Che 10110
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Engineering Encyclopedia Saudi Aramco DeskTop Standards
Other Heat Transfer Equipment
Note: The source of the technical material in this volume is the Professional Engineering Development Program (PEDP) of Engineering Services. Warning: The material contained in this document was developed for Saudi Aramco and is intended for the exclusive use of Saudi Aramco’s employees. Any material contained in this document which is not already in the public domain may not be copied, reproduced, sold, given, or disclosed to third parties, or otherwise used in whole, or in part, without the written permission of the Vice President, Engineering Services, Saudi Aramco.
Chapter : Process File Reference: CHE10110
For additional information on this subject, contact R. A. Al-Husseini on 874-2792
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CONTENTS
PAGES
PROCESSING STEPS IN REFRIGERATION SYSTEMS ..................................1 Determining Refrigeration Systems Requirements ....................................5 Calculating Percent of Design Duty ...........................................................9 Air-Cooled Exchangers ..............................................................................9 Fan Types .................................................................................................10 Air Control ...............................................................................................12 Ambient Air Effects..................................................................................12 CALCULATING COOLER DESIGN REQUIREMENTS..................................15 Cooler Configuration/Materials................................................................15 Tube Fin Types.........................................................................................15 Cooler Design Method .............................................................................19 Calculating Plot Area ...............................................................................30 OTHER TYPES OF HEAT EXCHANGERS ......................................................32 General .....................................................................................................32 Plate/Frame...............................................................................................34 Hairpin......................................................................................................37 Box Coolers..............................................................................................39 Enhanced Heat Transfer ...........................................................................41 CALCULATING FURNACE EFFICIENCY ......................................................45 General .....................................................................................................45 Direct Fired...............................................................................................46 Forced Draft Furnaces ..............................................................................52 Combustion Air Preheaters.......................................................................52 Efficiency .................................................................................................52
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Part 2 - Calculate Furnace Efficiency.......................................................54 Routine Furnace Startup and Operations..................................................58 Startup...........................................................................................58 Optimum Excess Air Levels .........................................................61 Monitoring Devices and Techniques........................................................65 Controls/Safety Devices/Burners..............................................................65 HEATER ALARM/SHUTDOWN SYSTEM DESCRIPTION............................67 Monitoring Tube Metal Temperature .......................................................71 NOMENCLATURE.............................................................................................73 KEY FORMULAS...............................................................................................75 WORK AID 1 - PROCEDURES FOR CALCULATING PERCENT OF DESIGN DUTY.........................................................................77 WORK AID 2 - PROCEDURES FOR CALCULATING EXTENDED SURFACE AND FACE AREA REQUIREMENTS..................79 WORK AID 3 - PROCEDURES FOR CALCULATING FURNACE EFFICIENCY.............................................................................82 GLOSSARY ........................................................................................................84 REFERENCES.....................................................................................................87 APPENDICES .....................................................................................................88
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PROCESSING STEPS IN REFRIGERATION SYSTEMS Refrigeration systems are used to lower the operating temperature of certain processing schemes to temperatures that cannot be obtained via air and/or water cooling. The lower process operating temperatures are usually required because of: •
The process fluid's thermodynamic properties.
•
Fouling tendencies at higher temperatures.
•
The economic attractiveness of operating the process at a lower temperature.
•
The system refrigerant is usually selected based on:
•
The refrigerator temperatures needed.
•
The type of refrigerants available from the process plant.
•
A need to not contaminate the process plant fluid if equipment were to leak.
Some common refrigerants and their properties are listed below.
ASHRAE Refrigerant Number
Chemical Name
11
Trichlorofluoromethane
114
Dichlorotetrafluoroethane
Chemical Formula
CC13F CC1F2OC1F
Molecula r Weight
137.4
Normal Boiling Point °F @ 14.696 psia
74.8
Critica l Temp. °F 388.4
Liquid Viscosity Centipoise
Liquid Thermal Conductivity Btu (hr ft2 °F) ft
-168
0.421@NBT 0.395@86°F
0.0506@NBT 0.0498@86°F
1.13
5
0.0405@NBT 0.0366@86°F
1.09
6
Critical Pressure psia
Freezing Point °F @ 14.696 psia
640.0
Specific Heat Ratio k = Cp/Cv
Toxicity UL Group Classification
170.0
38.4
294.3
474.0
-137
0.44@NBT 0.32@86°F
120.9
-21.6
233.6
597.0
-252
0.358@NBT 0.206@86°F
0.0518@NBT 0.0392@86°F
1.14
6
-256
0.33@NBT 0.192@86°F
0.0695@NBT 0.0495@86°F
1.18
5a
-217
0.213@NBT 0.159@86°F
0.0663@NBT 0.061@86°F
1.09
5b
-305
0.21@NBT 0.101@86°F
0.076@NBT 0.056@86°F
1.14
5b
-301
0.15@NBT 0.089@86°F
0.082@NBT 0.057@86°F
1.15
5b
-297
0.168@NBT 0.039@86°F
0.082@NBT 0.048@86°F
1.19
5b
0.111@NBT 0.031@86°F
2 12
Dichlorodifiuoro Methane
22
Chlorodifiuoro Methane
600 290 1270 170
N-Butane Propane Propylene Ethane
CC12F2 CHC1F2 C4H10 C3H 8 C3H 6 C2H 6
86.5 58.1 44.1 42.1 30.1
-41.4 31.1 -43.7 -53.9 -127.4
204.8 305.6 206.0 197.1
716.0 550.7 616.3 667.2
9.01 707.8
1150
Ethylene
C2H 4
28.1
-154.8
48.6
731.1
-272
0.17@NBT 0.07@86°F
1.24
5b
50
Methane
CH4
16.0
-258.7
-116.7
667.8
-296
0.118@NBT
0.110@NBT
1.305
5b
-108
0.25@5°F 0.207@86°F
0.29@32°F 0.26@86°F
1.29
2
717
Ammonia
NH3
17.0
-28.0
270.4
1636.0
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. Saudi Aramco DeskTop Standards
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Table 1
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A very simplified process flow diagram showing how a refrigeration system works is shown in Figure 1. The refrigerant is compressed to a pressure that permits condensing the compressor discharge with water or air coolers. The condensed liquid is then fed to the process cooler (chiller or evaporator). An expansion valve on the refrigerant feed to the process cooler significantly lowers the pressure of the refrigerant. The depressured refrigerant in the cooler boils at a low temperature, thereby cooling the process fluid on the tube side of the cooler. Low-pressure vapors from the process cooler shell flow back to the compressor. The refrigeration cycle just discussed (from the compressor through the water or air cooler, expansion valve and process fluid cooler, back to the compressor) is also shown thermodynamically on the pressure/enthalpy diagram in Figure 1. The A, B, C points on the pressure/enthalpy diagram correspond to the letters on the process flow diagram in Figure 1. Figure 2 is a pressure/enthalpy diagram that would be used by an engineer to review an operating refrigeration system or design a new system.
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PROCESS FLOW DIAGRAM Flow Diagram
PRESSURE ENTHALPY DIAGRAM
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 1
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Determining Refrigeration Systems Requirements Engineers will often review pressure/enthalpy diagram when reviewing operating refrigeration systems or designing new systems. The various calculations involved with the four processing steps discussed earlier can also be determined using these diagrams. Figure 2 is an example of a diagram from which all four refrigeration processing steps, compression, condensation, refrigerant flashing, and vaporization may be determined. Figure 2, a Propane P-H diagram, is used for refrigerant systems calculations in which propane suitable for refrigeration is manufactured during a process. The primary outcome of these calculations is to determine the quantity of refrigerant needed to provide the appropriate process cooling duty. The enthalpy of the condensed compressor discharge can be read from Figure 2 at the intersection of the bubble point curve and the known condensing temperature of water in the system. After the appropriate compressor suction pressure is selected (this pressure will boil the propane at a low enough temperature to ensure the desired temperature difference across the tubes in a process cooler), the enthalpy of the vapor returning to the compressor may be read at the intersection of the psia pressure line and the dew point line. The enthalpy of the liquid refrigerant (in this case, propane) to the process cooler is read at the intersection of the psia line and the bubble point curve. Note that the lb/hr of refrigerant vaporized in the process cooler is determined using the following formula: Process Duty Refrigerant Heat of Vaporization After the lb/hr of refrigerant vaporized is determined, the lb/hr of refrigerant flashed across the expansion valve must be calculated. The refrigerated flash is an isenthalpic reaction and is calculated using the process cooler refrigerant rate and the enthalpies of the feed and products to and from the expansion valve. It is important to remember that the total pounds of refrigerant from the valve is equal to the total pounds to the valve. The calculation for refrigerant flashed lb/hr can be summarized as: (Enthalpy of Compressor Discharge Condensed Liquid, Btu/lb) (Compressor, Gas Rate, lb/hr) = (Vapor from Pressure Reduction Valve, lb/hr) (Vapor Enthalpy, Btu/lb) + (Liquid from Pressure Reduction Valve, lb/hr) (Liquid Enthalpy, Btu/lb)
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PRESSURE/ENTHALPY DIAGRAM
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 2
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This calculation will yield the lb/hr of refrigerant needed to provide process duty at the specified temperature. Thermodynamic tables for refrigerants are available and can be used in place of the P-H diagram (Figure 2) for enthalpy data. When selecting the appropriate refrigerant, the following factors should be considered: • • • •
Atmospheric boiling point. Availability. Fire hazard possibilities. Toxicity.
For additional information, Figure 3 illustrates how a simple system can be expanded into a complex one with multiple stages and refrigerants. Figure 4 shows one of several types of curves available for use by engineers in the design and evaluation of refrigeration systems. In summary, the four processing steps in a refrigeration system can be calculated using a P-H diagram. CASCADE REFRIGERATION SYSTEM
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 3
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SINGLE-STAGE PROPANE REFRIGERATION SYSTEM
(use photostat)
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 4
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Calculating Percent of Design Duty Calculating the percent of design duty a cooler can perform during normal or abnormal operations involves a number of calculations. The next three sections, fan types, air control, and ambient air effects will provide sufficient information needed for calculations of percent of design duty. Air-Cooled Exchangers The extent to which air-cooled exchangers are used is one of the first major economic evaluations made in connection with a process system design. Generally, air-cooled exchangers are more attractive for the higher temperature heat removal services, and watercooled exchangers are more attractive for the lower temperature heat removal services. An economic evaluation will establish the break point temperature for the change from air-cooled to water-cooled exchangers. The availability of plot space for air coolers and the availability of makeup water for the cooling water system also significantly influence the air versus water evaluation, particularly for unusual circumstances. The design alternatives available for specifying an air-cooled exchanger will now be discussed. Initially, an engineer should refer to Saudi Aramco documents AES-E-001 and ADP-E-001 for the preferred air-cooled exchanger configurations used by Saudi Aramco. Whenever possible, the final design should conform to these preferred configurations. See Figure 5 for a typical air-cooled installation. TYPICAL SIDE ELEVATIONS OF AIR COOLERS
Figure 5
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TYPICAL PLAN VIEWS OF AIR COOLERS
Figure 5 (Cont'd) Fan Types Air-cooled exchangers can be specified with induced draft or forced draft fans. If the fan is located below the tube bundle and the fan discharges upward, the cooler has a forced draft fan. If the fan is above the tube bundle and discharges upward, the cooler has an induced fan. The advantages and disadvantages of each are listed below. The advantages of induced draft fans are: •
Better distribution of air across the section.
