Basics of Air Cooled Heat Exchangers

Home , Cooling, Fan (machine), Heat capacity rate

The Basics of AIR-COOLED HEAT EXCHANGERS

HUDSON Products Corporation A Subsidiary of Hudson Products Holdings, Inc.

www.hudsonproducts.com CONTENTS Page

FOREWORD ...... i NOMENCLATURE ...... ii

I. DESCRIPTION OF AIR-COOLED HEAT EXCHANGERS ...... 1 Components ...... 1 Tube Bundle ...... 1 Axial Flow Fans ...... 5 Plenum ...... 6 Mechanical Equipment ...... 6 Structure ...... 6 Comparison of Induced and Forced Units ...... 7 Induced Draft ...... 7 Forced Draft ...... 7 II. THERMAL DESIGN ...... 8 Typical Heat Transfer – Case I ...... 13 Application of Design Method ...... 19 Sample Problem ...... 20 Fan Selection – Horsepower Requirements ...... 21 III. PERFORMANCE CONTROL OF ACHEs ...... 21 Varying Air Flow ...... 23 Extreme Case Controls ...... 23 Internal Recirculation ...... 23 External Recirculation ...... 23 Co-current Flow ...... 23 Auxiliary Heating Coils – Steam or Glycol ...... 23 IV. NOISE CONTROL ...... 24 V. DESIGN OF ACHEs FOR VISCOUS LIQUIDS ...... 24 VI. COST ...... 25

REFERENCES ...... 25 TABLES Page

1. Guide to First Estimates of Bundle Rows ...... 19 2. Typical Heat Transfer Coefficients for Air-Cooled Heat Exchangers ...... 19

FIGURES Page

1. Typical Components of an Air-Cooled Heat Exchanger ...... 2 2. Typical Construction of Tube Bundles with Plug and Cover Plate Headers ...... 3 3. Fin Types ...... 4 4. Header Types ...... 5 5. Axial Flow Fan ...... 5 6. Mechanical Components ...... 6 7. Comparison of Induced and Forced Draft Units ...... 7 8. Thermal Optimization Parameters ...... 8 9. MTD Correction Factors, One Pass – Cross Flow ...... 10 10. MTD Correction Factors, Two Pass – Cross Flow ...... 11 11. MTD Correction Factors, Three Pass – Cross Flow ...... 12 12. Air Cooler Sizing Chart – One Pass ...... 15 13. Air Cooler Sizing Chart – Two Pass ...... 16 14. Air Cooler Sizing Chart – Three Pass ...... 17 15. Air Cooler Sizing Chart – Four Pass ...... 18 16. Unit Weight and Surface per Unit Fan HP as a Function of Bundle Depth ...... 22 17. Air Flow Control ...... 23 18. Unit Price as a Function of Total Surface and Bundle Depth ...... 26 FOREWARD

This brochure is designed to familiarize user with the types, components, and features of Air-Cooled Heat Exchangers (ACHEs) by means of simplified explanations and procedures. It discusses the advantages and disadvantages of each type of help designers become more discriminating, competent, and confident as users of this equipment.

The brochure also provides a method of estimating ACHE size, weight, price and power consumption in the planning stage, but is not intended to provide information sufficient for detailed and final design. If specific assistance is needed, please contact Hudson Products Corporation at (281) 275-8100.

NOMENCLATURE

English Letter Symbols

a=heat transfer surface area per unit length of tube ft 2/ft A=total exchanger bare tube heat transfer surface ft 2 Aw = average wall thickness in BWG = Birmingham wire gauge cp = specific heat Btu/(lb•°F) Cair =Ccold = Q / ∆t = Q / (t 2-t 1) = air-side heat capacity rate Btu/(hr•°F) = 1.08 • FV • L • W Ctube =Chot = Q / ∆T = Q / (T 1-T 2) = tube-side heat capacity rate Btu/( hr•°F) = Mc p Cmin = minimum heat capacity rate Btu/( hr•°F) Cmax = maximum heat capacity rate Btu/( hr•°F) CMTD = corrected mean temperature difference °F = F • LMTD E=exchanger thermal effectives Dimensionless = Chot • (T 1-T 2) Cmin • (T 1-t 1) = Ccold • (t 2-t 1) Cmin • (T 1-t 1) F=MTD correction factor Dimensionless f=Fanning friction factor Dimensionless

i FA = face area ft 2 FV = standard air face velocity std ft/min G = mass velocity lb/(sec•ft 2) h=individual heat transfer coefficient Btu/( hr•ft 2•°F) ID = inside diameter of tube ft K=thermal conductivity Btu/( hr•ft•°F) k=parameter, NTU or NTU • R Dimensionless = n • N • a 1.08 • FV • (I / U) L=tube length ft LE = calmed length ft LMTD = log mean temperature difference °F M= mass flow rate lb/hr mw = minimum wall thickness in n=tubes per row, per foot of exchanger width l/ft

N= rows of tubes in direction of air flow Dimensionless NRe = Reynolds Number Dimensionless Ntu = number of heat transfer units Dimensionless OD = outside diameter of tube ft P=number of tube-side passes Dimensionless Q= total exchanger heat load (duty) Btu/hr r=individual heat transfer resistance (hr•ft 2•°F)/Btu R=Cmin / C max = heat capacity rate ratio Dimensionless S=specific gravity Dimensionless t=air temperature °F T=hot fluid temperature °F Tavg = bulk average temperature °F U= overall heat transfer coefficient (rate) Btu/( hr•ft 2•°F) = 1 (r i + r air + r f + r m) W= width of exchanger ft Z=parameter, E or E • R Dimensionless =T1 – T 2 T1 – t 1

ii Greek Letter Symbols

µ=viscosity centipoise 3 ρ = density lb/ft

Subscripts

air = air side max = maximum cold = cold fluid = air min = minimum f=tube-side fouling t=total hot = hot fluid = tube-side fluid 1=inlet i=tube side 2=outlet m= tube metal

iii AIR COOLED HEAT EXCHANGERS · A plenum between the bundle or bundles and the air-moving device.

A proven means for cooling in the process · A support structure high enough to allow air to and power industries enter beneath the ACHE at a reasonable rate.

I. DESCRIPTION OF AIR-COOLED · Optional header and fan maintenance walkways HEAT EXCHANGERS with ladders to grade.

An ACHE is a device for rejecting heat from a fluid · Optional louvers for process outlet temperature directly to ambient air. This is in contrast to rejecting control. heat to water and then rejecting it to air, as with a shell and tube heat exchanger and a wet cooling · Optional recirculation ducts and chambers for tower system. protection against freezing or solidification of high pour point fluids in cold weather. The obvious advantage of an ACHE is that it does not require water, which means that plants requiring large · Optional variable pitch fan hub for temperature cooling capacities need not be located near a supply control and power savings. of cooling water. Tube Bundle An ACHE may be as small as an automobile radiator A tube bundle is an assembly of tubes, headers, side or large enough to reject the heat of turbine exhaust frames, and tube supports as shown in Figure 2. steam condensation from a 1,200 MW power plant – Usually the tube surface exposed to the passage of air which would require 42 modules, each 90 feet wide has extended surface in the form of fins to by 180 feet long and served by two 60-foot diameter compensate for the low heat transfer rate of air at fans driven by 500-horsepower motors. atmospheric pressure and at a low enough velocity for Components reasonable fan power consumption.

