points along the streamline. Bleeding — Migration to the surface of plasticizers, waxes, or similar materials to form a film or beads. Boyle’s Law — The absolute pressure of a fixed mass of gas varies inversely as the volume, provided the temperature remains constant. Break-out — Force necessary to inaugurate sliding. Expressed in same terms as friction. An excessive break-out value indicates the development of adhesion. Breathing capacity — A measure of flow rate through an air breather. Bulk modulus — The measure of resistance to compressibility of a fluid. It is the reciprocal of compressibility. Cavitation — A localized gaseous condition within a liquid stream which occurs where the pressure is reduced to the liquid's vapor pressure, often as a result of a solid body, such as a propeller or piston, moving through the liquid; also, the pitting or wearing away of a solid surface as a result of low fluid levels that draw air into the system,
producing tiny bubbles that expand explosively at the pump outlet, causing metal erosion and eventual pump destruction. Charles’ Law — The volume of a fixed mass of gas varies directly with absolute temperature, provided the pressure remains constant. Circuit, meter-in — A speed control circuit in which the control is achieved by regulating the supply flow to the actuator. Circuit, meter-out — A speed control circuit in which the control is achieved by regulating the exhaust flow from the actuator. Circuit, open — A circuit in which return fluid is directed to the reservoir before reciprocation. Circuit, regenerative — A circuit in which pressurized fluid discharged from a component is returned to the system to reduce power input requirements. Circuit, sequence — A circuit which established the order in which two or more phases of a circuit occur. Circuit — An arrangement of interconnected components and parts. Cold flexibility — Flexibility following exposure to a predetermined time. Cold flow — Continued deformation under stress. Compatibility, seal — Ability of an elastomer to resist the action of a fluid on its dimensional and mechanical properties. Compressibility — The change in volume of a unit volume of a fluid when subjected to a unit change in pressure. Compression modulus — The ratio of the compressive stress to the resulting compressive strain (the latter expressed as a fraction of the original height or thickness in the direction of the force). Compression modulus may be either static or dynamic. Compression set — The amount by which a rubber specimen fails to return to original shape after release of the compressive load. Compressor — A device which converts mechanical force and motion into pneumatic fluid power. Condensation — The process of changing a vapor into a liquid condensate by the extraction of heat. Conditioner, air — An assembly comprising a filter, a pressure reducing valve with gage, and a lubricator, intended to deliver compressed air in suitable condition. Conductor — A component whose primary function is to contain and direct fluid. Contaminant — Any material or substance
1998/1999 Fluid Power Handbook & Directory
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GLOSSARY
which is unwanted or adversely affects the fluid power system or components, or both. Control — A device used to regulate the function of a component or system. Controller — A device which senses a change of fluid state and automatically makes adjustments to maintain the state of the fluid between predetermined limits, e.g., pressures, temperatures, etc. Copolymer — A polymer consisting of two different monomers chemically combined. Creep — The progressive relaxation of a given rubber material while it is under stress. This relaxation eventually results in permanent deformation or “set.” Cushion — A device which provides controlled resistance to motion. Cylinder cap — A cylinder end closure which completely covers the bore area. Cylinder capacity, extending — Volume required for one full extension of a cylinder. Cylinder capacity, retracting — Volume (annular) absorbed by one full retraction of the cylinder. Cylinder capacity — The volume of a theoretically incompressible fluid that would be displaced by the piston during a complete stroke. (For double acting cylinders it must be given for both directions of stroke.) Cylinder force, theoretical — The pressure multiplied by the effective piston area, ignoring friction. For double acting cylinders, the value must be given for both directions of stroke. Cylinder, adjustable stroke — A cylinder equipped with adjustable stops at one or both ends to limit piston travel. Cylinder, area, piston rod — Cross-sectional area of the piston rod. Cylinder, area, piston, effective — Area upon which fluid pressure acts to provide a mechanical force. Cylinder, bore — The internal diameter of the cylinder body. Cylinder, cushioned — A cylinder with a piston-assembly deceleration device at one or both ends of the stroke. Cylinder, differential — A double acting cylinder in which the ratio of the area of the bore to the annular area between the bore and the piston rod is significant in circuit function. Cylinder, double acting — A cylinder in which fluid force can be applied to the moveable element in either direction. Cylinder, double rod — A cylinder with a single piston and a piston rod extending from each end. Cylinder, dual stroke — A cylinder combination which provides two working strokes. Cylinder, duplex — A unit comprised of two cylinders with independent control, mechanically connected on a common axis
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to provide three or four positions depending on the method of application. Cylinder, piston type — A cylinder in which the piston has a greater cross-sectional area than the piston rod. Cylinder, plunger (ram) — A cylinder in which the piston has the same cross-sectional area as the piston rod. Cylinder, rotary actuator — A cylinder which translates piston reciprocation into oscillation of an output shaft. Cylinder, rotating — A cylinder in which the piston and piston rod, plunger or ram, is permitted to rotate with reference to the cylinder housing. Cylinder, single acting — A cylinder in which the fluid force can be applied to the movable element in only one direction. Cylinder, tandem — Arrangement of at least two pistons on the same rod moving in separate chambers on the same cylinder body allowing the compounding of force on the piston rod. Cylinder, telescoping — Cylinder with two or more stages or extensions, achieved by hollow piston rods sliding one within the other (may be single or double acting). Cylinder, tie rod — A cylinder with head and cap end closures that are secured