•
Less possibility of the hot effluent air recirculating around to the intake of the sections. The hot air is discharged upward at approximately 2-1/2 times the velocity of intake, or about 1500 ft/min.
•
Less effect of sun, rain, and hail, since 60% of the face area of the sections is covered.
•
Increased capacity in the event of fan failure, since the natural draft stack effect is much greater with induced draft.
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•
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
The disadvantages of induced draft fans are: •
Higher horsepower, since the fan is located in the hot air. (Horsepower varies directly with the absolute temperature.)
•
Effluent air temperature should be limited to 200°F to prevent potential damage to fan blades, bearings, V-belts, or other mechanical components in the hot air stream.
•
The fan drive components are less accessible for maintenance, which may have to be done in the hot air generated by natural convection.
•
For inlet process fluids above 350°F, forced draft design should be used; otherwise, fan failure could subject the fan blades and bearings to excessive temperatures.
The advantages of forced draft fans are: • • • •
Slightly lower horsepower because the fan is in cold air. Better accessibility of mechanical components for maintenance. Easily adaptable for warm air recirculation for cold climates. Capable in cooling process fluids with higher inlet temperatures.
The disadvantages of forced draft fans are: • • • • •
Poor distribution of air over the section. Greatly increased possibility of hot air recirculation, due to low discharge velocity from the sections and absence of stack. Low natural draft capability on fan failure, due to small stack effect. Total exposure of tubes to sun, rain, and hail.
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
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Air Control It is often attractive to regulate the air flow to the airfin cooler and/or condenser because seasonal variations in air temperature can result in considerable overcooling of the process stream. Overcooling can cause exchanger plugging or other processing problems. The air can be regulated by variable speed drives for the fan or variable pitch blades on the fans. Louvers on the face of the tube bundle will also control the flow of air. Louvers will not reduce fan motor electrical consumption, whereas the variable fan pitch and variable speed drive will. Since the air-cooled exchanger must be specified for the hottest time of the year (typically 94 to 96% of the maximum expected air temperature, +5°F to allow for air recirculation), for most of the year there is an opportunity to reduce electrical costs with variable pitch blades or a variable speed motor. Designing the air cooler at 94 to 96% of the maximum temperature versus the maximum air temperature often reduces the cost of the air cooler by 50 to 60%. Refer to Saudi Aramco AES-E-002 (4-14-3) for Saudi Aramco design air temperature. Appendix A is a sample of the input to and output from a Saudi Aramco computer program from an airfin exchanger. Ambient Air Effects Changes in air temperature significantly affect an operating air cooler's process cooling capability and the horsepower requirement for the fans. Example Problem 1 shows that when the air temperature is at the minimum design level, about half the airfin cooling surface on this service could be shut down to compensate for the Æte increasing from 75 to 164°F. Actually, the partial shutdown mode is accomplished by shutting down some of the airfin fans on a large duty air cooler that has three or more fans during the winter, when air temperatures are consistently below design. Shutting down fans or specifying variable pitch fan blades or drivers for the fans can often result in 20 to 80% of rated horsepower savings, annually, due to seasonal changes in the air temperature. The effect of changes in ambient air temperature on cooling capacity is illustrated by Example Problem 1. Example Problem 1 In order to calculate the percent of design duty, Æte must first be calculated. Use Part 1 of Work Aid 1 to calculate Æte for the design air temperature operation and the minimum air temperature application, using FT factors based on curves in Figure 6. Assume that the air cooler is cooling a product stream from 280°F to 170°F, the exchanger duty is 80 MBtu/hr, the design air temperature is 120°F, the minimum air temperature is 38°F (night time in the winter), and the air temperature rise across the cooler is 45°F when inlet air to the air cooler is at the design value (120°F). Also assume that the air cooler has four tube rows.
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EXAMPLE PROBLEM 1 Assume that the air cooler is cooling a product stream from 280°F to 170°F, the exchanger duty is 80 MBtu/hr, the design air temperature is 120°F, the minimum air temperature is 38°F (night time in the winter), and the air temperature rise across the cooler is 45°F when inlet air to the air cooler is at the design value (120°F). Also assume that the air cooler has four tube rows. The calculated Dte for the design air temperature operation and the minimum air temperature application is as follows, using FT factors based on curves in Figure 6. X = T1 - T2 = 280 - 170 = 110 = 0.688 T1 - t1 280 - 120 160 Y = t2 - t1 = 165 - 120 = 45 = 0.409 T1 - T2 280 - 170 110 F T = 0.96
(Using the multiple-tube-pass curve of Figure 6.)
Using the LMTD chart from the TEMA Manual, Pg. 111. GTTD LTTD LMTD Design Æte
= 280 - 165 = 115 = 170-120 = 50 = 78°F = (0.96) (78) = 75°F
For the minimum air temperature operation, FT = 1.0, because the difference between the air outlet temperature and the process outlet temperature will drastically increase. The LMTD for the minimum air temperature operation, assuming no cooling of process liquid below 170°F product temperature and same air rate (same air temperature rise) is:
GTTD
LTTD
= 280 - (38 + 45) = 197
= 170 - (38) = 132
LMTD = 162°F = Æte Æte
x =
280 − 170 = 0.455 280 − 38
y =
83 − 38 = 0.40 280 − 170
FT = 0.99
= .99 (162) = 161
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CORRECTION FACTORS FOR LIMITED CROSS FLOW EXCHANGERS
(use photostat)
Basis: where: t1 t2 T1 T2
= = = =
X = T1 - T2 T1 - t1
Y = t2 - t1 T1 - T2
Ambient Air, °F Average Hot Air, °F Tubeside Inlet, °F Tubeside Outlet, °F
Corrected LMTD = (FT) (tm) = Æte Figure 6
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CALCULATING COOLER DESIGN REQUIREMENTS Cooler Configuration/Materials Air cooler configuration is greatly influenced by the type of exchanger that Saudi Aramco uses as a standard design (see ADP-E-001). The standard size is 1-in. O.D. tubes, and 1 1/2in. O.D. tubes are used for high viscosity fluids. The standard tube lengths are 12-, 20-, 24-, 30-, 32-, 36-, or 40-ft long. Standard fins are aluminum and have a height in the 3/8- to 5/8in. range (10 fins per linear inch of tube), depending on the quality of the fluid being cooled (hi value). The higher the tubeside coefficient (hi) is, the more the optimum height of the fins increases. Saudi Aramco standardizes on exchangers that have tube rows in the range of 3 to 8 deep. As the temperature difference between the process fluid inlet or air inlet temperatures increases or the overall coefficient (Uo) decreases, the optimum number of tube rows increases (see Figure 7). Figure 7 should be used only for the exchanger configuration noted on the figure. This type of exchanger is one of the standard Saudi Aramco configurations. The optimum number of tube rows indicated in Figure 7, as well as the allowable tubeside pressure drop, influences the number of tube passes selected. For tube pitch, Saudi Aramco has standardized on a 2 3/8-in. triangular spacing. This standard tube pitch should always be used, except for unusual services. An unusual service would be a circumstance where the 2 3/8-in. triangular pitch, in combination with the specified number of tube rows, face area, and air rate, results in a static air pressure requirement too high for the fans. Air fan static head availability covers only a narrow range. The maximum permissible value for static head is about 0.7 in. H2O with a normal value being closer to 0.5 in. H2O. The acceptable static head range corresponds to a face velocity range of 6.5 to 13 ft/s, depending on the temperature of the air and the number of tube rows in the exchanger. The preferred way of correcting an air pressure drop problem would be to change the face area of the air-cooled exchanger or the number of tube rows before changing the standardized tube pitch. Air-cooled exchanger materials for the tubes and headers (parts in contact with the process fluid) are specified in the same manner as shell and tube exchangers. A tubewall and header wall thickness is selected so that at the end of the desired equipment life (say 10 years or 15 years) the corroded wall will still be thick enough to contain the process fluid pressure. Tube Fin Types Saudi Aramco has standardized on two types of air cooler tube fins: •
5/8-in. high fins, 10 fins/inch of tube, extruded fin type.
•
5/8-in. high fins, 11 fins/inch of tube, embedded fin type.
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Both types are made of aluminum. Other types of fins available are wrapped fins that fit into spiral grooves cut into the host tube or fins first wrapped around the tube. The fins may also be serrated or plain. Some of the different types of tube fins are shown in Figure 8.
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EFFECT OF TEMPERATURE LEVEL AND OVERALL HEAT TRANSFER RATE UPON OPTIMUM BUNDLE DEPTH
(use photostat)
Basis:
1-in. O.D. x 24-in. long steel tube with extruded aluminum fins and 2 3/8-in. triangular spacing.
where:
T1 t1
= =
Inlet temperature of fluid to be cooled, °F. Inlet air temperature.
Uo
=
Overall heat transfer coefficient (related to bare tube outside diameter) Btu/hr ft 2 °F.