An ACHE consists of the following components: The prime tube is usually round and of any metal suitable for the process, due consideration being given (See Figure 1): to corrosion, pressure, and temperature limitations. Fins are helical or plate type, and are usually of · One or more bundles of heat transfer surface. aluminum for reasons of good thermal conductivity and economy of fabrication. Steel fins are used for · An air-moving device, such as a fan, blower, very high temperature applications. or stack.

· Unless it is natural draft, a driver and power transmission to mechanically rotate the fan or blower.

1 Typical Components of an Air-Cooled Heat Exchanger Hudson Products Corporation Sugar Land, Texas, USA Figure 1

2 9 3 4 5

6 7 8

INDUCED DRAFT

5 4 3 1

9 7 FORCED DRAFT 8

1. Fan 4. Nozzle 7. Drive Assembly 2. Fan Ring 5. Header 8. Column Support 3. Plenum 6. Tube Bundle 9. Inlet Bell

2 Typical Construction of Tube Bundles with Plug and Cover Plate Headers Hudson Products Corporation Sugar Land, Texas, USA Figure 2

16 9 10 3 5 13 14 1 4

11 2 7 8 6

3 15 16 PLUG HEADER

16 9 10 18 3 13 14 18 17 1 17

6 3 4 15 COVER PLATE HEADER

1. Tube Sheet 10. Side Frame 2. Plug Sheet 11. Tube Spacer 3. Top and Bottom Plates 12. Tube Support Cross-member 4. End Plate 13. Tube Keeper 5. Tube 14. Vent 6. Pass Partition 15. Drain 7. Stiffener 16. Instrument Connection 8. Plug 17. Cover Plate 9. Nozzle 18. Gasket

3 Fins are attached to the tubes in a number of ways: also affords a continued predictable heat transfer and should be used for all coolers operating above 600°F 1) by an extrusion process in which the fins are and below 750°F. The wrap-on footed fin tube can be extruded from the wall of an aluminum tube used below 250°F; however, the bond between the fin that is integrally bonded to the base tube for and the tube will loosen in time and the heat transfer is the full length. not predictable with certainty over the life of the cooler. It is advisable to derate the effectiveness of the 2) by helically wrapping a strip of aluminum to wrap-on tube to allow for this probability. embed it in a pre-cut helical groove and then peening back the edges of the groove against There are many configurations of finned tubes, but the base of the fin to tightly secure it, or manufacturers find it economically practical to limit production to a few standard designs. Tubes are 3) by wrapping on an aluminum strip that is footed manufactured in lengths from 6 to 60 feet and in at the base as it is wrapped on the tube. Figure diameters ranging from 5/8 inch to 6 inches, the most 3 shows a cutaway view of these finned tubes. common being 1 inch. Fins are commonly helical, 7 to 11 fins per inch, 5/16 to 1 inch high, and 0.010 to 0.035 inch thick. The ratio of extended to prime surface varies from 7:1 to 25:1. Bundles are rectangular and typically consist of 2 to 10 rows of finned tubes arranged on triangular pitch. Bundles may be stacked in depths of up to 30 rows to suit unusual services. The tube pitch is usually between 2 and 2.5 tube diameters. Net free area for air flow through bundles is about 50% of face area. Tubes are rolled or welded into the tube sheets of a pair of box headers.

The box header consist of tube sheet, top, bottom, and Figure 3 end plates, and a cover plate that may be welded or bolted on. If the cover is welded on, holes must be Sometimes serrations are cut in the fins. This causes an drilled and threaded opposite each tube for interruption of the air boundary layer, which increases maintenance of the tubes. A plug is screwed into each turbulence which in turn increases the air-side heat hole, and the cover is called the plug sheet. Bolted transfer coefficient with a modest increase in the air- removable cover plates are used for improved access side pressure drop and the fan horsepower. to headers in severe fouling services. Partitions are welded in the headers to establish the tube-side flow The choice of fin types is critical. This choice is pattern, which generates suitable velocities in as near influenced by cost, operating temperatures, and the countercurrent flow as possible for maximum mean atmospheric conditions. Each type has different heat temperature difference. Partitions and stiffeners transfer and pressure drop characteristics. The extruded (partitions with flow openings) also act as structural finned tube affords the best protection of the liner tube stays. Horizontally split headers may be required to from atmospheric corrosion as well as consistent heat accommodate differential tube expansion in services transfer from the initial installation and throughout the having high fluid temperature differences per pass. life of the cooler. This is the preferred tube for Figure 4 illustrates common heat types. operating temperatures up to 600°F. The embedded fin

4 Figure 4 Figure 5 Bundles are usually arranged horizontally with the air entering below and discharging vertically. Axial Flow Fans Occasionally bundles are arranged vertically with the Figure 5 displays the air moving device for an ACHE air passing across horizontally, such as in a natural which is commonly an axial flow, propeller type fan draft tower where the bundles are arranged vertically that either forces the air across the bundles (forced at the periphery of the tower base. Bundles can also be draft) or pulls it across (induced draft). To provide arranged in an “A” or “V” configuration, the principle redundancy in case a mechanical unit fails and to advantage of this being a saving of plot area. The provide the basic control achievable by running one disadvantages are higher horsepower requirements for fan or two, a bundle or set of bundles is usually a given capacity and decreased performance when provided with two fans. winds on exposed sides inhibit air flow. Even distribution of the air across the tube bundle is With practical limits, the longer the tubes and the critical for predictable, uniform heat transfer. This is greater the number of rows, the less the heat transfer achieved by adequate fan coverage and static pressure surface costs per square foot. One or more bundles of loss across the bundle. Good practice is to keep the the same or differing service may be combined in one fan projected area to a minimum of 40% of the unit (bay) with one set of fans. All bundles combined projected face area of the tube bundle and the bundle in a single unit will have the same air-side static static pressure loss at least 3.5 times the velocity pressure loss. Consequently, combined bundles pressure loss through the fan ring. For a two fan unit having different numbers of rows must be designed for this is generally assured if the ratio of tube length to different velocities. bundle width is in the range of 3 to 3.5 and the number of tube rows is held to 4 rows minimum with the net free area for air flow at about 50% of the face area of the bundle.

Fans can very in size from 3 to 60 feet in diameter and can have from 2 to 20 blades. Blades can be made of wood, steel, aluminum, or fiberglass-reinforced plastic, and can be solid or hollow. Blades can have

5 straight sides or be contoured. The more efficient type drives are used with very large motors and fan has a wide chord near the center and tapers to a diameters. Fan speed is set by using a proper narrow chord at the tip, with a slight twist. The twist combination of sprocket or sheave sizes with timing and taper compensate for the slower velocity of the belts or V-belts, and by selecting a proper reduction blade nearer the center to produce a uniform, efficient ratio with gears. Fan tip speed should not be above air velocity profile. 12,000 feet per minute for mechanical reasons, and may be reduced to obtain lower noise levels. Motor Fans may have fixed or adjustable pitch blades. Except and fan speed is sometimes controlled with variable for small diameters (less than 5 feet), most ACHEs frequency drives. Figure 6 provides a breakdown of have adjustable pitch blades. Adjustable pitch fans are the mechanical equipment. manufactured in two types. One is manually adjustable (with the fans off) and the other is automatically adjustable (while running). Most automatically adjustable pitch fans change their pitch by means of pneumatically actuated diaphragm working against large springs inside the hub.