by tie rods. Cylinder — A device which converts fluid power into linear mechanical force and motion. It usually consists of a movable elements such as a piston and piston rod, plunger or ram, operating within a cylindrical bore. Darcy’s Formula — A formula used to determine the pressure drop due to flow friction through a conduit. Deliquescent — Moisture is separated by using the absorptive properties of special hygroscopic compounds. Desiccant — Material that tends to remove moisture from compressed air. Dew point — The temperature at which vapors in a gas condense. For practical purposes, it must be referred to a stated pressure. Digital — Of or pertaining to the general class of fluidic devices or circuits whose output varies in discrete steps (i.e., pulses or “on-off” characteristics). Displacement, volumetric — Volume absorbed or displaced per stroke of a cylinder or per cycle of a pump or motor. Dissolved air — Air which is dispersed at a molecular level in hydraulic fluid to form a single phase. Dissolved water — Water which is dispersed at a molecular level in hydraulic fluid to form a single phase. Dither — A low amplitude, relatively high frequency periodic electrical signal, sometimes superimposed on the servovalve input to improve system resolution. Dither is ex-
pressed by the dither frequency (Hz) and the peak-to-peak dither current amplitude. Droop — The deviation between no flow secondary pressure and secondary pressure at a given flow. Dryer, compressed air — A device for reducing the moisture content of the working compressed air. Durometer — 1. An instrument for measuring the hardness of rubber. Measures the resistance to the penetration of an indentor point into the surface of rubber. 2. Numerical scale of rubber hardness. Efficiency — Ratio of output to the corresponding input. Elasticity — The property of a material which tends to return to its original shape after deformation. Elastomer — Any synthetic or natural material with resilience or memory sufficient to return to its original shape after distortion. Elongation — Generally means “ultimate elongation” or percent increase in original length of a specimen when it breaks. Emulsion, oil in water — A dispersion of oil in a continuous phase of water. Emulsion, water in oil — A dispersion of water in a continuous phase of oil. Emulsifier — additive that promotes formation of a stable mixture, or emulsion, of oil and water. Emulsion — A homogeneous dispersion of two immiscible liquids, generally of a milky or cloudy appearance. Entrained air — A mechanical mixture of air bubbles having a tendency to separate from the liquid phase. Expectancy, life — The predicted working period during which a component or system will maintain a specified level of performance under specified conditions. Sometimes expressed in statistical terms as a probability. Filter — 1. A device whose primary function is the removal by porous media of insoluble contaminants from a liquid or a gas. 2. Chemically inert, finely divided material added to the elastomer to aid in processing and improve physical properties. Filter, strainer — A coarse hydraulic filter usually of woven wire construction. This may be in the form of a complete filter or just an element. Filter, by-pass (reserve) –– A filter which provides an alternate unfiltered flow path around the filter element when a preset differential pressure is reached. Filter, spin-on — A filter with spin-on element sealed in its own pressure housing for independent mounting to the filter. Filtration ratio (bm) — The ratio of the number of particles greater than a given size (b) in the influent fluid to the number of particles greater than the same size (m) in
GLOSSARY
the effluent fluid. Fitting — A connector or closure for fluid power lines and passages. Fitting, compression — A fitting which seals and grips by manual adjustable deformation. Fitting, flange — A fitting which utilizes a radially extending collar for sealing and connection. Fitting, flared — A fitting which seals and grips by a pre-formed flare at the end of the tube. Fitting, flareless — A fitting which seals and grips by means other than a flare. Flash point — The temperature to which a liquid must be heated under specified conditions of the test method to give off sufficient vapor to form a mixture with air that can be ignited momentarily by a flame. Flip flop — A digital component or circuit with two stable states and sufficient hysteresis so that it has “memory.” Its state is changed with a control pulse; a continuous control signal is not necessary for it to remain in a given state. Flow characteristic curve — The change in regulated (secondary) pressure occurring as a result of a change in the rate of air flow over the operating range of the regulator. Flow rate — The volume, mass or weight of a fluid passing through any conductor per unit of time. Flow, laminar (streamline) — A flow situation in which fluid moves in parallel lamina or layers. Flow, output — Flow rate discharged at the outlet port. Flow, turbulent — A flow situation in which the fluid particles move in a random fluctuating manner. Flow — Movement of fluid generated by pressure differences. Fluid capacity — The liquid volume coincident with the “high” mark of the level indicator. Fluid friction — Friction due to the viscosity of fluids. Fluid logic—A branch of fluid power associated with digital signal sensing and information processing, using components with or without moving parts. Fluid miscibility — Capacity of fluids to be mixed in any ratio without separation into phases. Fluid power system — A system that transmits and controls power through use of a pressurized fluid within an enclosed circuit. Fluid power — Energy transmitted and controlled through use of a pressurized fluid. Fluid stability — Resistance of a fluid to permanent changes in properties. Fluid stability, oxidation — Resistance of a fluid to permanent changes caused by
chemical reaction with oxygen. Fluid, anti-corrosive — A fluid containing metal corrosion inhibitors. Fluid, aqueous — A fluid which contains water as a major constituent besides the organic material. The fire resistance properties are derived from the water content. Fluid, fire resistant — A fluid difficult to ignite which shows little tendency to propagate flame. Fluid, hydraulic — A fluid suitable for use in a hydraulic system. Fluid, Newtonian — Fluid having a viscosity that is always independent of the rate of shear. Fluid, pneumatic — A fluid suitable for use in a pneumatic system, usually air. Fluid, rust protection — Capacity of a fluid to prevent the formation of rust under specified conditions. Fluid — A liquid, gas or combination thereof.