Figure 7
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TYPES OF FINNED TUBES USED IN AIR-COOLED HEAT EXCHANGERS
Figure 8
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Cooler Design Method Saudi Aramco Design Practice ADP-E-001 contains a simplified design calculation procedure that be used to check vendors' proposals for air-cooled exchangers or review the performance of an operating exchanger. The Saudi Aramco design practice procedure will be used for Example Problem 2 below, which will show how to review a vendor-specified air-cooled exchanger, checking the adequacy of the vendor-specified surface, face area, and fan horsepower. Example Problem 2 Oftentimes vendors will specify air-cooled exchangers. In order to determine if the vendor specified exchanger meets Saudi Aramco design expectations, a series of simplified design calculations can be used. These calculations can also be used to review the performance of operating exchangers. The following example problem will demonstrate how to review a vendor-specified air-cooled exchanger, checking the adequacy of the vendor-specified surface, face area, and fan horsepower. It is necessary to check the following vendor-proposed air-cooled exchanger. Calculate the air temperature rise, air rate, surface (extended) requirement, face area, number of fans, and horsepower for each fan to confirm the vendor's design. Refer to Work Aid 2 for procedures to calculate extended surface and face area requirements. Assume the following: Exchanger duty Q = 31 Mbtu/hr. Process fluid temperature in/out = 250/150°F. Process fluid flow rate Wt is 550,000 lb/hr of hydrocarbon with a specific heat Cp = 0.56 Btu/lb°F. Thermal conductivity k = 0.076 Btu/hr ft2°F/ft. Viscosity µ = 0.5 cP at the average temperature of 200°F. The tubeside fouling factor selected by the vendor is rdi = 0.0015 hr ft2°F/Btu. The tubeside configuration is 1-in. O.D. (0.87-in. I.D.), 30 ft long, 5 rows deep, on a 2 3/8-in. triangular tube pitch with 5/8-in. fins at 10 fins per inch.
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The vendor's design is also based on an overall heat transfer coefficient of 3.73 Btu/hr °F ft2 of extended surface, a 40°F rise in air temperature, a design air temperature of 120°F, and an air rate of 3.2 M lb/hr. The vendor specified six 11-ft-diameter fans with a 30 hp electric motor driver on each fan. The fan specified static head is 0.5 inH2O. The vendor's exchanger face and extended surface areas are 1200 and 160,800 ft2. TYPICAL DESIGN DETAILS FOR AIRFIN COOLERS Basis:
1-in. O.D. tubes 0.262 ft2/ft of bare tube Rs = ax/Ao 2 3/8-in. triangular spacing on tubing 5/8-in. fins at 10 fins per inch tube with a base-to-fin surface ratio of 21.2 = Rs
Depth in tube rows
3
4
5
6
7
8
Unit weight, lb/ft2 face area
66
75
80
88
97
115
Typical face velocity, ft/min
630
595
565
540
510
490
Ao/af, ft2 surface area/ft2 face area
3.80
5.04
6.32
7.60
8.84 10.0 8
Note: Table 2 is from Saudi Aramco ADP-E-001 Table 1, Pg. 103.
Table 2 Calculate the air temperature rise, air rate, surface (extended) requirements, face area, number of fans, and horsepower for each fan to confirm the vendor's design. Since the tubeside film coefficient for an air-cooled exchanger is calculated in the same manner as was covered in the shell and tube section of this course (ChE 101.09), assume that the tubeside film coefficient has been calculated and hi = 250 (Btu/hr °F ft2 of inside tube area). Confirm duty Q by using tubeside and shellside air cooler process conditions (refer to Figure 29 at the end of the text for nomenclature). For the configuration given, and from Table 2 (from ADP-E-001, Pg. 103), the bundle weight per square foot of face area is 80 lb/ft2, the recommended face air velocity (Vf) is 565 ft/min, the ratio of fin (ax) to base (Ao) surface is 21.2, and the ratio of bare tube surface area (Ao) to face area (af) is 6.32.
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Example Problem 2 (Cont'd) Tubeside: Q
= (Wt) (Cp) (T1 - T2) = (550,000) (0.56) (250 - 150) = 30.8 MBtu/hr
Shellside: Q
= (Ws) (Cp) (t2 - t1) = (3,200,000) (0.242) (160 - 120) = 30.9 MBtu/hr
Cp for air from Maxwell, Pg. 88 at average temperature of 140°F. Calculate Æte for the vendor's design, referring to Figure 6.
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X = T1 - T2 = 250 - 150 = 100 = 0.77 T1 - t1 250 - 120 130 Y = t2 - t1 = 160 - 120 = 40 = 0.4 T1 - T2 250 - 150 100 FT = 0.94 LMTD = GTTD - LTTD= 90 - 30 = 60 = 54.6 1.099 ln GTTD ln 90 LTTD 30 LMTD = 55 ² te = (FT) (LMTD) = (0.94) (55) = 52°F
T 250 1
150
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t2
160
t1 120
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Vendor's air cooler face area is 1200 ft2 Face velocity Vf is then:
Density of air @ 70°, 14.7 psia = PAIR =
29 60 + 460 0.07507 379 70 + 460 =
3.2 x 106 lb hr 1 ft 3 = 710, 415 ft 3 / min @ 70° F x x hr 60 min .07507 lb Air Rate @ 70°F = Air Rate @ 70° F and 14.7 psia af Vf = = Velocity over face area 710, 415 = 592 ft / min Vf = 1200 Note: See Figure 29 on page 69 at the end of the text for nomenclature. Example Problem 2 (Cont'd) 592 ft/min = 9.9 ft/s, which is in the middle of the acceptable range of 6 to 13 ft/s (from text) for fan discharge velocity. Saudi Aramco typical design detail guideline chart, ADP-E-001, Pg. 103 (see Participant Module, Table 2), recommends a Vf of 565 ft/min for a 5-row-deep airfin cooler.
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TABLE 2 TYPICAL DESIGN DETAILS FOR AIRFIN COOLERS Basis:
1-in. O.D. tubes 0.262 ft2/ft of bare tube Rs = ax/Ao 2 3/8-in. triangular spacing on tubing 5/8-in. fins at 10 fins per inch tube with a base-to-fin surface ratio of 21.2 = Rs
Depth in tube rows
3
4
5
6
7
8
Unit weight, lb/ft2 face area
66
75
80
88
97
115
Typical face velocity, ft/min
630
595
565
540
510
490
Ao/af, ft2 surface area/ft2 face area
3.80
5.04
6.32
7.60
8.84 10.0 8
Note: Table 2 is from Saudi Aramco ADP-E-001 Table 1, Pg. 103.
The next step in reviewing the vendor's design is confirming the overall coefficient used by the vendor. Tubeside: hi = 250 Btu/hr °F ft2 (given) Shellside: (See Figure 29.)
Vf Vf = Fraction of free flow area between tubes (P − d R ) P Vmax 592 592 = ( 2 3 / 8 -1) = .579 = 1023 ft / min 2 3/ 8 =
=
ho =
(1. 9)(d R )(V max )0.56 0.5
(P )[(N f )do ]
(1. 9)(1023)0.56 (2.375) (22. 50)
0.5
=
0.56
=
(1.9) (1)(V max )
0.5
(2 3 / 8)[(10 )(1 + 5 / 8 + 5 / 8)]
(1.9)( 48. 5) = 8.19 Btu / hr •F ft2 ( 2. 375)( 4. 74)
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Since • Rs = ax/Ao and dr/di = Ao/ai then
(Rs )(d R ) ( 21. 2) (1. 0) ax = = = 24.37 di 0.87 ai 1 = 1 ax + rdi ax + rm + 1 ai Ux hi ai ho
(From Table 2, RS = 21.2.)
Ignore rm, since the value is very small. 1 1 Ê 1 ˆ = (24. 37) + (0. 0015)(24.37) + Á ˜ U x 250 Ë8.19 ¯ 1 = 0.097 + 0.0366+ 0.12 = 0. 25 Ux
Ux = 3.94 Btu/hr °F ft2 (extended surface)
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The calculated value for Ux (heat transfer coefficient based on square feet of extended surface) is 3.94 Btu/hr °F ft2 as compared to the vendor-quoted value of Ux = 3.73 Btu/hr °F ft2. Therefore, the specified extended surface 160,800 ft2 is probably adequate. However, the equation Q = (Ux) (ax) (Æte) should be utilized to confirm the adequacy of the vendor's specified extended surface: Q = (Ux)(ax)( ∆ te) 31,000,000 = (3.94)(ax)(52) a=
31, 000, 000 = 151, 308 ft 2 (extended surface) (3.94 )(52 )
This calculation shows that 151,308 ft2 of extended surface is needed based on the calculated Ux [3.94 Btu/hr °F ft2 (extended)]. The vendor has specified 160,800ft2. This conclusively shows that the vendor's specified surface is adequate. The calculated face velocity Vf of 580 ft/min (9.7 ft/s) approximates the Saudi Aramco face velocity recommendation for this type of air cooler, which is about 565 ft/min. Also, the 9.7 ft/s calculated velocity falls within the acceptable Vf range of 6 to 13 ft/s. Therefore, it can be concluded that the vendor quoted face area of 1200 ft2 is acceptable. When reviewing a vendor's design, both the tubeside and air side pressure drops must be calculated and compared to the vendor's quoted value. Due to class time limitations, the pressure drop analysis on the tubeside will not be done as part of this example problem. The tubeside pressure drop calculation for air coolers is done in the same manner that was presented for the shell and tube exchangers. Next, the fan facilities (number of fans, horsepower, and head for fans) should be reviewed. In order to have proper air distribution over the face of the airfin exchanger, the total face area of the fans should be 40-60% of the airfin face area. The vendor specified 6 fans, with an 11-ft diameter for each. Calculate fan face area as a fraction of exchanger face area: π (Fan Diam )2 π (11)2 x No. Fans x6 Fan Face 4 4 = = = 0.48 Exchanger Face af 1200
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The ratio of fan face area to exchanger face area is 0.48, which is suitable since the acceptable range is 0.4 to 0.6. (Tube side pressure drop; the same calculation as for shell and tube exchangers.) Next, calculate the air side pressure drop and compare the calculated value with the fan static head that the vendor has specified. Fan velocity (face)Vo =
3,200,000 Ws = (60) (fan area) ρ( air ) (60) (570) (0.07507)
Vo = 1248 ft/min PT = Pst + Pv
(See Figure 9.)