Plenum

The air plenum is an enclosure that provides for the smooth flow of air between the fan and bundle. Plenums can be box type or slopesided type. The slopesided type gives the best distribution of air over the bundles, but is almost exclusively used with induced draft because hanging a machinery mount Figure 6 from a slopesided forced draft plenum presents structural difficulties. Structure

Mechanical Equipment The structure consists of columns, braces, and cross beams that support the exchanger at a sufficient Fans may be driven by electric motors, steam turbines, elevation above grade to allow the necessary volume gas or gasoline engines, or hydraulic motors. The of air to enter below at an approach velocity low overwhelming choice is the electric motor. Hydraulic enough to allow unimpeded fan performance and to motors are sometimes used when power from an prevent unwanted recirculation of hot air. To conserve electric utility is unavailable. Hydraulic motors also ground space in oil refineries and chemical plants, provide variable speed control, but have low ACHEs are usually mounted above, and supported by, efficiencies. pipe racks with other equipment occupying the space underneath the pipe rack. ACHE structures are The most popular speed reducer is the high-torque designed for appropriate wind, snow, seismic, piping, positive type belt drive, which uses sprockets that dead and live loads. mesh with the timing belt cogs. They are used with motors up to 50 or 60 horsepower, and with fans up to about 18 feet in diameter. Banded V-belts are still often used in small to medium sized fans, and gear

6 Comparison of Induced and Forced Draft Units temperatures may occur during fan-off or low air flow operation.

3) Fans are less accessible for maintenance, and maintenance may have to be done in the hot air generated by natural convection.

4) Plenums must be removed to replace bundles.

Forced Draft

Advantages:

1) Possibly lower horsepower requirements if Figure 7 the effluent air is very hot. (Horsepower varies inversely with the absolute temperature.) Induced Draft 2) Better accessibility of fans and upper bearings Advantages: for maintenance.

1) Better distribution of air across the bundle. 3) Better accessibility of bundles for replacement.

2) Less possibility of hot effluent air recirculating Accommodates higher process inlet temperatures. into the intake. The hot air is discharged upward at approximately 2.5 times the intake velocity, Disadvantages: or about 1,500 feet per minute. Less uniform distribution of air over the bundle. 3) Better process control and stability because the plenum covers 60% of the bundle face area, Increased possibility of hot air recirculation, resulting reducing the effects of sun, rain, and hail. from low discharge velocity from the bundles, high intake velocity to the fan ring, and no stack. 4) Increased capacity in the fan-off or fan failure condition, since the natural draft stack effect is Low natural draft capability of fan failure. much greater. Complete exposure of the finned tubes to sun, rain, Disadvantages and Limitations: and hail, which results in poor process control and stability. 1) Possibly higher horsepower requirements if the effluent air is very hot. In most cases, the advantages of induced draft design outweigh the disadvantages. 2) Effluent air temperature should be limited to 220°F to prevent damage to fan blades, bearings, or other mechanical equipment in the hot air stream. When the process inlet temperature exceeds 350°F, forced draft design should be considered because high effluent air

7 II. THERMAL DESIGN Q= U • A • (T – t) mean Where

(T – t) mean = CMTD = LMTD • F

(T 1 – t 2) –(T 2 – t 1) • F (T 1 – t 2) 1n [](T 2 –t1)

F is a factor that corrects the log mean temperature difference for any deviation from true counter-current flow. In ACHEs, the air flows substantially unmixed upward across the bundles and the process fluid can Figure 8 flow back and forth and downward as directed by the pass arrangement. With four or more downward There are more parameters to be considered in the passes, the flow is considered counter-current; and, so thermal design of ACHEs than for shell and tube the factor “F” is 1.0. The correction factors for one, exchangers (see Figure 8). ACHEs are subject to wide two, and three passes, given by Figures 9 – 11, were variety of constantly changing climatic conditions calculated from the effectiveness values developed by which pose problems of control not encountered with Stevens, Fernandez, and Wolf 1 for the appropriate shell and tube exchangers. Designers must achieve counter-cross flow arrays. an economic balance between the cost of electrical power for the fans and the initial capital expenditure As is apparent, initially neither the area nor the overall for the equipment. A decision must be made as to heat transfer rate nor the effluent air temperatures are what ambient air temperature should be used for known. The traditional approach in the design of design. Air flow rate and exhaust temperature is ACHEs entailed an iterative trial and error procedure initially unknown and can be varied in the design both on the CMTD and the transfer rate until the area stage by varying the number of tube rows and thus satisfied both. Specifically, an air rise was assumed, varying the face area. the CMTD was calculated, an overall heat transfer coefficient was assumed, and an exchanger size was Because the number of tube rows, the face area, the selected with the expected necessary area. An air face velocity, and the geometry of the surface can appropriate face velocity was then used to calculate all be varied, it is possible to generate many solutions an effluent air temperature, and the process was to a given thermal problem. However, there is repeated until the assumed effluent air temperature obviously an optimum solution in terms of capital and matched the calculated value. The individual operating costs. coefficients and the overall coefficient were then The basic heat transfer relationships that apply to calculated, and the whole process was repeated until shell and tube exchangers also apply to ACHEs. the calculated “U” and CMTD were sufficiently close The fundamental relation is the Fourier equation: to the assumed values.

8 However, there is another method that eliminates trial We define air flow in terms of standard cubic feet per and error on the CMTD and leaves only the trial and minute (scfm) as the product of the effective width and error on the tube-side film coefficient. The following length of the exchanger in feet, and the face velocity discussion presents the Ntu Method described by Kays (FV) in standard feet per minute (sfm). For any ACHE and London in Compact Heat Exchangers 2, as applied service, it will not necessarily be apparent at the to ACHES. design stage whether the air or the hot tube-side fluid will have the minimum heat capacity rate, since the The following are definitions based on Compact Heat mass flow rate of the air is unknown. The two cases Exchangers : presented below will cover both design situations.

1.Hot fluid heat capacity rate = C h = C tube

Q = (Mc p)tube = T1 – T 2

2. Cold fluid heat capacity rate = C c = C air

Q = (Mc p)air = t2 – t 1

A•U 3. Number of heat transfer units = Ntu = Cmin

Cmin 4. Heat capacity rate ratio = R = Cmax

5. ACHE heat transfer effectiveness = E

Ch (T 1 –T2) Cc (t 2 –t1) E = = Cmin (T 1 – t 1) Cmin (T 1 – t 1)

See pages i – iii for additional definitions and nomenclature.

