Vertical Turbine Pump Information

 

Section B -- Pump Application Data 1. DATUM OR GRADE GRADE  - The elevation of the surface from which the pump is supported. 2. STATIC STATIC LIUID LE!EL LE !EL  - The vertical distance from grade to the liquid level when no liquid is being drawn from the well or source. ". DRA#DO#$ DRA#DO#$ - The distance between the static liquid level and the liquid level when pumping at required capacity. %. PUMPI$G LE!EL LE!EL  - The vertical distance from grade to liquid level when pumping at rated capacity. acity . PumpingLIUID liquid level equals static water level plus drawdown. &. SETTI$G SETTI$G - The distance from grade to the top of the pump bowl assembly assembly.. '. TPL (TOTAL PUMP LE$GT)* LE$GT)*  - The distance from grade to lowest point of pump. +. RATED PUMP )EAD )EAD  - Lift below discharge plus head above discharge plus friction losses in discharge line. This is the head for which the customer is responsible and does not include any losses within the pump. ,. COLUM$ A$D DISC)ARGE )EAD RICTIO$ LOSS LOSS  - Head loss in the pump due to friction in the column assembly and discharge head. Friction loss is measured in feet and is dependent upon column size shaft size setting and discharge head size. !alues !alues given in appropriate charts in "ata #ection. . BO#L )EAD )EAD  - Total head which the pump bowl assembly will deliver at the rated capacity. This is curve performance. 1/. BO#L EICIE$C0EICIE$C0- The efficiency of the bowl unit only. This value is read directly from the performance curve.

11. BO#L )ORSEPO#ER)ORSEPO#ER- The horsepower - required by the bowls only to deliver a specified capacity against bowl head.

 

12. TOTAL PUMP )EAD )EAD  - $ated pump head plus column and discharge head loss. %ote& This is new or final bowl head. 1". S)AT RICTIO$ LOSS - The horsepower required to turn the lineshaft in the bearings. These values are given in appropriate table in "ata #ection. 1%. PUMP BRAE )ORSEPO#ER )ORSEPO#ER - #um of 'bowl horsepower plus shaft loss (and the driver thrust bearing loss under certain conditions). 1&. TOTAL PUMP EICIE$C0 (#ATER TO #ATER* -The #ATER*  -The efficiency of the complete pump less.the driver with with all pump losses ta*en into account.

1'. O!ERALL EICIE$C0 (#IRE TO #ATER*-The #ATER* -The efficiency of the pump and motor complete. +verall efficiency , total pump efficiency  motor efficiency. 1+. SUBMERGE$CE-"istance SUBMERGE$CE-"istance from liquid level to suction bell.

Vertical Turbine Pumps Section 2 Section B -- !etical Tu3ine Pump4 Tu3ine $omenclatue !etical Tu3ine Pump4 Calculatin5 A6ial T7u4t nder normal circumstances !ertical !ertical Turbine Pumps have a thrust load acting parallel to the pump shaft. This load is due to unbalanced pressure dead weight and liquid direction change. +ptimum selection of the motor bearing and correct determination of required bowl lateral for deep d eep setting pumps require accurate *nowledge of both the magnitude and direction (usually down) of the resultant of these forces. /n addition but with a less significant role thrust influences shaft H.P. rating rating and shaft critical speeds. IMPELLER T)RUST /mpeller Thrust in the downward direction is due to the unbalanced discharge pressure across the eye area of the impeller. #ee diagram 0.

1ounteracting this load is an upward force primarily due to the change in direction of the liquid passing through the impeller. The resultant of these two forces constitutes impeller thrust. 1alculating this thrust using a thrust constant (2) will often produce only an appro3imate thrust value because a single constant cannot e3press the upthrust component which varies with capacity.