For more fluid terms, see glossary in hydraulic fluids chapter. Force motor — A type of electromechanical transducer having linear motion used in the input stages of servovalves. Free air — Any compressible gas, air or vapor trapped within a hydraulic system that does not condense or dissolve to form a part of the system fluid. Free water — Water droplets or globules in the system fluid that tend to accumulate at the bottom or top of the system fluid depending on the fluid’s specific gravity. Frequency response — The changes, under steady-state conditions, in the output variable which are caused by a sinusoidal input variable. Gage damper (snubber) — A device employing a fixed or variable restrictor inserted in the pipeline to a pressure gage, to prevent damage to the gage mechanism caused by rapid fluctuations of fluid pressure. Gage protector — A device inserted in the pipeline to a pressure gage and arranged to isolate the pressure gage from the fluid pressure if this exceeds a predetermined limit. The device can usually be adjusted to suit the range of the pressure gage. Gage, bourdon tube — A pressure gage in which the sensing element is a curved tube that tends to straighten out when subjected to internal fluid pressure. Gage, diaphragm — A gage in which the sensing element is relatively thin and its inner portion is free to deflect with respect to its periphery. Gage, instrument — An instrument or device for measuring, indicating, or comparing a physical characteristic.
Gage, pressure — A gage which indicates the pressure in the system to which it is connected. Head — The height of a column or body of fluid above a given point expressed in linear units. Head is often used to indicate gage pressure. Pressure is equal to the height times the density of the fluid. Head, cylinder — The cylinder end closure which covers the differential area between the bore area and the piston rod area. Head, friction — The pressure required to overcome the friction at the interior surface of a conductor and between fluid particles in motion. It varies with flow, size, type and condition of conductors and fittings, and the fluid characteristics. Head, pressure — The pressure due to the height of a column or body of fluid. Head, static — The height of a column or body of fluid above a given point. Heat exchanger — A device which transfers heat through a conducting wall from one fluid to another. (Typically to cool a system.) Heater — A device which transfers heat through a conducting wall from one fluid to another. (Typically to warm up a system.) Hose, wire braided — Hose consisting of a flexible material reinforced with woven wire braid. Hose — A flexible line or conductor whose nominal size is its inside diameter. Hydraulic amplifier — A fluid device which enables one or more inputs to control a source of fluid power and thus is capable of delivering at its output an enlarged reproduction of the essential characteristics of the input. Hydraulic amplifiers may utilize sliding spools, nozzle-flappers, jet pipes, etc. Hydraulic motor — A device which converts hydraulic fluid power into mechanical force and motion. It usually provides rotary mechanical motion. Hydraulic motor efficiency, hydromechanical — Ratio of the effective torque to the derived torque. Hydraulic motor efficiency, overall — Ratio of the output power to the effective hydraulic power. Hydraulic motor efficiency, volumetric — Ratio of the derived output flow to the effective input flow. Hydraulic motor, fixed displacement — A hydraulic motor in which the displacement per unit of output motion cannot be varied. Hydraulic motor, flow, input — Flow rate crossing the transverse plane of the inlet port. Hydraulic motor, gear, external — A motor having two or more external gears. Hydraulic motor, gear, internal — A motor with an internal gear in engagement with
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GLOSSARY
one or more external gears. Hydraulic motor, gear — A motor in which two or more gears act in arrangement as working members. Hydraulic motor, vane — A motor in which the fluid under pressure acting on a set of radial vanes causes rotation of an internal member. Hydraulic stepping motor — A hydraulic motor which follows the commands of a stepped input signal to achieve positional accuracy. Hydraulics — Engineering science pertaining to liquid pressure and flow. Hydrodynamics — The engineering science which governs the movement of liquids and the forces opposing that movement. Hydrokinetics — Engineering science pertaining to the energy of liquid flow and pressure. Hydropneumatics — Pertaining to the combination of hydraulic and pneumatic fluid power. Hydrostatic transmission — Combination of one or more hydraulic pumps and motors forming a unit. Hydrostatics — Engineering science pertaining to the energy of liquids at rest. Indicator, differential pressure — An indicator which signals a difference in pressure between two points in a fluid power system. Inhibitor — Any substance which, when present in very small proportions, slows, prevents or