C F R O T C A F N O I T C E R R O 0 . 9 8 7 6 1 . . . . 0 . 1 T

L 2 . T 0 9 . U S R O D

T E

4 T

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S O

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A O H 8

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C T F 0 1 O D U B T s , M s s a

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T / 1 U 9 • s O r e Y

r A 5

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g E t 1 i c B a 5 r F . r U a o T F 0 p L . r 2 A o n C I C P o i s Y t t T c 4 c 5 u . . e d 2 r o r r P o n 0 C o . s 3 d 3 u D . H E T D E E I S M D D 0 E I I 1 . 1 E t S S t D 4 B I – E R – S U I 1 2 B T T t A R U I E 0 E 2 T . A . R R 5 = E E U r U R R T T 0 U . A U A T 6 T R R A E A E R : P R P E E E 0 . M P M R P E 8 E U M T M 1 T . T E E R T T T A 2 T 1 E E L t T T L T L C E – T – E T L 2 N 1 L U t U E T N N I O I O M = = = = = O 1 2 1 2 R t T t T N 0

0 9 8 7 6 5

......

F R O T C A F N O I T C E R R O 1 C

F R O T C A F N O I T C E R R O C 0 . 9 8 7 6 1 . . . . 0 . 1 T E T L

E T

L .

U 0 N I O 9 . W S O R

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T 0 S X C I S A M O F R N N C U O

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0 9 8 7 6 5

......

F R O T C A F N O I T C E R R O 1 C

F R O T C A F N O I T C E R R O C 0 . 9 8 7 6 1 . . . . 0 .

. 0

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C 5 . o 4

. s 2 i t t c u c d e o r r r 3 P o n 3 o . C s E d D u E E I D S D D H E I I T 1 4 1 E t S S t D B I – E R – S M U I 1 2 B T T t A R U I E E 2 T 5 A . R R = E E U r U R R T T U A U A T T R R A E A E 7 R : P R P E E E M P M R P E E U M T 0 M 1 T . T E E 1 T T T A 2 T 1 E E L t T T L T L C E R – T – E T L 2 N 1 L U t U E T N N I O I O M = = = = = O 1 2 1 2 R T t t T N 0

0 9 8 7 6 5

......

F R O T C A F N O I T C E R R O 1 C

12 CASE I. C min = C air = C cold T1 – T 2 = T1 – t1 1. From Definition 4 above:

T1 – T 2 scfm • 1.08 If we let = Z, then T1 –t1 Cmin Cair Q R = = = C C max hot [](T 1 –T2) Z = E • R for Case I From Definition 3 above: FV • L • W • 1.08 A•U A•U n • N • a • W • L • U Q = Ntu = = = Cmin Cair 1.08 • W • L • FV [](T 1 –T2)

FV • L • W • 1.08 • (T –T ) n • N • a = 1 2 = Q 1.08 • FV • (r i + r air + r f + r m)

Note: 1.08=0.075 lb/ft 3 • 60 min/hr n • N • a Let k = · 0.24 Btu/(lb • °F) 1.08 • FV • (r i + r air + r f + r m) 2. From Definition 5 above and by Chot k=Ntu for Case 1 substitution for : Cair We can plot the expression for Case I with E • R and Ntu as coordinates and R as the parameter. Knowing Ch (T 1–T2) that Z = E • R and that k = Ntu, we can find R on the E = plot. Cmin (T 1–t1) From Equation 1, Case I: Q = FV • L • W • 1.08 • (T –T ) FV • L • W • 1.08 • (T –t ) 1 2 1 1 R = Q Multiplying Equation 1 by Equation 2: Q • R FV • L • W • 1.08 • (T 1–T2) W = 3. ER = 1.08 • FV • L • (T 1–T2) Q

(T 1 – T 2) Q t2 = + t 1 • R FV • L • W • 1.08 • (T 1–t1)

13 CASE II. C min = C tube = C hot T1 – T 2 and t 2 = + t 1 R From Definition 5 above:

Ch (T 1 – T 2) T1 – T 2 It can be shown that E, Ntu, and R can be related for E = = = Z any flow arrangement. For counter-current flow, the Cmin (T 1 –t1) T1 –t1 expression is:

From Definition 4 above: 1 – e –Ntu • (1 –R) E = –Ntu • (1 –R) Cmin Chot 1 – R • e R = = Cmax Cair Figures 12-15 show plots relating these variables for the following cases: Q = 1. Cross flow with both fluids unmixed (one-pass FV • L • W • 1.08 • (T 1–T2) ACHE, Figure 12).

2. Two-pass counter cross flow with both n • N • a • W • L Ntu = fluids unmixed in each pass but with the hot Q fluid mixed between passes (two-pass ACHE, •(r + r + r + r ) i air f m Figure 13). [](T 1 –T2) 3. Three-pass counter cross flow with both fluids n • N • a unmixed in each pass but with the hot fluid R • Ntu = mixed between passes (three-pass ACHE, 1.08 • FV • (r i + r air + r f + r m) figure 14).

k = R • Ntu for Case II 4. Counter flow (ACHE with four or more passes, Figure 15).

CASE I or II. Selection Criteria The case of the isothermal exchanger is much simpler. As R goes to zero, the equation for effectiveness We can plot the expression with E and R • Ntu as reduces to: coordinates and R as the parameter on the same graph E = 1 – e –Ntu as Case I, with R = 1 common to both plots. And Ntu is as in Case I for C min = C air : For values of R above the line R = 1: n • N • a Q Ntu = W = 1.08 • FV • (r i + r air + r f + r m) 1.08 • R• (T 1–T2) • FV • L Q Q and t 2 = R • (T 1 – T 2) + t 1 E = = Q 1.08 • (T 1–t1) • FV • L • E For values of R below the line R = 1: max

Q • R Q W = W = 1.08 • (T 1–T2) FV • L L • W • FV • 1.08 • (T 1–t1)

14 0 0 R . . • 1 1 1 1 V t R t F V = S = • p + • V + F S c p ) R ) F • R V 8 2 2 • c A • F 0 E 8 T T • Q P . • W 8 0 M V – – • R Q 1 ) . 0 8 ) • M O 1 1 . 1 1 2 O R 0 ) L ( T T 1 . R T 2 B ( ( E 1 E T –

= B R

1 – 1 N 1 2 = t – T =

R T 1 = t R

( O T

2 A = R • E = Z A

( t

2 1 T – O T O = F F 0 . 5 8 . 4 6 . 4 4 . 4 A S 2 . U 4 , s s 0 a . s 4 x a e T ) ) 8 P . m m , 3 r r d + + 1 n f f 6 . r r a 3 / + + L i i a a r r r t 4 • • . + + a r 3 r r N N i i g a a a • • r r u ( ( 2 . n n S h 3 V V F F 2 C • • • 0 1 . 8 8 3 0 0 e g n . . r 1 1 o n 8 i u . i t

0 = = 2 g

i 1

z . r F U U i

6 0

R 2 . V V o

• • . 2

F F

S p 0 a a • •

r R

• •

= 3

8 8 r 4 o . .

N N R

0 0 0 2 . . e

• • C 4

1 1

. l =

n n

s 0 2

t R

o 5

. 2 = =

c 0

u u o R

u 6

t t . = 0

d .

N 0 N

C R 2

o . =

= =

r 0 8

. R

k k r

P = 0 8

9 . i

R 1

n 0 0

A R

o =

1 6

s 9 . R .