Pst = (number of tube rows) (²P)
From Figure 9, Pv = 0.097. conversion of air velocity to velocity pressure Vo = air velocity for 70°F air and 29.92 inHg pressure, ft per min Pv = velocity pressure, in H2O Vo
Pv
Vo
Pv
Vo
Pv
100 200 300 400 500 600 700 800 900 1,000
0.001 0.002 0.006 0.010 0.016 0.022 0.031 0.040 0.050 0.062
1,100 1,200 1,300 1,400 1,500 1,600 1,700 1,800 1,900 2,000
0.075 0.090 0.105 0.122 0.140 0.160 0.180 0.202 0.225 0.249
2,100 2,200 2,300 2,400 2,500 2,600 2,700 2,800 2,900 3,000
0.274 0.302 0.328 0.359 0.390 0.421 0.454 0.489 0.524 0.561
Figure 9
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[(ÆP) (De)] is plotted as a function of k and Vmax (See Figure 10). k= where:
(N f )(d o ) = (10 )(1 + 58 + 5 8) = 10 (2.25) = 22.5 1 (1)0.2 (dR )0.2 Nf = Number of fins/in. tube length. do = Fin outside diameter, in. dR = Inside fin diameter, in. (Root diameter) 2
2 P d 2 3 / 8 (10)(1 + 5 / 8 + 5 / 8) De = (Nf ) o = 12 1 12 dR
= (5.64)(1.88) = 10.6 where: P = Tube pitch, in. From Figure 10: (ÆP) (De) = 0.85 at Vmax = 1023 ft/min
and k = 22.5
²P = 0.85 = 0.08 inH2O 10.6 Pst = (number of tube rows) (²P) = (5) (0.08) = 0.4 inH2O PT = Pst + Pv = 0.4 + 0.097 = 0.5 inH2O
As discussed, the most likely head available from an air fan is 0.5 in H2O. Therefore, the calculated value of 0.5 in H2O is acceptable.
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PRESSURE DROP CORRELATION FOR AIR
(use photostat)
Figure 10
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Calculating Plot Area The plot area required for a horizontal air-cooled exchanger is the exchanger face area plus a suitable buffer area between pieces of equipment on all sides of the air-cooled exchanger, to prevent the flow of hot air from other equipment into the exchanger air supply. This buffer area will vary, depending on the type and temperature level of the equipment upwind from the air-cooled exchanger. Plot area can be reduced by using an A-frame type of construction. The plot area for an Aframe configuration is about half that required for a horizontal configuration. However, when an A-frame configuration is used to meet plot area constraints, extra attention must be given to the possible flow of warm air into the air supply of the A-frame exchanger. A-frame exchangers are more susceptible than horizontal exchangers to the recirculation of their own hot air discharge into the air supply. Air-cooled exchangers are also very susceptible to damage from fire and maintenance equipment because of the large area of extended surface. Therefore, the refinery space guide should be consulted for the proper buffer area and the recommended limitation on what equipment can be under or over an air-cooled exchanger. A sample equipment spacing guide is shown in Figure 11. To use this guide, enter the chart from the equipment being located, read down to the row concerning the pieces of equipment nearby, and read the minimum distance required between them.
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ONSITE SPACING CHART
Figure 11
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OTHER TYPES OF HEAT EXCHANGERS General A number of secondary uses for heat-transfer equipment have not been discussed in the preceding modules of this course. See Figure 12 for a typical listing. Some of the more frequently utilized exchanger types will be discussed in this section of ChE 101.10. Note that not all types of exchangers will be discussed, rather, only the more common ones.
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OTHER HEAT EXCHANGER TYPES Type
Major Characteristics
Application
Double Pipe
Pipe within a pipe; inner pipe may be finned or plain.
For small units.
Extended Surface
Externally finned tube.
Services where the outside tube resistance is appreciably greater than the inside resistance. Also used in debottlenecking existing units.
Brazed Plate Fin
Series of plates separated by corrugated fins.
Cryogenic services; all fluids must be clean.
Spiral Wound
Spirally wound tube coils within a shell.
Cryogenic services; fluids must be clean.
Scraped Surface
Pipe within a pipe, with rotating blades scraping the inside wall of the inner pipe.
Crystallization cooling applications.
Bayonet Tube
Tube element consists of an outer and inner tube.
Useful for high temperature difference between shell and tube fluids.
Falling Film Coolers
Vertical units using a thin film of water in tubes.
Special cooling applications.
Worm Coolers (Box Coolers)
Pipe coils submerged in a box of water.
Emergency cooling.
Barometric Condenser Direct contact of water and vapor.
Where mutual solubilities of water and process fluid permit.
Cascade Coolers
Cooling water flows over series of tubes.
Special cooling applications for very corrosive process fluids.
Impervious Graphite
Constructed of graphite for corrosion protection.
Used in very highly corrosive heat exchange services.
Plate and Frame Heat Exchanger
Series of parallel corrugated plates separated by flexible gaskets. Process fluids flow in nearly true counterflow compact size.
Where high heat transfer effectiveness is required.
Spiral Heat Exchanger Two long parallel plates wound in spiral shape. Process fluids can be designed for counterflow or cross flow.
Where high fouling process fluids or slurries are present. Also used as lower mounted condensers or Thermosiphon Reboilers.
Figure 12
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Plate/Frame A plate and frame heat exchanger (PFHE) consists of multiple grooved plates compressed together by bolts. The liquid between the plates is contained by a gasket material compressed between the plates. At each end of the exchanger is a header plate containing the inlet and outlet parts. Plates can be added or removed as required for a service. See Figure 13 for an exploded view of a plate and frame exchanger. The PFHE has the following advantages and disadvantages compared to a conventional shell and tube heat exchangers. Advantages: •
It can be disassembled for cleaning.
•
The plates can be rearranged, added to or removed from the plate rack for different service conditions.
•
The fluid residence time is short (low ratio of fluid volume to surface area).
•
No hot or cold spots exist to damage temperature sensitive fluids.
•
Fluid leakage between streams cannot occur unless plate material fails.
•
Fluid leakage due to a defective or damaged gasket is external and easily detected.
•
Low fouling due to the high turbulence created by the plates.
•
A very small plot area required relative to a shell and tube type heat exchanger for the same service.
•
The maintenance service area required is within the frame size of the exchanger.
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Disadvantages: •
Care must be taken by maintenance personnel to prevent damage to the gaskets during disassembly, cleaning, and reassembly.
•
A relatively low upper design temperature limitation exists (usually set by gasket material at 300-400°F range).
•
A relatively low upper design pressure limitation exists (for most units, this is in the 140-230 psig range).
Gasket materials are not compatible with all fluids. With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. The plate and frame exchanger is generally considered to be a high-heat-transfer, highpressure-drop device. Of course, the plates can be arranged to give a low pressure drop, but then the heat transfer coefficient also decreases. In a typical high-pressure-drop service, the fluid flow through the exchanger is very turbulent and minimizes potential fouling. When alloy materials are required for a heat exchange service, the plate and frame exchanger is competitive with the shell and tube exchanger for conventional services.
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PLATE AND FRAME HEAT EXCHANGER Carrying Bar Plate Pack Fixed End Cover
Moveable End Cover
Carrying Bar
Compression Bolt
Courtesy Alfa-Laval Figure 13
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Hairpin Another type of exchanger that is used for small duty services is the hairpin heat exchanger, which is designed in a hairpin shape and fabricated in two types: double pipe and multitube. Figure 14 shows hairpin double pipe and multitube exchangers. The double pipe exchanger is manufactured with a bare or longitudinal finned tube. The advantages and disadvantages of a hairpin exchanger are as follows: Advantages: •
The use of longitudinal finned tubes will result in a compact heat exchanger for shellside fluids having a low heat transfer coefficient.
•
Countercurrent flow will result in lower surface area requirements for services having a temperature cross.
•
Potential need for expansion joint is eliminated due to U-tube construction.
•
Shortened delivery times can result from the use of stock components that can be assembled into standard sections.
•
Modular design allows for the addition of sections or the rearrangement of sections for new services.
•
Simple construction leads to ease of cleaning, inspection, and tube element replacement.
Disadvantages: •
Hairpin sections are specifically designed units which are normally not built to any industry standard other than ASME Code. However, TEMA tolerances are normally incorporated, wherever applicable.
•
Multiple hairpin sections are not always economically competitive with a single shell and tube heat exchanger.
•
Proprietary closure design requires special gaskets.
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
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Double Pipe Heat Exchanger
Multitube Heat Exchanger
Courtesy of Bastex Corp. Figure 14
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A screening tool that can be used to determine if a hairpin exchanger should be considered for a service under review, is to calculate the product of the estimated overall coefficient (Uo) and estimated area from the equation: Uo A =
Q LMTD
For values of (Uo) (A) greater than 150,000 Btu/hr °F, the hairpin exchanger is probably not economically attractive. For values below 50,000 Btu/hr °F, serious consideration should be given to both the double pipe and multitube hairpin exchangers. Of course, there are always exceptions to these general guidelines for services where a variable such as fluid temperature cross, pressure level, or viscosity of fluid, is an extreme value. It should be noted that hairpin exchangers are particularly useful for services where there is a large temperature cross, because they have true countercurrent flow. The FT factor for a hairpin exchanger is always equal to one. This type of exchanger is also available for high-pressure services with standard designs, for pressures such as 5000 psig on the tubeside and 500 psig on the shellside. Box Coolers A box cooler (worm cooler) is a third type of exchanger used to transfer heat for special services. The cooler consists of a pipe coil submerged in a flooded box, tank or trough (see Figure 15). A box cooler is used for emergency cooling services, where normally there is little or no cooling duty and within a very short time a significant emergency cooling duty must be accommodated. The sudden increase in cooling duty is handled by the pool of water in the tank around the coil. Automatic activation of a cooling fluid supply for this emergency service is not dependable because seldom-used instrumentation does not always respond to an emergency. However, box coolers are used on a restricted basis, because on a square foot of surface basis, they are very expensive. Sometimes, the emergency cooling service involves a solids-bearing stream (catalyst fines) that is very corrosive or erosive. The box cooler is valuable for a corrosive solids-bearing service because the coil can be made from standard types of pipe with extra heavy walls to withstand the fluid.
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Box cooler (Submerged Pipe Coil)
Source: Process Heat Transfer by Donald Q. Kern, page 724, copyright 1950 and 1978 by McGraw-Hill Book Company, Inc. With permission from McGraw-Hill Book Company, Inc.
Figure 15
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Enhanced Heat Transfer The heat transfer surface for the different types of exchangers can be improved (higher rate of heat transfer per exchanger unit size) by enhancing the heat transfer surface. The enhancement can be produced by: •
Adding fins.
•
Providing nucleate boiling surfaces.
•
Installing turbulence promoters.
•
Providing online mechanical cleaning of the surface.
Sometimes, enhancement techniques are used in combination. The enhancements can be used on an existing exchanger where more heat transfer is needed. A service duty increase or temperature driving force decrease, resulting from changes to the processing conditions, can justify surface enhancement. Examples of some high-performance fin geometrics and enhanced boiling surfaces are shown on Figures 16 and 17, respectively. Figure 18 summarizes the use of special surface geometrics and other enhancement techniques for various heat transfer modes. When an enhanced heat transfer surface is evaluated for an application: •
The dominant thermal resistance in the exchanger must be identified.