 

To accurately determine impeller thrust thrust-capacity curves based on actual tests are required. #uch curves now e3ist for the 404 Line. To determine determine thrust the thrust factor 424 is read from the thrust-capacity curve at the required capacity and given $P5. 424 is then multiplied by the Total Total Pump Head (Final Lab Head) times #pecific 6ravity of the pumped liquid. /f impeller thrust is e3cessively high the impeller can usually be hydraulically balanced. This reduces the value of 424. 7alancing is achieved by reducing the discharge pressure above the impeller eye by use of balancing holes and rings. #ee diagram 7. $OTE8  0lthough hydraulic hydraulic balancing reduces impeller impeller thrust it also decreases efficiency efficiency by 8 to 9 points by providing an addi-tional path for liquid recirculation. $OTE8  0lthough hydraulic hydraulic balancing reduces impeller impeller thrust it also decreas-es efficiency by one one to five points by providing an additional path for liquid recirculation. +f even greater concern is that should the hydraulic balancing holes become clogged (unclean fluids fluids with solid content intermittent services etc.) the impeller thrust will increase and possibly cause the driver d river to fail. Hydraulically balanced impellers cannot be used in applications requiring rubber bowl bear-ings because the flutes on the inside diameter of the bearings pro-vide an additional path to the top side of the impeller thus creating an additional down thrust. Hydraulically balanced impellers should be used as a 4last resort4 for those situations where the pump thrust e3ceeds the motor thrust bearing capabilities. Section B -- !etical Tu3ine Pump4 Tu3ine $omenclat DEAD #EIG)T /n addition to the impeller force dead weight (shaft plus impeller weight less the weight of the liquid displaced) acts downward. +n pumps with settings less than 9: feet dead weight may be neglect-ed on all but the most critical applications as it represents only a small part of the total force. +n deeper setting pumps dead weight becomes significant and must be ta*en into account. $OTE8 ;e normally only ta*e shaft weight into consideration as dead weight the reason being that impeller weight less its liquid displace-ment weight is usually a small part of the total. S)AT SLEE!ES Finally there can be an upward force across a head shaft sleeve or mechanical seal sleeve. /n the case of Finally can pumps with suction pressure there can be an additional upward force across the impeller shaft area.  0gain for most applications these forces are small and can be neglected< however when there is a danger danger of upthrusts or when there is high discharge pressure (above =:: psi) or high suction pressure (above >:: psi) these forces should be considered. MOTOR BEARI$G SI9I$G 6enerally spea*ing a motor for a normal thrust application has as standard a bearing adequate for shutoff thrust. ;hen practical motor bearings rated for shutoff conditions are preferred. For high thrust applications (when shutoff thrust e3ceeds the standard motor bearing rating) the motor bearing may be sized for the ma3imum anticipated operating range of the pump. #hould the pump operate to the left of this range for a short period of time anti-fraction bearings such as angular contact or spherical roller can handle h andle the overload. /t should be remembered however however that bearing life is appro3imately inversely proportional to the cube of the load. #hould the load double motor bearing life will be cut to 8 of its original original value. 0lthough down thrust overloading is possible the pump must never be allowed to operate in a continuous up thrust condition even for a short interval without a special motor bearing equipped to handle it. #uch upthrust will tail the motor bearing. CALCULATI$G MOTOR BEARI$G LOAD  0s previously stated stated for short setting setting non-hydraulic non-hydraulic balanced pumps below 9: feet with discharge pressures below =:: psi and can pumps with #uction pressures below 8:: psi only impeller thrust need be considered. nder these conditions&

 

5otor 7earing Load (lbs.) Timp, 2HL 3 #6 ;here& /mpeller Thrust (lbs.) 2,Thrust factors (lbs.?ft.) HL , Lab Head (ft.) #6 , #pecific 6ravity For more demanding applications the forces which should be considered are impeller thrust plus dead weight minus any sleeve or shaft area force. In e:uation ;om8  ;om8  5otor 7earing Load , Timp @ ;t(8) - sleeve force (A) -shaft area force(B) ,Tt

(8) ;t., #haft "ead ;t. 3 #etting /n Ft. (A) #leeve Force,#leeve area 3 "ischarge pressure (B) #haft 0rea 0rea Force , #haft area 3 #uction pressure +il Lube shaft does not displace liquid above the pumping water level and therefore has a greater net weight. T)RUST BEARI$G LOSS Thrust bearing loss is the loss of horsepower delivered to the pump at the thrust bearings due to thrust. /n equation form&

where& T7 L  , , Thrust loss (HP) 7HP 7ra*ebearing horsepower  Tt , 5otor 7earing Load (Lbs.) , Timp@ ;t(8) - sleeve force(A) - shaft area force(B)