modifies chemical reactions such as corrosion or oxidation. Intensification, ratio of — The ratio of the secondary pressure to the primary pressure or of the primary flow rate to the secondary flow rate. Intensifier, double acting — A unit which magnifies the secondary fluid pressure regardless of the direction of flow of the primary fluid. Intensifier, single acting — A unit which only magnifies the fluid pressure in one direction of flow of the primary fluid. Intensifier, single shot — An intensifier in which the continuous application of primary fluid at the inlet port can only give a limited volume of secondary fluid. Intensifier — A device which converts low pressure fluid power into higher pressure fluid power. Joint — A line positioning connector. Joint, rotary — A joint connecting lines which have relative operational rotation. Leakage rate — The rate at which a gas or liquid passes through a barrier. Total leakage rate includes the amounts that diffuse or permeate through the material of the barrier as well as the amount that escapes around it. Line, return — A pipe (conductor) to re-
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turn the working fluid to the reservoir. Line, working — A line which conducts fluid power. Line — A tube, pipe, or hose for conducting fluid. Lubricator — A device which adds controlled or metered amounts of lubricants into a fluid power system. Magnetic plug — A plug which attracts and holds ferromagnetic particles. Manifold — A conductor which provides multiple connection ports. Maximum inlet pressure — The maximum rated gage pressure applied to the inlet port of the regulator. Memory — Tendency of a material to return to original shape after deformation. Modulus of elasticity — One of the several measurements of stiffness or resistance to deformation, but often incorrectly used to indicate specifically static tension modulus. Modulus — Tensile stress at a specified elongation. (Usually 100% elongation for elastomers.) Moving parts logic — The technology of achieving logic control by means of fluid devices having moving parts. Muffler — A device for reducing gas flow noise. Noise is decreased by back pressure control of gas expansion. Newt — A unit of kinematic viscosity in the English system. It is expressed in square inches per second (see Stokes). NOR device — A control devices which has its output in the logical 1 state if and only if all the control signals assume the logical 0 state. NOT device — A control device which has its output in the logical 1 state if and only if the control signal assumes the logical 0 state. The NOT device is a single input NOR device. Oil swell — The change in volume of a rubber article due to absorption of oil or other fluid. OR device — A control device which has its output in the logical 0 state if and only if all the control signals assume the logical 0 state. Outgassing — A vacuum phenomenon wherein a substance spontaneously releases volatile constituents in the form of vapors or gases. In rubber compounds, these constituents may include water vapor, plasticizers, air, inhibitors, etc. Output stage — The final stage of hydraulic amplifications used in a servovalve. Ozone resistance — Ability to withstand the deteriorating effect of ozone (which generally causes cracking). Packing — A sealing device consisting of bulk deformable material of one or more mating deformable elements, reshaped by manually adjustable compression to obtain
and maintain effectiveness. It usually uses axial compression to obtain radial sealing. Pascal’s Law — A pressure applied to a confined fluid at rest is transmitted with equal intensity throughout the fluid. Permanent set — The deformation remaining after a specimen has been stressed in tension for a definite period and released for a definite period. Permeability — The rate at which a liquid or gas under pressure passes through a solid material by diffusion and solution. In rubber terminology, it is the rate of gas flow expressed in atmospheric cubic centimeters per second through an elastomeric material one centimeter square and one centimeter thick (atm cm3/cm2•cm•sec). Petroleum fluid — A fluid composed of petroleum oil which may contain additives and/or inhibitors. Pipe — A conductor whose outside diameter is standardized for threading. Pipe is available in standard, extra strong, or double extra strong wall thickness. Piston rod — The element transmitting mechanical force and motion from the piston. Plasticizer — A substance, usually a heavy liquid, added to an elastomer to decrease stiffness, improve low temperature properties, and improve processing. Pneumatics — Engineering science pertaining to gaseous pressure and flow. Poise — The standard unit of dynamic viscosity in the cgs (centimeter gram second) system. It is the ratio of the shearing stress to the shear rate of fluid and is expressed in millipascal sec. (equals 1 centipoise). Polymer — A material formed by the joining together of many (poly) individual units (mer) of one or more monomers; synonymous with elastomers. Port — A terminus