•
Exchanger service limitations such as flow rate, pressure drop, etc., must be quantified.
•
The design objective defined (will Q, Uo, or A be held constant, and which variable will be changed.)
Enhancement will not significantly change the overall Uo unless it alters the dominant thermal resistance. Finally, practical concerns of cost and possibility of fouling must be evaluated before the design is fabricated.
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HIGH-PERFORMANCE FIN GEOMETRICS
Enhanced surface for gases (A) to (F) and for condensing (G). (A) Offset strip fins used in plate-fin heat exchanger. (B) Louvered fins used in automotive heat exchangers. (C) Segmented fins for circular tubes. (D) Integral aluminum strip finned tube. (E) Louvered tube-and-plate fin. (F) Corrugated plates used in rotary regenerators. (G) Integral low-fin condenser tube.
Figure 16
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ENHANCED BOILING SURFACES
(A) (B) (C) (D) (E)
Rolled-over low fins. Tunnel and pore arrangement. Flattened low fins. Knurled low fins. Sintered porous metallic matrix surface.
Figure 17
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APPLICATION OF ENHANCEMENT TECHNIQUES Commercial
Forced
Boiling
Conden-
Availability
Convection
Mode
sation
Typical Material
Performance
Metal coatings
Yes
--
2
--
Al, Cu, Steel
High
Integral fins
Yes
2
3
4
Al, Cu
High
Flutes
Yes
4
4
4
Al, Cu
Moderate
Integral roughness
Yes
2
3
4
Cu, Steel
High
Wire coil inserts
Yes
3
4
4
Any
Moderate
Displaced promoters
Yes
2
4
4
Any
Mod. (lam)
Twisted tape inserts
Yes
2
3
4
Any
Moderate
Metal
Yes
--
2
4
Al, Cu, Steel
High (boil)
Nonmetal
No
--
4
4
"Teflon"
Moderate
Roughness (integral)
Yes
3
2
4
Al, Cu
High (boil)
Roughness (attached)
Yes
3
4
--
Any
Mod. (for conv)
Axial fins
Yes
1
4
4
Al, Steel
High (for conv)
Gases
Yes
1
--
--
Al, Cu, Steel
High
Liquids/two-phase
Yes
1
1
1
Any
High
Integral
Yes
--
--
2
Al, Cu
High
Nonintegral
Yes
--
--
4
Any
High
Metal coatings
Yes
--
3
--
Al
High
Surface roughness
Yes
4
3
4
Al
High (boil)
2
2
Al, Steel
High
Potential
Inside Tubes
Outside Circular Tubes Coatings
Transverse fins
Flutes
Plate-Fin Heat Exchanger
Configured or interrupted fins Flutes
No
--
--
4
Al
Moderate
Metal coatings
No
--
4
--
Steel
Low
Surface roughness
No
4
4
4
Steel
Low
Configured channel
Yes
1
3
3
Steel
High (for conv)
Plate Type Heat Exchanger
where: 1 2 3 4
= = = =
Common use. Limited use. Some special cases. Essentially no use. Figure 18
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CALCULATING FURNACE EFFICIENCY General All furnaces and heaters are classified in one of two categories; direct fired heaters or fire tube heaters. Because most furnaces and heaters in a refinery are direct fired, the following discussion will be limited to direct fired equipment. However, a brief summary of fire tube heater types, their characteristics, and how they compare with direct fired heaters is given below as general background. Direct Fired
Fire Tube Applications
Hot oil heater.
Indirect fired water bath heaters (line heaters).
Regeneration gas heaters.
Propane and heavier hydrocarbon vaporizers.
Amine and stabilizer reboilers.
Hot oil and salt bath heaters. Glycol and amine reboilers. Low pressure steam generators. Characteristics
More ancillary equipment and controls.
Heat duty usually less than 10 MBtu/hr.
Higher thermal efficiency.
Easily skid mounted.
Requires less plot space.
Forced or natural draft combustion.
Forced or natural draft combustion.
Less likely to have hot spots or tube rupture.
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
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Direct Fired There are two basic types of direct fired furnaces, cylindrical and cabin. Within each type there are many different configurations. The furnaces can have different coil arrangements: horizontal, vertical, helical, or serpentine. Also, the furnace can be all radiant (no convection section) or have a convection section. Several configurations for the vertical cylindrical and cabin type furnaces are shown in Figures 19 and 20. The all radiant cylindrical furnace is the simplest and least expensive. Typically, an all-radiant furnace operates with about a 60% efficiency and a stack temperature of about 1200°F. Adding a convection section to an all-radiant vertical cylindrical furnace increases the overall furnace efficiency to about 80%. Of course, the convection section significantly increases the furnace cost. Some of the advantages for the two types of direct-fired furnaces are as follows. Cylindrical Furnace Advantages: •
Require the smallest plot area for a given duty.
•
The cost is usually 10 to 15% lower in the larger sizes.
•
Can accommodate more parallel passes in the process coil.
•
For large duties, a cylindrical heater has a taller firebox and more natural draft at the burner.
•
The flue gas velocity is usually higher in the convection section, hence, the flue gas film coefficient is higher.
•
Few expensive tube supports or guides are required in the convection section.
•
The noise plenums or preheated combustion air plenums are smaller.
•
Fewer soot blowers are required in the convection section. (Soot blowers are not needed for gaseous fuel.)
•
If coil drainage is a problem, a helical coil may be used when there is only one pass.
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Cabin Furnace Advantages: •
The process coil can always be drained.
•
Two-phase flow problems are less severe (slug flow can generally be avoided).
•
Cabins can accommodate side-firing or end-firing burners instead of only vertically upward firing. This permits the floor of the heater to be closer to the ground. (Some burner manufacturers prefer to fire liquid fuels horizontally.)
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
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EXAMPLES OF VERTICAL CYLINDRICAL DIRECT FIRED FURNACES
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 19
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EXAMPLES OF CABIN DIRECT FIRED FURNACES
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 20 Saudi Aramco DeskTop Standards
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The major components of a furnace are the radiant section (firebox), convection section, stack, burner fuel system, and process fluid coil. The radiant section provides the high level heat to the process coil, with the process fluid flow leaving the furnace via the radiant section. The burner flame is contained in the radiant section. The combustion gases leaving the radiant section typically are in the 1500-1900°F temperature range. Heat is transferred from the flame to the process coil mainly by radiation from the flame. The hot gases of combustion leave the radiant section and flow into the convection section, which transfers the low-level heat to the cold process fluid as it enters the furnace. The combustion gases are cooled in the convection section from the 1500-1900°F range to less than 750°F. Heat is transferred from the gases of combustion (flue gas) to the process fluid coil via convection (hot gas moving over pipes). The stack sits on top of the convection section and generates sufficient draft to overcome the friction losses of the hot flue gas flowing over the convection section tubes. If pollution considerations set the stack height higher than is needed for draft, a damper in the stack can absorb the incremental available draft. Refer to Figure 21 for an illustration of furnace draft. The burner/fuel system includes the burner, which mixes air with fuel and burns the fuel in the radiant section of the furnace. The burner flame typically is about 60% of the height of the radiant section. Fuel and air are fed to the burner by separate pipe/duct systems. The process fluid coil carries the process fluid being heated in the furnace from the process inlet in the convection section (flue gas outlet) to the process outlet in the floor of the radiant section. The coil changes in configuration (horizontal, vertical, low tube finned tube) and type of materials throughout the furnace. The coil is exposed to relatively mild conditions at the process inlet in the convection section and to severe conditions in the radiant section.
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FURNACE NATURAL DRAFT PROFILES
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 21
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Forced Draft Furnaces Forced draft furnaces are used when firing fuel oil. As opposed to natural draft furnaces in this service, forced draft systems have the advantages of fewer burners, less burner maintenance, better air/fuel mixing, and closer excess air control. It is also easier to use an air preheater with this system. Combustion Air Preheaters This method of waste heat recovery is one of the two main methods of optimizing the thermal efficiency of fired equipment, the other being the waste heat boiler. An air preheater is a heat exchanger that is used to transfer heat from the flue gas, leaving a fired heater to the air used for combustion. In this manner, the heater efficiency is increased by reducing the stack temperature below that normally obtained. The cost of the air preheater system must be justified by the resulting fuel savings. In addition to the air preheater itself, the air preheater system consists of forced and induced draft fans, ducting for flue gas and air, tight shutoff and modulating dampers, and special safety controls and instrumentation. Efficiency Furnace efficiency is the ratio of the heat absorbed by the process fluid to the total heat fired in the furnace (QA/QF). When the heat fired is determined, the heat content of the fuel is expressed two ways, as a high heating value (HHV) and as a low heating value (LHV). Either heating value could be used in the furnace efficiency calculation, however, by convention the LHV is always used. The heat absorbed by the process fluid (QA) is calculated from the furnace operating conditions. The method usually used is QA = (Cp)(W)(Æt), the same method previously discussed in ChE 101.09 for calculating shell and tube heat exchanger duty. Once the type of fuel is defined, the furnace flue gas temperature is measured, and the percent excess air at the burner is calculated from the flue gas analysis; curves are used to determine how much of the heat fired (QF) is available for absorption by the process coil (HA). The required quantity of fuel (net) that has to be fired is: FN =
QA (Btu/hr) = lb fuel/hr HA (Btu/lb fuel)
See Maxwell, Pg. 184 for typical curves.
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The fuel to be fired required to meet the QA duty is further adjusted for heat loss from the furnace firebox (radiant section). This loss typically is about 2%. Therefore, the gross quantity of fuel fired to meet QA duty is FG = (FN) (1.02). QF can be determined from the equation QF = FG (lb/hr) x LHV (Btu/lb). Furnace efficiency is: Percent efficiency =
QA (100) QF
The following backup calculation can be done to check the furnace efficiency calculation. Fuel to a furnace is measured by a flowmeter. The actual rate of fuel should be determined from the fuel meter and a backup of QF value calculated from the fuel meter reading. If there is a significant disagreement between QF calculated from the efficiency equation and QF calculated from the fuel meter, this difference should be reconciled before the calculated furnace efficiency is accepted as a credible value. The percent excess air at the burner is calculated from the furnace flue gas analysis. The quantity of excess air at the burner affects the amount of energy in the fuel that is available for absorption by the process fluid coil. Example Problem 3 Part 1 - Calculate Percent Excess Air The measured flue gas composition is (all values vol% for dry flue gas) CO2 = 9.5, CO = 1.8, O2 = 2.0 and N2 = 86.7. The total oxygen to the furnace is determined from the quantity of N2 in the flue gas and the fact that air is 79 vol% N2 and 21 vol% O2.