 

 Key Factors to Consider Before Before App Applying lying a V Veertical Tu Turbine rbine Pump !ertical !ertical Turbine Pumps (figure 8) can be an attractive choice for many water and wastewater applications because of their  many advantages. The vertical construction ta*es up little floor space< priming problems can be avoided due to the impellers being submerged in the liquid< the first stage impeller can be lowered (by increasing the pit depth if necessary) to provide the desired %P#H margin< the multistage construction offers higher efficiencies on high head lower flow applications< and the modular construction allows the pumps to be customized for many applications.

Potential Citical I44ue4 1are must however be ta*en when applying a vertical turbine pump (!TP) in short-set applications due to a number of critical issues that can occur with this unique pump construction. !ertical !ertical Turbine pumps were originally developed for deep-water wells but the design was found to have certain benefits in industrial and municipal applications. 7ut the conditions that allow successful operation of !TP's in deep-water well applications are not always present in shorter set water and wastewater services. 8. !ertical !ertical turbine pumps do not balance the a3ial h hydraulic ydraulic pressure forces on the impeller which are substantial and with a downward direction. /nstead they rely on the thrust bearing in the motor to carry this high load. This motor thrust bearing must also carry the rotor weight which can be quite large for deep well settings. 0 0lthough lthough high a3ial thrust is normally a bad thing in most pumps in a !TP it is actually a good thing.

 

1lic* here to enlarge image

This high a3ial thrust actually compensates for alignment and resonant frequency issues that often e3ist in the standard vertical pump construction especially when applied on short set applications. The threaded line shaft couplings that are used in the standard !TP construction cannot hold accurate shaft alignment because the diameter of the line shaft ends which butt together are too small compared to the section shaft lengths. /n order to hold a reasonable concentricity (say .::9 inches) of the line shaft at the end of a typical 8: feet span each shaft end surface would have to be machined square to within a tolerance of C .::::A9 inches. This is not possible with current machining technology. technology. !ertical !ertical turbine pumps instead rely on the high a3ial thrust from the unbalanced pressure force on the impellers plus the rotor weight to *eep the shaft sections straight enough for successful operation. A. Further being so long and narrow !TP components components can be subDect to high vibration. The typical line shaft bearing spacing is 8: feet which means that most line shafts are operating above their first critical speed. Further vertical turbine pumps have more then one 4first critical speed4 depending on the shaft end condition. First there is the first 4normal4 (free-free end supports) critical speed but instead of the ne3t (normally second) critical speed being four times the first critical speed it is actually 8.9= times the first critical speed (for free-fi3ed end supports) and A.9E times the first critical speed for fi3ed-fi3ed end supports. Further if you add reasonable safety margins around these various critical speeds it becomes difficult to avoid operating very far away from a shaft critical speed especially with a variable speed drive. Here again it is the high a3ial thrust that helps typical vertical pumps in deep-well applications to achieve reasonable life spans. B. #o what happens when a !TP is applied in a short set application and operated at high flow rates ? low discharge pressures ;ell without the resulting high a3ial thrust line shaft vibration and bearing loads can be greatly increased reducing the life of the internal sleeve bearings. >. Further some or all of the internal sleeve bearings are lubricated ? wetted by the liquid pumped so they must be able to handle any corrosives and?or abrasives in the pumpage. Further the the bearings must be able to handle any air in the pumped liquid. The amount of entrained air can be substantially increased if there is any cavitation in the pump. 7ased on tests conducted by the writer for a Te Te3as 3as 0G5 Pump ser's #ymposium paper the amount of dissolved air released (to become entrained air - the damaging *ind) is greatly increased as the %P#H 5argin (%P#H0 ? %P#H$) is reduced (see figure 8). ;hile cavitation can be present in a centrifugal pump up to a %P#H 5argin $atios as high as > or more the entrained air liberated by cavitation begins to dramatically increase as the %P#H 5argin $atio approaches around 8.> near the best efficiency point (bep) flow rate and A.: around the low flow suction recirculation condition. This means that as the %P#H  0vailable from the sump (sy (system) stem) approaches the p pump ump %P#H $equired that that the internal sleeve sleeve bearings especially especially those in the bowl assembly assembly will see more air (dry operation). %ow if these bowl bearings are bronze (typical) and limited in leadincontent (which giveslife. bronze bearings their lubricity) due to regulations in the drin*ing water industry this can result very short bearing