of a passage in a component to which conductors can be connected. Port, differential pressure — A port which provides a passage to the upstream and downstream sides of a component. Post cure — The second step in the vulcanization process for the more exotic elastomers. Provides stabilization of parts and drives off decomposition products resulting from the vulcanization process. Pour point — The lowest temperature at which a liquid will flow under specified conditions (ASTM Designation D97). Power unit — A combination of pump, pump drive, reservoir, controls and conditioning components to supply hydraulic power to a system. Pressure — Force per unit area, usually expressed in pounds per square inch (bar). Pressure, absolute — The pressure above zero absolute, i.e., the sum of atmospheric and gage pressure. In vacuum related work it is usually expressed in millimeters of
GLOSSARY
mercury (mm-Hg). Pressure, atmospheric — Pressure exerted by the atmosphere at any specific location. (Sea level pressure is approximately 14.7 pounds per square inch absolute. 1 bar = 14.5 psi). Pressure, back — The pressure encountered on the return side of a system. Pressure, breakloose (breakout) — The minimum pressure which initiates movement. Pressure, burst — The pressure which causes failure of and consequential loss of fluid through the product envelope. Pressure, charge — The pressure at which replenishing fluid is formed into a fluid power system. Pressure, control range — The permissible limits between which system pressure may be set. Pressure, cracking — The pressure at which a pressure-operated valve begins to pass fluid. Pressure, differential (pressure drop) — The difference in pressure between any two points of a system or a component. Pressure, gage — Pressure differential above or below ambient atmospheric pressure. Pressure, induced — Pressure generated by an externally applied force. Pressure, inlet — The pressure at the apparatus inlet port. Pressure, intensified — In a fluid power cylinder, the outlet pressure required to slow the piston rod extending under regulated pressure introduced at the cap end. Pressure, maximum inlet — The maximum rated gage pressure applied to the inlet. Pressure, nominal — A pressure valve assigned to a component or system for the purpose of convenient designation. Pressure, outlet — Pressure at the apparatus outlet port. Pressure, override — The difference between the cracking pressure of a valve and the pressure reached when the valve is passing its rated flow. Pressure, peak — The maximum pressure encountered in the operation of a component. Pressure, pilot — The pressure in the pilot circuit. Pressure, precharge — The pressure of compressed gas in an accumulator prior to the admission of a liquid. Pressure, proof — The non-destructive test pressure, in excess of the maximum rated operating pressure, which causes no permanent deformation, excessive external leakage, or other resulting malfunction. Pressure, rated — The qualified operating pressure which is recommended for a component or system by the manufacturer.
Pressure, shock — The pressure existing in a wave moving at sonic velocity. Pressure, static — The pressure in a fluid at rest. Pressure, surge — The pressure resulting from surge conditions. Pressure, system — The pressure which overcomes the total resistances in a system. It includes all losses as well as useful work. Pump — A device which converts mechanical torque and motion into hydraulic fluid power. Pump, fixed displacement — A hydraulic pump in which the volume displaced per cycle cannot be varied. Pump, gear, external — Pump with two or more external gears. Pump, gear, internal — Pump with an internal gear in engagement with one or more external gears. Pump, gear — Pump in which two or more gears act in engagement as pumping members. Pump, hydraulic — A device which converts mechanical force and motion into hydraulic fluid power. Pump, multiple stage — Two or more hydraulic pumps in series. Pump, piston, axial — Pump having several pistons with mutually parallel axes which are arranged around and parallel to a common axis. Pump, piston, inline — Pump having several pistons with mutually parallel axes arranged on a common plane. Pump, piston, radial — Pump having several pistons arranged to operate radially. Pump, piston — Pump in which the fluid volume is displaced by one or more reciprocating pistons. Pump, screw — A hydraulic pump having one or more screws rotating in a housing. Pump, vane, balanced — Pump in which the transverse forces on the rotor are balanced. Pump, vane, unbalanced — Pump in which the transverse forces on the rotor are not balanced. Pump, vane — A hydraulic pump having multiple radial vanes within a supporting rotor. Pump, variable displacement — A hydraulic pump in which the volume displaced per cycle can be varied. Pump-motor — Unit which functions either as a pump or as a rotary motor. Quick disconnect coupling — A component which can quickly join or separate a fluid line without the use of tools or special devices. Refrigerated dryer — Moisture is separated by lowering the air temperature by means of refrigeration compressor and heat exchanger. Regenerative dryer — The capacity of the dryer to separate moisture can be restored