= 1

d 0

. R

u =

0 4

R 7 .

H = 1

6 = .

0 2

R .

5 1

R 0

= 0

4 1 .

R 8 . .

2 6

. .

= R

1 . .

= R 2 . 0 5 0 5 0 5 0 0 5 0 5 0 5 0 5 0 5 0 5 0 5

7 7 6 6 5 0 9 9 8 8 5 4 4 3 3 2 2 1 1 0 ......

1 1 1 t – T

= E = Z

2 1 T – T

15 0 0 R . . • 1 1 1 1 V S t R t F V E = = • p + • V + F S c p ) R ) F • R V S 8 2 2 • c • F 0 E 8 T T A • Q . • W 8 0 M V – P – • R Q 1 . 0 8 ) • M O 1 ) 1 . 1 2 O R 0 ) L T T 2 1 . R T 2 B ( ( ( E 1 T –

A =

B R

1 O – 1 1 2 = t – T =

R T 1 = t R

( W T

2 A = R • E = Z A

( t

T F

2 1 T – O T O = F F 0 . 5 8 . 4 6 . 4 4 . 4 A S 2 . U 4 , s s 0 a . s 4 x a e T ) ) 8 P . m m , 3 r r d + + 2 n f f 6 . r r a 3 / + + L i i a a r r r t 4 • • . + + a r 3 r r N N i i g a a a • • r r u ( ( 2 . n n S h 3 V V F F 3 C • • • 0 1 . 8 8 3 0 0 e g n . . r 1 1 o n 8 i u . i t

0 = = 2 g

i 1

z . r F U U i

6 0

R .

V V

o 2

• • . 2 F F

S = p

a a 0 • •

• •

= 3

8 8 r 4 o . . N N

0 0

R 0 2 . .

• • C 4

1 1

. l = n n

s 0 2

. t R

o 5

2 = = =

c 0

u u o R

t t

= 0

d . .

N N

C 0 2 R

o 7

= . =

r = 0

k k

r 8

P R 8

= . i

0 1

n R .

A 0

o 0 .

R 6

s = . 1

9 1

d R =

8 R

. 4

H R

7 = .

0 R

= 2

. .

0 R 1

0 0

= 1

R .

3 . .

R 6

2 . .

= R

. 1

= R 2 . 0 5 0 5 0 5 0 0 5 0 5 0 5 0 5 0 5 0 5 0 5

7 6 6 5 5 0 9 9 8 8 7 4 4 3 3 2 2 1 1 0 ......

1 1 1 t – T

= E = Z

2 1 T – T

16 0 0 R . . S • 1 1 1 1 V t E R t F V = = • S p + • V + F S c p ) R ) F • R V 8 2 2 • c A • F 0 E 8 T T • Q . P • W 8 0 M V – – • R Q 1 . ) 0 8 ) • M O 1 1 . 1 2 O 3 R 0 ) L T ( T 1 . R T 2 B ( ( E 1 T E –

A =

B R

1 E – 1 1 2 = t – T =

R T 1 = t R R

( T 2 A = R • E = Z A

H ( t R

2 T 1 T – O T O = F F 0 . 5 8 . 4 6 . 4 4 . 4 A S 2 . U 4 , s s 0 a . s 4 x a e T ) ) 8 P . m m , 3 r r d + + 3 n f f 6 . r r a 3 / + + L i i a a r r r t 4 • • . + + a r 3 r r N N i i g a a a • • r r u ( ( 2 . n n S h 3 V V F F 4 C • • • 0 1 . 8 8 3 0 0 e g n . . r 1 1 o n 8 i u . i t = = 2 g a i z r F U U i 6

. 1 V V

0 o

• • . 2 F F

S 2

p 0

a a . • • r

• •

R 3

8 8 r 4 o .

. N = N

0 0 R

2 0

. . 4

e • •

. R

1 1

l =

n 5 n

s . 2

t R

= o 0

6 2 = =

c .

u 0 u o

u 7

t t .

= 0

d 8 . 0 N . N

C 2

o 0

= 9 =

r .

k 0 k r

P 0 8 R

. . i =

1 1

n R

= A 9

o .

6 0

s R .

8 1 =

d .

u R

7 4

. .

H R 0 1

6 = .

R 2

5 1 .

R 0

R . 4

. 1 0

3 .

R 6

. 2

= R

. 1

= R 2 . 0 5 0 5 0 5 0 0 5 0 5 0 5 0 5 0 5 0 5 0 5

7 7 6 6 5 0 9 9 8 8 5 4 4 3 3 2 2 1 1 0 ......

1 1 1 t – T

= E = Z

2 1 T – T

17 S E S 0 0 R S . . • 1 1 1 1 A V t R t P F V = = • p + • V + F E c p ) R ) F • R V 8 2 2 R • c • F 0 E 8 T T • Q . • O W 8 0 M V – – • R Q 1 . 0 8 ) • M O 1 M 1 . 1 2 O R 0 ) L T T 1 . R T 2 B ( R ( E 1 T –

A = O

B R

1 – 1 1 2 ) = t – T =

R T 1 = t R

4 ( T 2 A = R • E = ( Z A

( t R

F F

2 R 1 T – O T O = F U F O F 0 . 5 8 . 4 6 . 4 4 . 4 A S 2 . U 4 , s s 0 a . s 4 x a e T ) ) 8 P . m m , 3 r r d + + 4 n f f 6 . r r a 3 / + + L i i a a r r r t 4 • • . + + a r 3 r r N N i i g a a a • • r r u ( ( 2 . n n S h 3 V V F F 5 C • • • 0 1 . 8 8 3 0 0 e g n . . r 1 1 o n 8 i u . i t = = 2 g a i

z r 1

F . U U i 6

0 2 .

V V

0 • • 3 2

F F

. = 0 p

a a =

• • 0 r

• •

R =

8 8

r . 4 o =

N 5 N

0 0 0

. . .

e • • C

1 = 1

n 6 n

= . 2 R . t

0 2

= = R

c 7 .

u u o 0

t t

8 0 = d . .

N N

C 2 0

o R

= = 9

r = .

k 0 k r 0 R

P . 8

. i

= 1 1

n 9

R =

A .

o 0

R 6

. = 8

. 1

d R 0

u = 7

. 4

. R

H 0 1

. R

0 2

5 1 .

R 0

R . 4

. 1 0

3 8

. .

= R

6 2 .

= R

1 .

= R 2 . 0 5 0 5 0 5 0 5 0 5 0 5 0 5 0 5 0 5 0 0 5

0 7 6 6 9 9 8 8 7 5 5 4 4 3 3 2 2 1 1 0 ......