 

9. 0 vertical vertical turbine pump can also e3perience e3cessive structural vibration of the discharge head and associated driver in the field even though the driver has been shop tested with low vibration and the rotating components were properly balanced. This could be the result of a 4reed4 (natural) frequency of the motor head and foundation assembly. Thi This s is occasionally caused by the stiffness of the field foundation being different then anticipated. 0 4reed resonance4 effect will result if the natural frequency of the assembly is at or near (within 89 of) the running frequency of the pump. !a !ariable riable speed operation drastically increases this li*elihood.

UPGRADES OR S)ORT SETTI$GS8 #o does this mean that vertical turbine pumps should not be used for short settings or variable speed operation %ot necessarily if consideration is given to the following pump up grades& necessarily 8. $eplace the standard threaded line shaft c couplings ouplings with a clamp type type (or equivalent) coupling that insures line shaft straightness. A. Provide bowl line shaft bearings with better lubricity lubricity such as those made of !esp !espel el 1$-=8::. B. $educe the bearing spacing from 8: feet to 9 feet to insure that the shaft first critical frequency is at least 89 above the ma3imum operating speed. >. 1hange from threaded to flanged column pipe connections connections for improved housing housing straightness. 9. Perform a Finite Ilement 0 0nalysis nalysis of the motor head and foundation assembly assembly.. =. se a #uction 7ell inst instead ead of a #uction 1ase to minimi minimize ze the required submergence and %P#H $equirement $equirement of the pump. J. /nsure that the %P#H 5argin is a att least 48.>4 if the liquid end (bowl (bowl assembly) has 4High #uction #uction Inergy4 (see +ctober A::J Pump Tips 1olumn).

 

 Back to Basics: How to mpro!e mpro!e V Veertical Tu Turbine rbine Pump "eliab "eliability ility t#roug# t#roug# $ptimum Bearing %election !ertical !ertical turbine pumps (!TPs) offer many unique advantages for many applications. For instance the vertical construction ta*es up little floor space< priming problems can be avoided due to submersion of the impellers in liquid< the first stage impeller can be lowered (by increasing the pit depth if necessary) to provide the desired %P#H margin< multistage construction and midrange specific speeds offer high efficiencies< and modular construction allows the pumps to be customized for many applications. !TPs are available in deep well wet pit (short setting or close-coupled) canned and submersible motor configurations.  0ccordingly my  0ccordingly my  #eptember A::E column  column on the advantages and cautions of using !TPs on water and wastewater applications concentrated primarily on cavitation vibration and a3ial thrust as well as how ho w to avoid the associated field problems.  0n additional *ey issue that s should hould be considered w when hen applying a !TP on liquids that c contain ontain solids abrasives and?or air  is the selection of the bearing material and?or construction given the fact that the bearings are immersed in and lubricated by the fluid pumped during most typical applications. 0s 0s such they are also generally the first component to deteriorate in a !TP.

Con8= stainless steel bowl shaft without any replaceable sleeves (see Fig. 8).

Further the column assembly connects the bowl assembly to the aboveground discharge head. Typical column bearings are either constructed of a cutless rubber (see Fig. A) operating against a stainless steel shaft sleeve (lubricated by the fluid pumped) or bronze enclosed in a tube (lubricated by either an oil drip or water flush introduced at the discharge head and e3iting into the well or sump at the top of the bowl assembly< see Fig. B).