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FLUE GAS OXYGEN VERSUS EXCESS AIR
Figure 22 Part 2 - Calculate Furnace Efficiency Calculate the thermal efficiency for a furnace where the operating duty QA has been confirmed to be 150 MBtu/hr [specific heat equation was used to confirm duty, Q = (W)(Cp)(Æt)]. Also, the flue gas temperature is 700°F, the flue gas composition is the same as in Part 1 (use 5% for excess air at burner), heat losses from firebox are 3%, and fuel gas fired in the furnace has a LHV of 19,700 Btu/lb.
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Example Problem 3, Part 1 - Percent Excess Air The measured flue gas composition is (all values vol% for dry flue gas) CO2 = 9.5 CO = 1.8 O2 = 2.0 and N2 = 86.7. The total oxygen to the furnace is determined from the quantity of N2 in the flue gas and the fact that air is 79 vol% N2 and 21 vol% O2. moles of N2 moles of O2 O2 supplied to furnace = 100 moles of flue gas 100 moles of air 100 moles of flue gas moles of N2 100 moles of air (86.7) (21) = 23.05 moles of O2 (79) 100 moles of flue gas
Excess oxygen at the burner is expressed on the basis of complete combustion; namely, all hydrogen and carbon in the fuel are burned to water and carbon dioxide. Therefore, the amount of free oxygen in the flue gas analysis must be reduced by the amount needed to complete the combustion of the flue gas CO to CO2. The formula for this adjustment is: (flue gas CO content ) + 0.5 moleO2 = CO2 mole CO 0.9 moleO2 = 1.8 moles CO2 1.8 moles CO + 100 moles flue gas 100 moles flue gas 100 moles of flue gas Adjusted excess O2 =
2.0 molesO2 0.9 moleO2 1.1 molesO2 = 100 moles of flue gas 100 moles of flue gas 100 moles of flue gas
From the flue gas analysis and these calculations, the excess oxygen at the burner can be determined as follows: adjusted moles excess O2 (100) 100 moles flue gas Percent excess O2 = (moles O2 burnt in furnace) 100 moles flue gas adjusted moles excess O2 (100) 100 moles flue gas = (moles O2 supplied to furnace) - adjusted moles excess O2 100 moles flue gas 100 moles flue gas =
(1.1) (100) = 110 = 5% 23.05 - 1.1 21.95
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Example Problem 3 (Cont'd)
Since the percent excess O2 is the same as the percent excess air, the sample calculation shows that the furnace under study is operating with 5% excess air. The nomogram in Figure 22 cqn also be used to determine percent excess air after the adjusted percent oxygen in the flue gas has been calculated. FLUE GAS OXYGEN VERSUS EXCESS AIR
(drawing)
Figure 22
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Example Problem 3, Part 2 - Furnace Efficiency Calculate the thermal efficiency for a furnace where the operating duty QA has been confirmed to be 150 MBtu/hr [specific heat equation was used to confirm duty, Q = (W) (Cp)(Dt)]. Also, the flue gas temperature is 700°F, the flue gas composition is the same as in Part 1 (use 5% for excess air at burner), heat losses from firebox are 3%, and fuel gas fired in the furnace has a LHV of 19,700 Btu/lb. Q Net fuel = FN = A = 150,000,000 HA 16,600 = 9040 lb/hr fuel (See Maxwell, Pg. 184 for HA) Gross fuel = FG = (FN) (firebox heat loss) = (9040) (1.03) = 9310 lb/hr fuel Heat fired = QF = (FG) (LHV fuel) = (9310) (19,700) = 183,350,000 Btu/hr Percent efficiency=
(Q A )(100) QF
=
150,000,000 = 82% 183,350,000
For a quick approximation of furnace efficiency, the following shortcut formula can be used in conjunction with Figure 22 (percent excess air versus percent O2 in flue gas curve). Percent efficiency 100 = [(100 - (0.0237 + (0.000189) (EA))) T( ST - TA)] 100 + QL (LHV)
where:
EA = Percent excess air. TST = Stack temperature, °F. TA = Ambient air temperature, °F. QL = Casing heat loss.
For Example 3, Part 2 conditions, the furnace efficiency calculated from the shortcut formula is as follows: Percent efficiency = [(100 - (0.0237 + (0.000189) (5))) (700 - 80)] 100 100 + 3 Percent efficiency = [100 - (0.0246) (620)] (0.971) = 82.3
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Routine Furnace Startup and Operations Startup The complexity of fired heaters is increasing. Today, furnace complexity often dictates that a furnace startup advisor be present for major startups. The advisor, together with mechanical, instrument, and burner specialists, review in detail the heater piping and instrumentation. Upon completion of this review and corrective actions required, the heater is ready to be lit for lining dryout. The following activities are expected from the startup personnel during the lining dryout and initial furnace operation. •
Review the Operating Manual and revise the fired heater section as necessary (prestartup, oil in, normal operations, shutdown procedures, troubleshooting, and auxiliary equipment instructions).
•
Ensure that hydrostatic test water has been removed from the coil to the maximum extent practical.
•
Ensure that all fuel lines have been steam blown (not through the burner guns).
•
Check the performance of all the burners during refractory dryout.
•
Monitor thermal movements of tubes, tube support systems, and refractory during dryout. Watch for debris on the heater floor.
•
Investigate any performance data for the fired heater and attendant equipment, that appears to differ from design specification values. Listed below are some of the more important general observations to be made and problems to look for during an initial startup. _ _ _ _ _ _ _ _ _ _ _ _
Coil and external piping movements. Lining condition as heater reaches operating temperature. Pass flows, pass crossover, and outlet temperature. Tube hot spots and overheated passes. Burner and pilot combustion performance. Watch for problems such as fuel dripback, gun orifice plugging, gun tip coking, wet atomizing steam, uneven burner firing rates, leaning flames, flame impingement, burner noise, etc. Draft conditions, particularly at bridgewall. Combustion air pressure. Expansion joint movement. Damper positions. Stack vibration. Fan-induced vibration and noise. Unsafe operating practices.
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•
Discuss special problems related to the specific fired heater in the operating manual.
When all prestartup activities have been completed (equipment checkout completed, linings dried, etc.), at the appropriate time in the unit oil-in operation, the furnace will be lit and put online. The following furnace startup steps are listed for background information and should not be considered complete. In each startup procedure, certain aspects of the procedure are unique to a particular service. •
Check to see if all fuel and pilot systems are active up to unit battery limits.
•
Check to see if all drains and vents of the on-fuel/pilot systems are closed.
•
Check to see if all instrumentation is working and that automatic shutdown devices are deactivated.
•
Commission any fuel oil steam tracing and open blinds at battery limits on fuel/pilot supplies.
•
Steam out fuel oil system to bring piping up to temperature.
•
Open furnace stack damper fully.
•
Start snuffing steam to furnace firebox and shut off snuffing steam when a good flow of steam can be observed from the stack.
•
Fully open air dampers on each burner.
•
Open valves on pilot gas system to purge inert gases.
•
Ignite fuel gas pilots, one burner at a time.
•
Open valve to bring fuel oil and tracing steam into the burner supply systems.
•
Start atomizing steam to the first burner.
•
Slowly open the fuel oil valve to the first burner and observe ignition. Adjust oil and air rates to give a stable, nonluminous flame. Set firing at a minimum stable rate consistent with a good flame pattern.
•
Repeat for each burner.
•
When all burners are lit, check for proper operation of pilot and burner flames.
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•
Activate furnace instrumentation and raise furnace coil outlet temperature at the rate of about 50°F/hr.
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Optimum Excess Air Levels As part of the discussion on furnace efficiency, some of the furnace operating variables have been discussed, namely: •
Checking the furnace duty QA by the enthalphy heat equation QA = W(Æh), and
•
Utilizing the stack flue gas temperature and oxygen content as part of burner operating conditions review.
For these variables operating the burner with the correct amount of excess air (determined from O2 level in flue gas) has the most significant effect on the entire operation of the furnace. Therefore, this discussion will further explore proper excess air levels for furnace burners. Figures 23 and 24 show the effect of different levels of excess air on the furnace efficiency and level of combustibles in the stack flue gas. An excess air target should be established for each furnace, and the operating level versus the target level should be monitored by the plant engineer or by automatic instrumentation. The target excess air level is established by plugging air leaks in the furnace walls and then reducing the air rate to the burner in increments while monitoring the carbon monoxide and smoke level in the stack flue gas. When the carbon monoxide level reaches the 100-200 ppm range, the minimum acceptable excess air level has been reached. The actual monitored target level for excess air will be at or close to this minimum level, as determined by the actual furnace service and associated instrumentation under study. Significant fuel savings can be made by monitoring burner excess air: a 40°F decrease in flue gas temperature usually produces about a 1% increase in furnace efficiency. The cost effect of unplugged air leaks is shown in Figure 25. Lowering the stack temperature to improve efficiency is usually limited by return on investment and the acid dew point in the flue gas (discussed in ChE 101.02). Operating Guidelines Low Draft
High Draft
Low Excess Air (O2)
Open Damper
Open Burner Air
High Excess Air (O2)
Close Burner Air
Close Damper
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OPTIMUM EXCESS AIR FOR A FIRED HEATER
Figure 23
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TYPICAL COMBUSTIBLES EMISSION FROM FIRED HEATERS
Figure 24
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COST OF FURNACE AIR LEAKS
(use photostat)
Figure 25
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Monitoring Devices and Techniques A more extensive discussion on monitoring furnace operations will be presented in a later course. However, for background information, the following are some typical items that can be monitored: •
Process fluid flow pass balancing.
•
Tube metal temperatures.
•
Bridge wall temperature.
•
General firebox conditions.
•
Safety equipment.
•
Turnaround checklist.
There are computer applications that can be used for furnace monitoring. Controls/Safety Devices/Burners Furnace controls and safety devices vary considerably depending on the furnace service and the refinery location. Figure 26 shows an example control system for a direct fired heater. It should not be considered complete, but only representative of the type of instrumentation that should be carefully considered in designing a control system for a furnace service.
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EXAMPLE OF DIRECT FIRED REBOILER CONTROLS/SAFETY DEVICES
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book.
Figure 26 The alarms and a description of the shutdown systems shown in Figure 26 are listed on the following page.