1 1 1 t – T

= E = Z

2 1 T – T

18 Application of Design Method TABLE 1 100 For any ACHE service at the design stage, the given Z • U Rows FV (ft/min) information includes the process terminal 0.4 4 650 temperatures, the heat load, and the air ambient 0.5 5 600 temperature as well as desirable tube dimensions. The 0.7 6 550 Ntu approach to design determines the optimum 0.8 to 1.0 8 to 10 400 to 450 values for the face area of the bundle and the airside outlet temperature. Using this data, a cooler design is TABLE 2 selected which must be checked rigorously, but the selection most likely will be close to the final best TYPICAL HEAT TRANSFER COEFFICIENTS FOR design for a given service. AIR-COOLED HEAT EXCHANGERS Condensing service U The value of Z can be calculated from the given data, Amine reactivator ...... 100 – 120 and the overall heat transfer coefficient estimated from values shown in Table 2. This allows the determination Ammonia ...... 105 – 125 of the number of tube rows and face velocity Refrigerant 12 ...... 75 – 90 corresponding to the value of Z • 100/U in Table 1. Heavy naphtha ...... 70 – 90 The value of k or Ntu is calculated and, using the Light gasoline ...... 95 assumed number of tube passes, R is read using the Light hydrocarbons ...... 95 – 105 appropriate curve in Figures 12 through 15. This value of R is applied in the appropriate equations to predict Light naphtha ...... 80 – 100 values of FA and the air outlet temperature. This Reactor effluent Platformers, design can be assumed to be accurate enough to be Hydroformers, Rexformers ...... 80 – 100 used for estimating purposes. An example of this Steam (0 – 20 psig) ...... 135 – 200 process is shown in the sample problem on the following page. Gas cooling service

This selection must be checked rigorously to create Air or flue gas @ 50 psig the final design. This is done by application of heat (∆P = 1 psi) ...... 10 transfer and pressure drop correlations that have been Air or flue gas @ 100 psig developed empirically and confirmed by (∆P = 2 psi) ...... 20 experimentation and observation of air cooler Air or flue gas @ 100 psig performance. Many of these correlations are generally (∆P = 5 psi) ...... 30 known. Hudson Products applies criteria that have Ammonia reactor stream ...... 90 – 110 been developed over the past 40 years of experimentation and observation. The influence of the Hydrocarbon gasses @ 15 – 50 psig various parameters that affect both heat transfer and (∆P = 1 psi) ...... 30 – 40 pressure drop are continually being updated by Hydrocarbon gasses @ 50 – 250 psig Hudson Products to assure the application of the latest (∆P = 3 psi) ...... 50 – 60 techniques in the design of air-cooled heat Hydrocarbon gasses @ 250 – 1500 psig exchangers. (∆P = 5 psi) ...... 70 – 90

19 Liquid cooling service From Table 1, we see that proper number of tube rows Engine jacket water ...... 130 – 155 is 6, and that the face velocity should be about 550 sfm. From Table 2, the overall heat transfer coefficient should Fuel Oil ...... 20 – 30 be about 90 Btu/(hr • ft 2 • °F). We can then calculate k: Hydroformer and Platformer liquids ...... 85 n = Tubes per row, per foot of width Light gas oil ...... 70 – 90 12 12 Light hydrocarbons ...... 90 – 120 = = = 4.8 pitch 2.5 Light naphtha ...... 90 Process Water ...... 120 – 145 π Residuum ...... 10 – 20 a = • OD = 0.2618 Tar ...... 5 – 10 12

Coefficients are based on outside bare tube surface for n • N • a 1-inch OD tubes with 10 plain extruded aluminum k = fins per inch, 5/8 inch high, 21.2:1 surface ratio. 1.08 • FV • (1/U)

Sample Problem 4.8 • 6 • 0.2618 = = 1.1424 Cool 273,000 lb/hr of light hydrocarbon liquid from 1.08 • 550 • (1/90) 250°F to 150°F with 100°F ambient air, and with DPallowable = 5 psi, fouling = 0.001, elevation = sea We will assume three tube-side passes, and find the level, and properties at the average tube-side heat capacity ratio R from Figure 14 to be 0.70 and temperature of 200°F as follows: above the R = 1.0 line, so the R term in FA is in the denominator. We can then calculate the face area: cp = 0.55 Btu/lb • °F Q ki = 0.055 Btu/(hr • ft • °F) FA = = 360.8 FV • 1.08 • R • (T 1–T2) µ = 0.51 centipoise = 1.234 lb/ft • hr Considering the given 32 foot long tubes, The specified tube material is 0.085 inch MW (0.093 inch AW) x 32 foot long carbon steel, and we select 1- 360.8 inch OD tubes with 10 extruded fins per inch, 5/8 Width = = 11.28 ft 32 inch high, with 2.5 inch transverse tube pitch.

We calculate Q = 273,000 • 0.55 • (250 – 150) = 11.28 ft • 12 in/ft • 6 rows 15,015,000 Btu/hr, and then use the Ntu Method to The tube count = = 325 2.5 inch tube pitch select an exchanger size to try.

The width of the structural components including air 250 – 150 100 seals adds about 6 inches, and we round the size up Z = • = 0.741 to the nearest standard size, namely 12 feet, with 250 – 100 90 [] 336 tubes.

20 Fan Selection – Horsepower Requirements mechanical efficiency. The value of driver output horsepower from the equation above must be divided The fan diameter must assure that the area occupied by the motor efficiency to determine input power. by the fan is at least 40 percent of the bundle face area. The fan diameter must be 6 inches less than the For estimating purposes, refer to Figure 16 to bundle width. Fan performance curves are used to approximate the horsepower requirement. This chart select the optimum number of blades and pitch angle plots bare tube surface divided by horsepower versus as well as the horsepower. tube bundle depth for the normal range of velocities. Applying the above criteria to our sample problem, To calculate the required horsepower we determine that we must use two 10-foot diameter for the fan driver: fans to have 40% of the bundle face area. Entering Motor Shaft Horsepower = Figure 16, we find that for a 6-row bundle, the area/horsepower is between 68 and 92 square feet of Actual ft 3/min (at fan) • Total Pressure Loss (inches water) bare tube surface. If we use an average value of 80, 6356 • Fan (System) Efficiency • Speed Reducer Efficiency the horsepower requirement for each fan is (336 • .2618 • 32) / (2 • 80) = 17.5 horsepower at The actual volume at the fan is calculated by maximum design ambient temperature. Power multiplying the standard volume of air (scfm) by the consumption must be calculated for the coldest density of standard air (0.075 lb/ft 3) divided by the expected ambient temperature, since at a fixed fan density of the air at the fan. From this relationship, it blade angle, fan horsepower consumption is inversely can be seen that the ratio of the fan horsepower proportional to the absolute temperature. The power required for a forced draft unit to that required for an required for this minimum ambient temperature will induced draft unit is approximately equal to the ratio set the required motor size. of the exit air density to the inlet air density, which is in III. PERFORMANCE CONTROL OF ACHEs turn equal to the ratio of absolute air temperatures (t 1 + 460) / (t + 460). The total pressure difference across 2 In addition to the fact that the process flow rate, the fan is equal to the sum of the velocity pressure for composition, and inlet temperature of the fluid may the selected fan diameter, the static pressure loss vary from the design conditions, the ambient air through the bundle, (which is determined from the temperature varies throughout a 24-hour day and from equipment manufacturer’s test data for a given fin type day to day. Since air coolers are designed for and tube spacing), and other losses in the aerodynamic maximum conditions, some form of control is system. Fan diameters are selected to give good air necessary when overcooling of the process fluid is distribution and usually result in velocity pressures of detrimental, or when saving fan power is desired. approximately 0.1 inch of water. Although control could be accomplished using by- The design of the fan, the air plenum chamber, and passing of process fluid, this is rarely done, and the the fan housing, (in particular fan tip clearance), can usual method is air flow control. materially affect system efficiency, which is always lower than on fan curves based on idealized wind tunnel tests. Industrial axial flow fans in properly designed ACHEs have fan (system) efficiencies of approximately 75%, based on total pressure. Poorly designed ACHEs may have system efficiencies as low as 40%. Speed reducers usually have about 95%