 

Open Line47a;t +pen column line shaft bearing construction is recommended for ease of maintenance and?or whenever a special bearing material is required< it is not recommended for longer settings greater than about 8:: feet. $enewable shaft sleeves or hard facing on the shaft are available for longer life and typical bearing spacing is 8: feet for well applications. However However for shorter settings the shaft size and spacing should be selected so that the shafting will operate below its first critical speed (see Fig. >). For e3ample at 8E:: $P5 the ma3imum bearing spacing f or a 8 88 88?8= ?8= inch (8.=K inch) shaft would be five feet.

Enclo4e= Column /n this configuration an enclosing tube provides the lineshaft with protection from the pumped liquid and ensures clean lubrication to the bearings prior to startup which is especially important for deeper settings (over about 8:: feet). The lineshaft bearings are typically spaced at five-foot intervals to support the lineshaft. 0 0n n internal spiral groove allows the lubricant to flow between the shaft and the inner face of the bearing while the outside of the bearing is threaded to connect the enclosing tube sections.

 

This construction minimizes maintenance maintenance of the column bearings in abrasive services. The oil (or water) lubrication for the enclosed construction is introduced at the surface. 0 tan* attached to the discharge head provides oil through a solenoid valve to the tension bearing in the stuffing bo3. /t then flows by gravity into the enclosing tube and through the bypass port in the bowl assembly discharge case. 0 0lternate lternate lubricants such as clean water or grease can also be used with enclosed lineshaft construction. Lubricating Lubricating oils are available that are acceptable for discharge into the pumped liquid even when it is intended for drin*ing water.

Bo>l A44em3l? Beain54 Line shaft bearings can be protected from abrasive wear by either constructing them of cutless rubber (which can tolerate fairly high levels of suspended solids) or by the use of an enclosed tube around the lineshaft< bowl bearings however must operate in the pumped liquid. This means that in most cases the choice of bearing material (see Table Table 8) is normally the only option that will allow the pump to handle higher levels of solids or air?vapor. 0 0ir ir can enter a !TP when the well or sump levels are low (vorte3ing) can be entrained in the pumpage due to mi3ing or can be released from entrainment due cavitation in the first and?or second stage of the pump (as discussed in the #eptember A::E  A::E column). +nce in the bowl assembly assembly the air and vapors (being lighter then water) can be centrifuged into the bearings. #ome !TP manufacturers do however offer rifle-drilled bowl assembly shafts with an e3ternal water flush that can greatly improve the bowl bearing life when handling solids?abrasives and?or air?vapor as shown in Figure 8.

Beain5 Mateial Option4 Ta Table ble 8 lists the classes of column and bowl assembly bearing materials generally available for !TPs with each manufacturer typically offering their own specific alloys. The bronze alloys typically offered have very low (if any) lead due to environmental concerns which reduces their dry-running ability. 1arbon graphite bearings probably offer one of the most efficient dry-running capabilities but have very low solids?abrasive tolerance< they are also available with a variety of fillers. Teflon Teflon bearings also have e3cellent dry-running capability and poor abrasive tolerance plus they are available with a variety of fillers. However Teflon Teflon bearings cannot be retained with a press fit due to the potential of cold flow. !espel !espel bearings are also much li*e Te Teflon flon bearings but with improved dimensional stability and they are also available with a variety of fillers including Teflon and carbon. $ubber bearings are primarily used for open lineshaft column applications and are very proficient at handling solids as long as they are not too sharp. However they have poor lubricity and should be wetted prior to startup which could pose a problem with deep settings (over about 8:: feet) especially if the pump does not have a foot (chec*) valve and it ta*es too long for the pumped liquid to reach the upper bearings. $ubber bearings are seldom used for bowl bearings due to the larger required running clearance. Finally harden hardened ed surface metal bearings are also available such as chromium o3ide and tungsten carbide for abrasive?solid applications but they are e3pensive and also have poor lubricity for handling air and?or vapors.  About the Author: Author: All Allan an R. Budris, P P.E .E., ., is an independent consult consulting ing engineer who specializes in training, failure analysis, troubleshooting, reliability, efficiency efficiency audits, and litigation support on pumps and pumping systems. With offices in Washington, .!., he can be contacted "ia email at budrisconsulting#comcast.net.

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