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HEATER ALARM/SHUTDOWN SYSTEM DESCRIPTION Caution: The alarms and shutdowns shown do not necessarily meet any minimum safety requirement, but are representative of the types used for control systems. Basic Criterion: The failure of any one device will not allow the heater to be damaged. Schematic Label
Alarm/Shutdown Description
Regeneration Gas Heater
Hot Oil Heater and Direct Fired Reboiler
TSH-1
High stack temperature.
See Note 1.
See Note 1.
TSH-2
High outlet temperature.
See Note 1.
FSL
Low mass flow through tubes.
See Notes 2 and 4.
See Notes 3 and 4.
BSL
Flame failure detection.
See Note 5.
See Notes 5 and 6.
PSL
Low fuel pressure.
PSH-1
High fuel pressure.
See Note 7.
See Note 7.
PSH-2
High cabin pressure.
See Note 8.
Not applicable, if natural draft.
See Note 6.
Notes: 1. A direct immersion jacketed thermocouple is preferred because the response is ten times faster than a grounded thermocouple in a well. A filled bulb system is a poor third choice. The high stack gas temperature shutdown should be set approximately 200°F above normal operation. 2. An orifice plate signal should be backed up by a low pressure shutdown to ensure adequate process stream flow under falling pressure conditions. 3. The measurement should be on the heater inlet to avoid errors from two-phase flow. 4. Differential pressure switches mounted directly across an orifice plate are not satisfactory due to switch does not turn on at the same pressure as it turns off. An analog differential pressure transmitter with a pressure switch on the output is recommended. The analog signal should be brought to the shutdown panel so that the flow level can be readily compared with the shutdown point. With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. 5. The flame scanner should be aimed at the pilot so that a flameout signal will be generated if the pilot is not large enough to ignite the main burner. Saudi Aramco DeskTop Standards
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6. If the heater design precludes flame scanners, a low fuel gas pressure shutdown should be installed to prevent unintentional flameout. This shutdown should detect gas pressure at the burner. 7. Either burner pressure or fuel control valve diaphragm pressure may be used. This shutdown should be used whenever large load changes are expected. It prevents the heater from overfiring when the temperature controller drives the fuel wide open to increase heat output with insufficient air. 8. This shutdown should block in all lines to the heater because the probable cause of its activation is tube rupture. Gas is probably burning vigorously outside the heater. 9. With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. The parts of the fired heater instrumentation related to safety should be given special attention and regularly inspected as well as tested for functionality. The following items, either observed by the plant engineer or indicated by instrumentation, usually indicate a problem with furnace operations. • • • • • • • • • •
The burner flame is not symmetrical, pulsates or breathes, is unusually long or lazy, lifts off the burner, etc. The burner is not aligned and/or the flame is too close to the tubes. There is a lack of negative pressure (draft) at the top of the firebox. The stack gas is smokey. The gas in the firebox appears hazy. There are unequal temperatures, differing by more than 10°F, on the process pass outlets. The stack temperature increases steadily with no change in the process heat duty. The fuel gas control valve is wide open. The fuel gas composition or pressure varies widely. The tubes in the heater are not straight.
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With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. Burner selection is very important because an improper burner will reduce furnace efficiency and service factor. Plant personnel need to have general knowledge of burners so that: • • •
The proper maintenance will be performed. Burner operating problems can be properly diagnosed and corrected. The burner operation can be optimized.
Four types of burners are commonly used in direct fired heaters. •
Inspirating Pre-Mix Burners - The passage of fuel gas through a venturi pulls in the combustion air. These burners have short dense flames that are not affected by wind gusts.
•
Raw Gas Burners - Some of the air required for combustion is pulled in by a venturi. The rest of the air is admitted through a secondary air register. These burners have larger turndown ratios, require lower gas pressures, and are quieter.
•
Low NOx Burners - The addition of a tertiary air register reduces the amount of nitrogen oxides in the flue gas. This type also can be operated with less excess air than inspirating pre-mix or raw gas burners.
•
Combination Gas and Oil Burner - An oil burner is added to the gas spider so that fuel oil can also be used. One-tenth pound of steam per pound of fuel is usually required to atomize the oil.
With permission from the Gas Processors Suppliers Association. Source: Engineering Data Book. Figure 27 shows a cross-sectional view of a combination gas/oil burner. This module (ChE 101.10) contains limited information on the selection and operation/maintenance of burners. Future courses will explore these subjects in greater depth.
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NATURAL DRAFT OIL/GAS BURNER
Figure 27
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Monitoring Tube Metal Temperature One critical variable to monitor in many furnaces after startup is the tube metal temperatures (temperature of the process coil on the firebox side) in the radiant section. Tube failures account for more than half of the furnace fires and explosions. Excessive tube metal temperatures accelerate tube creep (sagging tubes), hydrogen attack, and external (vanadium attack, oxidation) and internal corrosion of the tubewall. Monitoring tube metal temperatures also helps define the end of the current furnace run, the point at which the furnace is due for a shutdown and decoking. Tube metal temperatures increase with coke laydown in the tube, assuming all other variables are held constant. Tube metal temperatures are monitored by thermocouples attached to the tube, or by a pyrometer that measures the radiation emitted by the furnace tubewall. An expert on tubewall temperature instruments should be consulted about the installation of thermocouples or the purchase of a pyrometer. Both measuring devices are sophisticated pieces of equipment that vary in type depending on the service. Usually, the highest tube metal temperature in a direct fired heater occurs in the radiant section, where the process fluid temperature on the inside of the tube is the highest. The maximum allowable operating tube metal temperature for any one tube material is a function of the tubewall stress level and the severity of the tubewall corrosive atmosphere. Figure 28 gives allowable stress levels for furnace tubes as a function of wall temperature for several material types.
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ALLOWABLE ELASTIC AND CREEP RUPTURE STRESS FOR TYPICAL HEATER TUBE MATERIALS (Elastic and Creep Rupture Stress, psi)
Temp. °F(1) 700 750 800 850 900 950 1,000 1,050 1,100 1,150 1,200 1,250
Medium Carbon Steel Elastic Creep Stress Stress
15,800 15,500 15,000 14,250 13,500 12,600 11,500
20,800 16,900 13,250 10,200 7,500 5,400 3,700
C-1/2 Mo Elasti Creep c Stress Stress 15,700 15,400 15,000 14,500 14,000 13,400 12,700 11,900 10,900
17,000 10,250 5,900 3,400 2,000
1-1/4 Cr-1/2 Mo Elastic Creep Stress Stress
15,250 15,000 14,600 14,250 13,800 13,300 12,800 12,100 11,400
17,500 10,900 6,700 4,150 2,600
2-1/4 Cr-1 Mo Elasti Creep c Stress Stress 18,000 18,000 17,900 17,500 17,100 16,500 15,750 14,750 13,600 12,300 10,700
16,700 12,100 8,700 6,400 4,600 3,150 1,750
5 Cr-1/2 Mo Elasti Creep c Stress Stress 16,800 16,500 15,900 15,200 14,400 13,500 12,400 11,300 10,250 9,200 8,200
13,250 9,600 7,000 5,100 3,700 2,700 1,950
Notes: (1) For intermediate temperatures, stresses can be obtained by graphical interpolation. Source: API Recommended Practice 530, Calculation of Heater Tube Thickness in Petroleum Refineries, Third Edition, September 1988. Reprinted courtesy of the American Petroleum Institute. Figure 28
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NOMENCLATURE af ai Ao ax
Face area of tube bundle, ft2. Tube inside area, ft2. Heat transfer surface area based on bare tube O.D., ft2. Tube finned area, ft2.
Cp De
hi ho k Nf
Specific heat, Btu/lb °F. Equivalent diameter of the bank of finned tubes, (P/dr)2 (Nf) (do/12), ft. Inside diameter of tube, in. Logarithmic mean diameter of tube, in. Outside diameter of fins, in. Inside diameter of fins, in. Fan efficiency. Driver efficiency. Gross fuel fired, lb/hr. Net fuel fired, lb/hr. LMTD correction factor, from Figure 5. Heat available from fuel, Btu/lb fuel. Air-side film heat transfer coefficient based on bare tube O.D., Btu/hr ft2. Inside film heat transfer coefficient, Btu/hr ft2 °F. Outside film heat transfer coefficient for finned tube, Btu/hr ft2 °F. (Nf) [do/(dR)0.2]. Number of fins per inch of tube length.
P Pst Pv PT ÆP
Center-to-center distance (pitch) of tubes, in. Air static pressure at fan, inH2O. Air velocity pressure at fan, inH2O. Pst + Pv, inH2O. Pressure drop, inH2O/row of tubes.
Q QA QF
Total heat load of exchanger, Btu/hr. Furnace process heat absorbed, Btu/hr. Heat fired, Btu/hr.
di dm do dR etaf etad FG FN FT HA ha
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R rdi rdo rm RS T1 T2 t1 t2 Æte Ætm Uo
Overall resistance to heat flow, hr ft2 °F/Btu. Fouling resistance inside tube, hr ft2 °F/Btu. Fouling resistance outside tube, hr ft2 °F/Btu. Resistance to heat flow of tube wall, hr ft2 °F/Btu. Outside surface area finned tube Outside surface area bare tube
Ux
Inlet temperature of fluid to be cooled, °F. Outlet temperature of fluid to be cooled, °F. Inlet air temperature, °F. Outlet air temperature, °F. Effective temperature difference, °F. Logarithmic mean temperature difference (LMTD), °F. Overall heat transfer coefficient (related to bare tube O.D.), Btu/hr ft2 °F. Overall heat transfer coefficient (related to finned tube), Btu/hr ft2
Vf Vo Vmax
°F. Face velocity of cooling air, ft/min at 70°F, 29.92 inHg. Air velocity at outlet of fan (70°F, 29.92 inHg), ft/min. Air velocity through the minimum free flow area of tube banks.
Ws Wt X Y
Ft/min of air at 70°F, 29.92 inHg. Air rate, lb/hr. Process fluid rate, lb/hr. (T1 - T2)/(T1 - t1) (see Figure 5). (t2 - t1)/(T1 - T2) (see Figure 5). Figure 29
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KEY FORMULAS The following is a brief summary of the more important and useful formulas in this module. Note that many of them can be easily programmed for personal computer use. LMTD correction factorFT: (Figure 5) X = T1 - T2 T1 - t1 where: T1 T2 t1 t2
Y = t2 - t1 T1 - T2
= Inlet temperature of fluid to be cooled, °F. = Outlet temperature of fluid to be cooled, °F. = Inlet air temperature, °F. = Outlet air temperature, °F.