without replacing the drying compound. Regulator, air line pressure — A regulator which transforms a fluctuating air pressure supply to provide a constant lower pressure output. Regenerative circuit — see Circuit, regenerative. Reinforcing agent — Material dispersed in an elastomer to improve compression, shear, or other stress properties. Reservoir (tank) — A container for storage of liquid in a fluid power system. Reservoir, hydraulic — A reservoir for storing and conditioning a liquid in a hydraulic system. Reservoir, pressure sealed — A sealed reservoir for storage of fluids under pressure. Resilient — Capable of returning to original size and shape after deformation. Reyn — The standard unit of absolute viscosity in the English system. It is expressed in pound-seconds per square inch. Reynolds Number — A numerical ratio of the dynamic forces of mass flow to the shear stress due to viscosity. Flow usually changes from laminar to turbulent between Reynolds Numbers 2000 and 4000. Ring, O — A ring which has a round crosssection. Ring, piston — A piston sealing ring. It is usually one of a series and is often split to facilitate expansion or contraction. Ring, scraper — A ring which removes material by a scraping action. Rotation — The direction of rotation is always quoted as viewed looking at the shaft end. In dubious cases, provide a sketch. Seal, cup — A sealing device with a radial base integral with an axial cylindrical projection at its outer diameter. Seal, dynamic — A sealing device used between parts that have relative motion. Seal, elastomer — A material having rubber-like properties; i.e., having the capacity for large deformation and rapid and substantially complete recovery on release from the deforming force. Seal, rod (shaft) — A sealing device which seals the periphery of a piston rod. Seal, static (gasket) — A sealing device used between parts that have no relative motion. Sensor — A device which detects and transmits changes in external conditions. Separator — A device whose primary function is to isolate contaminants by physical properties other than size. (Separators remove gas from liquid medium or remove liquid from gaseous medium). Servovalve — A valve which modulates output as a function of an input command. Servovalve, electrohydraulic — A servovalve which is capable of continuously controlling hydraulic output as a function of an
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GLOSSARY
electrical input. Servovalve, electrohydraulic, flow control — An electrohydraulic servovalve whose primary function is control of output flow. Servovalve hysteresis — The difference in the servovalve input currents required to produce the same output during a single cycle of valve input current when cycled at a rate below that at which dynamic effects are important. Servovalve null leakage — Total internal leakage from the valve in the null position. Servovalve, pressure control — A hydraulic servovalve whose primary function is the control of output pressure. Shrinkage — Decreased volume of seal, usually caused by extraction of soluble constituents by fluids followed by air drying. Silencer — A device for reducing gas flow noise. Noise is decreased by tuned resonant control of gas expansion. Snubber — see gage damper. Solenoid, digital — Electrically energized device which generates on-off signals. Solenoid, proportional — An electrical device that reacts proportionally to strength of electrical signal. Sorption — The term used to denote the combination of absorption and adsorption processes in the same substance. Specific gravity, liquid — The ratio of the weight of a given volume of liquid to the weight of an equal volume of water. Squeeze — Cross section diametral compression of O-ring between surface of the groove bottom and surface of other mating metal part in the gland assembly. Stage — A hydraulic amplifier used in a servovalve. Servovalves may be single stage, two stage, three stage, etc. Standard — A document, or an object for physical comparison, for defining product characteristics, products, or processes; prepared by a consensus of a properly constituted group of those substantially affected and having the qualifications to prepare the standard for voluntary use. Stokes — The standard unit kinematic viscosity in the cgs (centimeter gram second) system. It is expressed in square centimeters per second; 1 centistokes equals 0.01 stokes. Strainer — see filter, strainer. Surface tension — The surface force of a liquid in contact with a fluid by which it tends to assume a spherical form and to present the least possible surface. It is expressed in pounds per foot or dynes per centimeter. Surge — A transient rise of pressure or flow. Swell — Increased volume of specimen caused by immersion in a fluid (usually a liquid). Switch, float — An electric switch which is responsive to liquid level.