21 0 0 2 7 N 2 h 4 1 G 1 t I S 0 p E 9 1 e D 0 1 R 4 1 1 D O F 0 8 e T 1 l O 1 0 3 N d 1 1 – n 0 Y 7 L u 1 6 N 2 9 B O 1 f G 0 6 N I 1 O 5 T 1 8 A 1 n M I 0 T 5 o 1 S i E 7 t 7 9 R c A O S 0 n F 4 U 1 u 8 , 6 8 s F a 0 x 3 A e 1 T 0 s 5 , 8 d A 0 n 2 a r 1 L 5 e 4 7 r R a E w 0 g 1 W u . o 1 T S 6 O F p 6 3 . P 6 • e Q E 1 0 S s S 0 } , e 1 n r E R r o C i O u o A t s g F H a i } d R a 0 H r / F n e 9 U s o r u S A a w p o n E o r e p r B a c , o t e U a F 0 F h T C b . 8 g u E t i s t t f R t e , i . A c h q w t B s n u t i p r 0 = d e n e 7 U o A p U D r r P e n 0 o P 6 s d e u c 0 H a 5 f r u 0 S 4 d n 0 3 A t h 0 2 g i e 0 W 1 t i n

9 8 7 6 5 4 3 2 1 4 3 2 1 0 U S W O R E B U T N I H T P E D E L D N U 1 1 1 1 1 B

22 Varying Air Flow Extreme Case Controls

Varying air flow can be accomplished by: Extreme case controls include: (See Figure 17) 1. Adjustable louvers on top of bundles. 1. Internal Recirculation 2. Two-speed fan motors. By using one fixed-pitch fan blowing upward and one AUTO-VARIABLE pitch fan, which is capable 3. Fan shut-off in sequence for multifan units. of negative pitch and thus of blowing air ® 4. AUTO-VARIABLE fans. downward, it is possible to temper the air to the 5. Variable frequency for fan motor control. coldest portion of the tubes and thus prevent freezing. Normally forced draft units have the Louvers operate by creating an adjustable restriction negative pitch fan at the outlet end, while induced to air flow and therefore do not save energy when air draft units have the positive pitch fan at the outlet flow is reduced. In fact, louvers impose a permanent end. In hot weather, both fans can blow upward. energy loss, even in the open position.

Two-speed motors, AUTO-VARIABLE fans, and 2. External Recirculation variable frequency fan motor control do save power This is a more positive way of tempering coolant when air flow is reduced. In temperate climates, as air, but is practical only with forced draft units. much as 67% of the design power may be saved over Hot exhaust air exits the bundle and enters a top the course of a year with AUTO-VARIABLE pitch fans. plenum covered by a louver. When no circulation AUTO-VARIABLE hubs will thus pay back their is required, the top louver is wide open, and the additional cost in about one year or less. heated air exits through it. When the top louver is partially closed, some of the hot air is diverted to a duct, through which it flows downward and back into the fan intake, mixing with some cold ambient air. An averaging air temperature sensor below the bundle controls the amount of recirculated air, and thus the average air intake temperature, by varying the louver opening.

3. Co-current Flow Four high pour-point streams, it is often advisable to ensure a high tube wall temperature by arranging the flow co-currently, so that the high inlet temperature process fluid is in contact with Figure 17 the coldest air and the low temperature outlet Both louvers and AUTO-VARIABLE fans may be process fluid is in contact with the warmed air. operated automatically through an instrument that senses temperature or pressure in the outlet header. 4. Auxiliary Heating Coils – Steam or Glycol For extreme cases of temperature control, such as Heating coils are placed directly under bundles. prevention of freezing in cold climates in winter, or Closing a louver on top of a bundle will allow the prevention of solidification of high pour-point or high heating coil to warm the bundle or keep it warm in melting point materials, more sophisticated designs freezing weather, so that on start-up or shut-down are available. the material in the bundle will not freeze or

23 solidify. Heating coils are also occasionally used to transition region. However, this usually results in temper very cold air to the bundles while the fan is pressure drops of 30 to 100 psi. In view of the operating and the exhaust louver is open. disadvantages of designing for laminar flow, this increased pressure drop is normally economically IV. NOISE CONTROL justifiable because the increase in the operating and capital cost of the pump is small compared with the In recent years concerns about industrial noise have decrease in the cost of the turbulent exchanger. grown. Since ACHEs were not originally one of the serious sources, it has only been after the abatement The biggest problem with laminar flow in tubes is that of the more serious contributors that attention has the flow is inherently unstable. The reasons for this can focused on ACHEs. be demonstrated by a comparison of pressure drop and heat transfer coefficient for turbulent versus laminar ACHE noise is mostly generated by fan blade vortex flow, as functions of viscosity (µ) and mass velocity (G): shedding and air turbulence. Other contributors are the speed reducer (high torque drives or gears) and the Delta P Heat Transfer motor. The noise is generally broad band, except for Flow Type Function Function occasional narrow band noise produced by the motor or speed reducer, or by interaction between these Turbulent µ0.2 , G 1.8 µ-0.47 , G 0.8 sources and the ACHE structure. Laminar µ1.0 , G 1.0 µ0.0 , G 0.33 The evidence is that for efficient fans at moderate fan tip speeds, this noise is proportional to the third power In an air-cooled heat exchanger, because of imperfect of the fan blade tip speed and to the first power of the air-side flow distribution due to wind, or because of consumed fan horsepower. It is at present quite multiple tube rows per pass, it is likely that the flow practical and usually economical to reduce the sound through some of the tubes in a given pass is cooled pressure level at 3 feet below an ACHE to 85 dB(A); more than that through other tubes. but, below 80 dB(A), noise from the drives predominates and special measures must be taken. With turbulent flow, pressure drop is such a weak function of viscosity (0.2 power) and such a strong V. DESIGN OF ACHEs FOR VISCOUS LIQUIDS function of mass velocity (1.8 power), that the flow in the colder tubes must decrease only slightly in order Film coefficients for laminar flow inside tubes are very for the pressure drop to be the same as that in the low and of the same order of magnitude as film hotter tubes. Also, as the flow slows and the viscosity coefficients for air flowing over the outside of bare increases, the heat transfer coefficient drops tubes. Therefore, there is generally no advantage in significantly, (-0.47 power of viscosity, 0.8 power of using fins on the air side to increase the overall heat G), so the over-cooling is self-correcting. transfer rate since the inside laminar flow coefficient will be controlling. Bare tube bundles with a large With laminar flow, pressure drop is a much stronger number of rows are usual. function of viscosity (1.0 power) and a much weaker function of mass velocity (1.0 power), so the flow in For process fluids with outlet viscosities up to 20 the colder tubes must decrease much more to centipoises, it is possible by using large diameter compensate for the higher viscosity. Viscosity of tubes and high velocities (up to 10 ft./sec) to achieve a heavy hydrocarbons is usually a very strong function Reynolds number at the outlet above the 2,000 critical of temperature, but with laminar flow, the heat transfer Reynolds number, and to keep the flow in the coefficient is independent of viscosity, and only a