LMTD: LMTD = GTTD - LTTD ln GTTD LTTD where: GTTD = Greater temperature difference. LTTD = Lesser temperature difference. Air fin face velocity,Vf Vf =
lb/hr air = ft/min 60 29 0.98 379
Air velocity through min free flow area, Vmax: Vmax =
Vf = ft/min (P - dR) af P af
where: Vf = Face velocity, ft/min. P = Center-center of tubes, in. dR = Inside diameter of fins, in. af = Face area, ft2.
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Outside film heat transfer coefficient, ho: ho =
1.9 dR (Vmax)0.56 P Nf (do)
0.5
= Btu/hr ft2 °F
where: Nf = Number of fins/in. of tube. do = Outside diameter of fins, in. dR = Inside diameter of fins, in. P = Center-to-center distance (pitch) of tubes, in. Vmax = Air velocity through the minimum free flow area of tube banks. Ft/min of air at 70°F, 29.92 inHg. Ratio finned surface to bare surface, ax/ai: ax = RS dR ai di where: ax = Tube finned surface, ft2. ai = Tube inside area, ft2. RS = Outside surface area finned tube. Outside surface area bare tube dR = Inside diameter of fins, in. di = Inside diameter of tube, in. Overall heat transfer coefficient, Ux: 1 = 1 ax + rdi ax + rm + 1 ai Ux hi ai ho where: ax ai rdi rm hi ho
= Tube finned surface, ft2. = Tube inside area, ft2. = Inside tube fouling resistance. = Tubewall resistance. = Inside tube film coefficient. = Outside film heat transfer coefficient for finned tube, Btu/hr ft2 °F.
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WORK AID 1 - PROCEDURES FOR CALCULATING PERCENT OF DESIGN DUTY To calculate the percent of design duty an airfin exchanger can perform when the air temperature rises above design, use the following procedure: A. Step 1: Calculate X and Y at design conditions: X = T1 - T2 T1 - t1
Y = t2 - t1 T1 - T2
Step 2: Determine FT, using Figure 6. Step 3: Calculate LMTD: GTTD - LTTD lnGTTD LTTD
or, use chart from TEMA Manual, Pg. 111.
Step 4: Calculate Æte: Æte = FT (LMTD)
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B. At the heat wave conditions, use a trial-and-error calculation where the air temperature use is affected by the decrease in the exchanger's capability to transfer heat. For the first trial calculation, estimate that the duty transferred by the exchanger is decreased by 15%. Also, since the exchanger surface is constant, the change in duty versus air temperature must be considered. Use the formula: Q1 Q2 = A1 = A2 = U1 ∆te1 U2 ∆te2 Therefore,
AssumeU1 = U2
Q1 Q = 2 ∆te1 ∆te2
Step 1: Calculate the process fluid temperature drop at the assumed 85% duty. Step 2: Calculate the air temperature rise at the assumed 85% duty. Step 3: Calculate the process and air outlet temperatures. Step 4: Determine X and Y values. Step 5: Determine FT, using Figure 6. Step 6: Determine LMTD. Step 7: Calculate Æte. Step 8: Check the assumption of 15% heat duty reduction, using: Q2 = ∆T2 = Q1 ∆T1
should equal the assumed duty (in this case, 0.85)
If not, assume a new duty and recalculate.
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WORK AID 2 - PROCEDURES FOR CALCULATING EXTENDED SURFACE AND FACE AREA REQUIREMENTS To calculate the extended surface and face area requirement of an airfin exchanger: Step 1: Calculate the estimated air temperature use, using the formula: Estimated use = Ux + 1 T1 + T2 - t1 10 2 Step 2: Calculate X and Y: X = T1 - T2 T1 - t1
Y = t2 - t1 T1 - T2
Step 3: Using Figure 6, determine FT. Step 4: Calculate LMTD. Step 5: Calculate Æte: Æte = (LMTD) (FT) Step 6: Calculate ax, using the formula: ax =
Q Ux ∆ te
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Step 7: Calculate: T1 - t1 Uo Step 8: Use Figure 7 and determine optimum bundle depth. Step 9: Calculate af, using the formula: af =
ax Ao/af ax/Ao
Step 10: Calculate Ws, using the formula: Ws =
Q ∆ta Cp
Step 11: Calculate ACFM air, using the formula: ACFM =
Ws 60 air density at 120°F and 1 atm
Step 12: Calculate Vf, using the formula: Vf = ACFM air af Step 13: Calculate Vmax, using the formula: Vmax = Vf
af open area between tubes
where: Open area between tubes =af P - dR . P
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Step 14: Calculate ho, using the formula: ho =
1.9 dR (Vmax)0.56 P Nf (do) 0.5
Step 15: Calculate ha, using the formula: ha = (RS) (ho) where: RS = ax/ab Step 16: Calculate Uo, using the formula: d d d r 1/Uo = 1 + rm R + rdi R + 1 R + do ha dm di hi di RS Neglect wall resistance (rm) and outside fouling resistance (rdo). Step 17: Calculate Ux, using the formula: 1/Ux = 1 + rdi ax do + 1 ax do ho Ao di hi Ao di Step 18: Compare Ux calculated versus assumed and recalculate if necessary. (For Exercise 2, the participant is instructed to stop here.)
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WORK AID 3 - PROCEDURES FOR CALCULATING FURNACE EFFICIENCY To determine a furnace thermal efficiency, follow the steps listed below: Step 1: Calculate oxygen to furnace, using the formula: moles N2 moles O2 100 moles flue gas 100 moles of air O2 to furnace/100 moles flue gas = moles N2 100 moles of air moles O2 21 100 moles flue gas = 79 Step 2: Calculate percent excess oxygen, using the formula: moles O2 from furnace 100 100 moles flue gas Percent excessO2 = moles O2 to furnace - moles O2 from furnace 100 moles flue gas 100 moles flue gas Percent excess O2 = percent excess air. Note: If there is CO in the flue gas, refer to Example Problem 3 for adjustment of excess oxygen. Step 3: Determine furnace duty, QA: QA = (W) (Cp) (Æt)
(If vaporization occurs, use enthalpy data.)
Step 4: Determine heat available (HA) per lb of fuel from Maxwell, Pg. 185.
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Step 5: Calculate net fuel fired, FN: FN =
QA FN
Step 6: Calculate gross fuel fired, FG. Assume furnace heat losses are 2 1/2%. FG = 1.025 (FN) Step 7: Calculate heat fired, QF, Btu/hr. QF = (FG) (LHV fuel) Step 8: Calculate furnace efficiency: % efficiency =
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GLOSSARY acid dew point
The temperature at which an acidic component in the flue gas condenses.
A-frame exchanger
Two sections of air coolers are hooked together with each sloping 45° in opposite directions, forming an A shape. This is done to reduce plot space.
air plenum chamber
An empty chamber under or over an air cooler that allows air to spread or be collected from the air cooler face area.
battery limits
The real estate boundary line denoting the start of a processing area.
bubble point curve
Denotes the family of pressure/temperature points at which the fluid starts to vaporize.
burner registers
The openings, equipped with a regulating device, in the burner to feed air to the burner.
contamination
The quality of a substance is made unacceptable by a contaminant. This reduction in quality is called contamination.
corbeled wall
An irregular wall in the convection section of a furnace. The irregularities in the wall match the staggered tubes in the convection section to prevent flue gas from bypassing around the tubes.
cross flow exchanger
The tubes carrying the process fluid in an air-cooled exchanger usually run horizontally. The air flow is vertical; therefore, it flows in a cross flow manner across the tubes.
dew point curve
Denotes the family of pressure/temperature points at which the fluid starts to condense.
embedded type fin
The fin is wrapped around the tube, sometimes held in place by grooves cut in the tube.
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extended surface
The metal surface area for a tube including the tube fin area.
extruded type fin
A design in which the sleeve that slides over the bare tube to hold the fin and the fins are extruded from one piece of metal.
face area
The flat projected area of an air cooler.
flame impingement
The burner flames touch the tubes.
furnace bridgewall
The top of the radiant section where the hot gases enter the convection section.
isenthalpic expansion
An expansion (drop in pressure) during which the enthalpy of the stream stays constant.
lining dryout
Most furnace refractory linings have been mixed with water at installation; this excess water is slowly boiled out of the lining during dryout.
louvers
Strips of metal that open and close to regulate the flow of air through the tube bundle of an air cooler.
maximum air velocity
The velocity of the air for an air cooler, flowing between the tubes in a tube row.
nonluminous flame
An improperly adjusted burner flame is yellow and luminous (it gives off light). A properly adjusted flame is a light blue and gives off very little light.
plot area
The ground area required by an air cooler.
pressure/enthalpy diagram
Heat content (enthalpy) and pressure are the abscissa and ordinate of this diagram with constant temperature and enthalpy lines. Used for design of refrigeration systems.
refrigerant
The fluid that is circulated in the refrigeration system to provide refrigeration.
serrated fin
A fin surface with many cuts, breaks.
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slug flow
A very uneven flow with alternate slugs (sections) of liquid followed by vapor.
snuffing steam
Steam injected into furnace firebox to extinguish a fire or purge the firebox to prevent an explosion when burners are lit.
stack gas combustibles
Material in the stack gas that was not fully burnt (combusted) in the furnace.
static head
A fan develops a low discharge pressure, called static head, in order to create an air discharge velocity.
toxicity
The adverse effect of a substance on human health.
variable pitch fans
A fan in which the angle (pitch) of the blades can be changed to regulate the flow of air through the tube bundle of an air cooler.
velocity pressure
The air velocity from a fan can be expressed in an equivalent set of units called static pressure.
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REFERENCES 1. Chemical Engineer's Handbook, 6th Edition, R. H. Perry and D. Green (Physical properties, general information, calc equations). 2. Data Book on Hydrocarbons, J. B. Maxwell (Hydrocarbon physical properties). 3. Engineering Data Book, Gas Processors Suppliers Association, 10th Edition, 1987. 4. Process Heat Transfer, Kern, 1950. 5. Standards of the Tubular Exchanger Manufacturers Association (TEMA), 7th Edition, 1988. 6. AES-E-001, Basic Design Criteria for Unfired Heat Transfer Equipment. 7. ADP-E-001, Exchangers.
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APPENDICES AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET
(photostat)
Figure 30
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AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET (CONT'D)
(photostat)
Figure 30 (cont'd)
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Process Other Heat Transfer Equipment
AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET (CONT'D)
(photostat)
Figure 30 (cont'd)
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Engineering Encyclopedia
Process Other Heat Transfer Equipment
AIR COOLED HEAT EXCHANGER SPECIFICATION SHEET (CONT'D)
(photostat)
Figure 30 (cont'd)
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