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Switch, flow — An electric switch operated by fluid flow. Switch, pressure differential — An electric switch operated by a difference in pressure. Switch, pressure — An electric switch operated by fluid pressure. Synthetic fluid, silicate ester — A fluid compound of organic silicates. It may contain additives. Synthetic fluid — Fluid other than mineral on which has been artificially compounded for use in a fluid power system. Temperature, ambient — The temperature of the environment in which an apparatus is working. Tensile strength — Force in pounds per square inch required to cause the rupture of a specimen of a rubber material. Terpolymer — A polymer consisting of three different monomers chemically combined. Tie rod — An axial external cylinder element which traverses the length of the cylinder. It is prestressed at assembly to hold the ends of the cylinder against the tubing. Tie rod extensions can be a mounting device. Torque motor — A type of electromechanical transducer having rotary motion used in the input stages of servovalves. Torque — Rotary force transmitted by the driving shaft of the pump. Torr — A unit of pressure equal to 1/760 of an atmosphere. Torricelli’s Theorem — The liquid velocity at an outlet discharging into the free atmosphere is proportional to the square root of the head. Transducer, flow — A device which converts fluid flow to an electrical signal. Transducer, pressure — A device which converts fluid pressure to an electrical signal. Trunnion — A mounting device consisting of a pair of opposite projecting cylindrical pivots. The cylindrical pivot pins are at right angle or normal to the piston rod centerline to permit the cylinder to swing in a plane. Tube — A conductor whose size is its outside diameter. Tube is available in varied wall thickness and material. Vacuum — Pressure less than ambient atmospheric pressure. Vacuum pump — A device which uses mechanical force and motion to evacuate gas from a connected chamber to create subatmospheric pressure. Valve — A device which controls fluid flow direction, pressure or flow rate. Valve actuator — The valve part(s) through which force is applied to move or position flow-directing elements. Valve, air — A valve for controlling air. Valve, cartridge — A valve with working parts contained in a cylindrical body. The cy-
lindrical body must be inserted into a housing for use. Ports through the body cooperate with ports in the containing housing. Valve, directional control — A valve whose primary function is to direct or prevent flow through selected passages. Valve, directional control, 3-way — A directional control valve whose primary function is to pressurize and exhaust a port. Valve, directional control, 4-way — A directional control valve whose primary function is to pressurize and exhaust two ports. Valve, directional control, check — A directional control valve which permits flow of fluid in only one direction. Valve, directly operated — A valve in which the controlling forces acting on the element directly influence the movement of the control elements. Valve, electrohydraulic, proportional — A valve which responds proportionally to input signals. Valve, flow control (flow metering) — A valve whose primary function is to control flow rate. Valve, flow control, bypass — A pressure compensated flow control valve which regulates the working flow diverting surplus fluid to reservoir or to a second service. Valve, flow control, deceleration — A flow control valve which gradually reduces flow rate to provide deceleration. Valve, flow control, pressure compensated — A flow control valve which controls the rate of flow independent of system pressure. Valve, flow dividing, pressure compensated — A flow dividing valve which divides the flow at a constant ratio regardless of the difference in the resistances of the branches. Valve, flow dividing — A valve which divides the flow from a single source into two or more branches. Valve, hydraulic — A valve for controlling liquid. Valve, needle — A flow control valve in which the adjustable control element is a tapered needle. Its usual purpose is the accurate control of the rate of volume of flow. Valve, pilot operated (indirect) — A valve in which a relatively small flow through an integral vent line relief (pilot) controls the movement of the main element. Valve, pilot — A valve applied to operate another valve or control. Volume change — A change in the volume of a seal as a result of immersion in a fluid expressed as a percentage of the original volume. Vulcanization — A thermo-setting reaction involving the use of heat and pressure, resulting in greatly increased strength and elasticity of rubber-like materials.
Organizations
This list, while not all-inclusive, gives information on many organizations that may prove useful sources of fluid power information. For a more complete list, visit our web site at www.fpweb.com.
American Institute of Motion Engineers (AIME), Kohrman Hall, 1201 Oliver St., Western Michigan University, Kalamazoo, MI 49008-5064 Phone: 616/387-6533; Fax: 616/387-4024 Web: http://www.aime.net; e-mail:
[email protected]
Compressed Air and Gas Institute (CAGI) 1300 Sumner Avenue Cleveland, OH 44115-2851 Phone: 216/241-7333; Fax: 216/241-0105 Web: http://www.taol.com/cagil e-mail :
[email protected]
Web: http://www.enre.umd.edu//iie.htm
American Iron and Steel Institute (AISI) 1101 17th Street NW, 13th Floor Washington, DC 20036 Phone:202/452-7100; Fax: 202/463-6573 Web: http://www.steel.org
Construction Industry Manufacturers Assn. 111 E. Wisconsin Avenue, Suite 1000 Milwaukee, WI 43202-4879 Phone: 414/272-0943; Fax: 414/272-1170 Web: http://www.cimanet.com e-mail :
[email protected]
American National Standards Institute (ANSI) 11 West 42nd Street, 13th Floor New York, NY 10036-8002 Phone: 212/642-4900; Fax: 212/398-0023 Web: http://www.ansi.org e-mail :
[email protected] American Petroleum Institute (API) 1220 L Street NW, Suite 900 Washington, DC 20005-4070 Phone: 202/682-8000; Fax: 202/682-8569 Web: http://www.api.org e-mail :
[email protected]
Equipment Manufacturers Institute (EMI) 10 S. Riverside Plaza, Suite 1220 Chicago, IL 60606-3710 Phone: 312/321-1470; Fax: 312/321-1480 Web: http://www.emi.org e-mail :
[email protected] Fluid Power Distributors Assn. PO Box 1420 Cherry Hill, NJ 08034-0054 Phone: 609/424-8998; Fax : 609/424-9248 Web: http://www.fpda.org e-mail :
[email protected]
American Society of Agricultural Engineers 2950 Niles Road St. Joseph, MI 49085-9659 Phone: 616/429-0300 Fax: 616/429-3852 Web: http://www.asae.org e-mail :
[email protected]
Fluid Power Educational Fdn. 3333 N. Mayfair Road, Suite 311 Milwaukee, WI 53222 Phone: 414/778-3364; Fax : 414/778-3361 Web: http://www.fpef.com e-mail :
[email protected]
American Society of Mechanical Engineers 345 East 47th Street New York, NY 10017-2392 Phone: 212/705-7722 or 800/843-2763 Fax: 212/705-7739 Web: http://www.asme.org e-mail :
[email protected]
Fluid Power Institute Milwaukee School of Engineering 1025 N. Broadway Street Milwaukee, WI 53202-3109 Phone: 414/277-7191; Fax : 414/277-7470 Web: http://www.msoe.edu e-mail :
[email protected]
American Society for Testing & Materials (ASTM) 100 Barr Harbor Drive West Conshohocken, PA 19428-2959 Phone: 610/832-9500 Fax: 610/832-9555 Web: http://www.astm.org e-mail :
[email protected]
Fluid Power Society (FPS) 2433 N. Mayfair Road, Suite 111 Milwaukee, WI 53226-1406 Phone: 414/257-0910; Fax : 414/257-4072 Web: http://www.ifps.org e-mail :
[email protected]
British Fluid Power Assn. (BFPA) Cheriton House Cromwell Business Park Chipping Norton Oxon OX7 5SR UNITED KINGDOM Phone: 44-1608-644114 Fax: 44-1608-643738 Web: http://www.bfpa.co.uk Canadian Fluid Power Assn. (CFPA) 6519-B Mississauga Road Mississauga, ON L5N 1A6, Canada Phone: 905/812-0978; Fax: 905/567-7191 Web: http://www.incan.com/cfpa e-mail :
[email protected]
Hydraulic Institute 9 Sylvan Way Parsippany, NJ 07054-3802 Phone: 973/267-9700; Fax : 973/267-9055 Institute of Electrical & Electronics Engineers 405 Hoes Lane Piscataway, NJ 08855 Phone: 732/981-0060; Fax : 732/981-1721 Web: http://www.ieee.org e-mail :
[email protected] Institute of Industrial Engineers 25 Technology Park/Atlanta Norcross, GA 30092 Phone: 770/449-0460; Fax : 770/263-8532
Instrument Society of America (ISA) 67 Alexander Drive, PO Box 12277 Research Triangle Park, NC 27709 Phone: 919/549-8411; Fax : 919/549-8288 Web: http://www.isa.org e-mail :
[email protected] International Standards Organization (ISO) 1, rue de Varembe, Case postale 56 CH-1211, Geneva 20, Switzerland Phone: 41-22-749-0111; Fax : 41-22-733-3430 Web: http://www.iso.ch e-mail :
[email protected] National Electrical Manufacturers Assn. (NEMA) 1300 N. 17th Street, Suite 1847 Rosslyn, VA 22209 Phone: 703/841-3200; Fax : 703/841-3300 Web: http://www.nema.org e-mail : mal_o’
[email protected] National Fluid Power Assn. (NFPA) 3333 N. Mayfair Road Milwaukee, WI 53222-3219 Phone: 414/778-3344; Fax : 414/778-3361 Web: http://www.nfpa.com e-mail :
[email protected] Power Transmission Distributors Assn. (PTDA) 6400 Shafer Court, Suite 670 Rosemont, IL 60018 Phone: 847/825-2000; Fax : 847/825-0953 Web: http://www.ptda.org e-mail :
[email protected] Society of Automotive Engineers (SAE) 400 Commonwealth Drive Warrendale, PA 15096-0001 Phone: 412/776-4841; Fax : 412/776-5760 Web: http://www.sae.org e-mail :
[email protected] Society of Manufacturing Engineers (SME) One SME Drive, PO Box 930 Dearborn, MI 48121-0930 Phone: 313/271-1500 or 800/733-4763; Fax : 313/271-2861 Web: http://www.sme.org e-mail :
[email protected] Society of Tribologists and Lubrication Engineers (STLE) 840 Busse Hwy. Park Ridge, IL 60068-2376 Phone: 847/825-5536; Fax : 847/825-1456 Web: http://www.stle.org e-mail :
[email protected] Taiwan Hydraulics & Pneumatics Assn. Rm. 3C-31 Taipei Trade Center Exhibition, Hall 5 Hsinyi Rd., Sec. 5, Taipei, TAIWAN Phone: 886-2-723-3768; Fax : 886-2-723-3813 Web: http://wwwthpa.org.tw e-mail :
[email protected]
1998/99 Fluid Power Handbook & Directory
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