24 weak function of mass velocity (0.33 power), so the promoters and designing for a lower velocity. It is self-correction of turbulent flow is absent. then advantageous to use external fins to increase the air-side coefficient also. In addition to the increase in The result is that many of the tubes become heat transfer coefficient, turbulence promoters have virtually plugged, and a few tubes carry most of the the great advantage that the pressure drop is flow. Stability is ultimately achieved in the high proportional to the 1.3 power of mass velocity, and flow tubes as a result of high mass velocity and only to the 0.5 power of viscosity, so that non- increased turbulence, but because so many tubes isothermal flow are much more stable. The simplest carry little flow and contribute little cooling, a and probably the most cost-effective promoters are the concurrent result is high pressure drop and low swirl strips, a flat strip twisted into a helix. performance. The point at which stability is reached depends on the steepness of the viscosity VI. COST versus temperature curve. Fluids with high pour points may completely plug most of an exchanger. The approximate purchase price may be determined from Figure 18, which gives the price per square foot This problem can sometimes be avoided by designing of bare tube surface as a function of the total bare deep bundles to improve air flow distribution. tube surface and the number of tube rows. The prices Bundles should have no more than one row per pass indicated are FOB factory, and do not include freight and should preferably have at least two passes per or export crating charges. The prices are based on 1- row, so that the fluid will be mixed between passes. inch OD X 12 BWG X 32 foot long steel tubes with extruded aluminum fins, fabricated steel headers with When fluid has both a high viscosity and a high pour steel shoulder plugs, 100 psig design pressure, TEFC point, long cooling ranges should be separated into motors, and HTD drives. Price multiplication factors stages. The first exchanger should be designed for are included for different tube materials. turbulent flow, with the outlet temperature high enough to ensure an outlet Reynolds number above 2,000 even It can be seen from these curves that the price per with reduced flow. The lower cooling range can be square foot varies little for installations in excess of accomplished in a serpentine coil (a coil consisting of 7,000 square feet of bare tube surface. It is also tubes or pipes connected by 180° return bends, with a evident that the reduction in unit price as a function of single tube per pass). The low temperature serpentine the number of tube rows becomes progressively less coil should, of course, be protected from freezing by as the number of rows increases. external warm air recirculation ducts. REFERENCES Closed loop tempered water systems are often more 1. Steven, R.A., J. Fernandez, and J. R. Wolf: “Mean Temperature economical, and are just as effective as a serpentine Difference in One, Two and Three-pass Crossflow Heat coil. A shell and tube heat exchanger cools the Exchangers,” Trans, ASME, Paper Nos. 55-A-89 and 55-A99, viscous liquid over its low temperature range on the 1955. shell side. Inhibited water is recirculated between the 2. Kays, William M., and Al. L. London, Compact Heat Exchangers, tube side of the shell and tube and an ACHE, where Third Edition, McGraw-Hill Book Company, New York, 1984. the heat is exhausted to the atmosphere. 3. Seider, E. N. and G. E. Tate, “Heat Transfer and Pressure Drop of Liquids in Tubes.” Ind. Eng. Chem., Vol. 28, 1429-1435, 1936. For viscous fluids which are responsibly clean, such as 4. Kern Donalds Q., Process Heat Transfer, McGraw-Hill Book lube oil, it is possible to increase the tube side Company, New York, 1950. coefficient between four- and tenfold, with no 5. Rohsenow, Warren M., and James P. Hartnett, Handbook of Heat Transfer, McGraw-Hill Book Company, New York, 1973. increase in pressure drop, by inserting turbulence

25 0 0 0 , 0 2 , s e 0 b . u 0 s t r . 4 0 t . o , t S 1 . o 8 5 y C 1 h b m . 1 d t 1 y e C l . F y d p l p d i E b : t e e T l s y e , 0 d l w u s w . u p r l i 0 m o 5 g t e l c , l l D 1 L 3 0 d s n ' u o 0 i r , a . 0 f 2 . e t e 1 m s 3 1 6 e d o , h a s l a x y y n 1 g e e e b b r e u h W b d r l a y y l l a u p d M s t p p r n l i i g n l t t A G e o 0 a l l e n t i e u S e u u s t c t W 0 t s e a a 5 B s m m f r U 0 b B 1 0 , , s n c . 2 t , u s 9 e e , o t 1 . 1 n t r r e r s 4 b l l e y x y u u e r o d e n a . s s b b g i a 1 m p s s e . t x c D a n x t y y . e e n s t l t l , s a r r e e s O u e p p s h p p j i i F r T d " 4 s t t c A d l l u n n n , . 1 0 e 0 s a l u u o a g g 3 s t i i d n n - , q e 0 i e e s s m m t t o P c r n a , , i e e c h S 0 T t p r s s p c e d d a s , g j u e e , p i i i G n L b 4 s s b e b a 2 d e e g r 0 u t W i u p p u f e f r t s t s 3 1 c a , s B - 0 ' 0 ' e s a r s a 5 6 4 0 0 P a m y d e i b g 2 1 2 5 T 4 f a i c x u i s s r r r r r r u r r o w e o o o o o o p r 0 P S k S v l u F F F F F F p r 0 * a 0 p u 0 l 8 S • 1 W C A 0 1 , a e t 0 e n b r 1 o o u i u t S T g T a i f W r F e o 0 r O p O R a 0 r 0 E B o , n l B C 8 0 a U o s t T 3 4 5 6 8 1 i t t o c u T c 0 d n o 0 r u 0 , P F 6 n o A s d s u 0 A H 0 0 , e 4 c i r P 0 t 0 i 0 , n 2 U 0 5

5 0 5 0 5 0 5 0 5 0 0

7 2 2 6 6 5 5 4 4 3 3

7 . t F . q S / s r a l l o D , e c i r P g n i l l e S

26 HUDSON Products Corporation

9660 Grunwald Road Beasley, Texas 77417-8600 Phone: (281) 275-8100 Fax: (281) 275-8211 1-800-634-9160 (24 Hours)

E-Mail: [email protected] Internet Home Page Address: http://www.hudsonproducts.com

Hudson Products Corporation manufactures

• Auto-Variable ® fan hubs • Combin-Aire ® water- and air-cooled heat exchangers • Exact-A-Pitch ® digital protractor for setting fan pitch angles • Fin-Fan ® air-cooled heat exchangers • Solo-Aire ® air-cooled heat exchangers • Hy-Fin ® finned tubes • Tuf-Edge ® erosion resistant leading edge protection • Tuf-Lite ® fan blades • Cofimco fan blades

Hudson, Auto-Variable, Combin-Aire, Exact-A-Pitch, Fin-Fan, Hy-Fin, Tuf-Edge, and Tuf-Lite are registered trademarks of Hudson Products Corporation.