Fin Fan Theory

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SECTION 10

Air-Cooled Exchangers

An air-cooled exchanger is used to cool fluids with ambientair. Several a rticles have been published describing in deta iltheir a pplication and economic ana lysis. (See Bibliography a tth e end of this section.) This section describes the genera ldesign of air-cooled exchangers a nd presents a method of ap-proximat e sizing.

ARRANGEMENT AND MECHANICALDESIGN

Figs. 10-2 a nd 10-3show typical elevat ion and plan views ofhorizonta l air-cooled exchang ers a s comm only used. The bas ic

components are one or more tube sections served by one more axial flow fans, fan drivers, speed reducers, and a n eclosing a nd supporting stru cture.

Air-cooled exchangers a re classed a s forced dra ft w hen thtube section is located on th e discha rge side of the fan , and induced draft when t he tube section is located on the suctiside of the fan .

Advantages of induced draft are:

Bet ter distribution of air a cross the section. Less possibility of the hot effluent air recirculatin

ar ound to the inta ke of the sections. The hot a ir is d

Ai = inside surface of tube, sq f t

Ab = outs ide bare tube surface, sq f t

Ax = outs ide extended surface of tube, sq f t

At = tube inside cross-sectional ar ea, sq in. (see Fi g. 9-25)

AC F M = a c t ua l cu bic f ee t p er min u t e

APF = tota l ex terna l a rea/f t o f fin tube, sq f t/f t

APS F = externa l a rea of f in tube, sq f t/sq f t o f bundle face a rea

AR = area ra t io of f in tube compared to the exter ior a reaof 1 in. OD ba re tube

B = cor r ect ion f a ct or, ps i (s ee Fig . 10-14)

C p = specif ic heat a t a verage temperat ure, Btu/(lb F)

CMTD = corrected mean tempera ture di fference, F

D = fa n d ia met er, ftD i = inside tube diameter, in .

D o = outs ide tube diameter, in .

D R = densi ty ra t io, the ra t io of ac tua l a i r dens i ty to thedensity of dry a ir at 70 F a nd 14.7 psia, 0.0749lb/cu ft (see Fig . 10-16)

f = f ri ct ion fa ct or (s eeFig . 10-12)

F = cor re ct ion f a ct or (s ee Fi g. 10-8)

F a = to t a l f ace a rea o f bundles , sq f t

F p = air pressure drop factor , in . of wa ter per rowof tubes

FAP F = fa n a rea per fa n , f t2/fa n

g = loca l accelera t ion due to grav i ty, f t/s2

G = ma s s ve loci t y, lb /(s q f t s)

G a = air fa ce ma ss velocity, lb/(hr sq ft) of face areaG t = tubeside mass velocity, lb/(sq ft s)

h a = air side film coefficient B tu/(h sq ft F)

h s = shell side film coefficient based on outside tubea rea , B tu/(h sq ft F)

h t = tube s ide fi lm coefficient based on inside tube area ,B tu/(h sq f t F)

J = J fa ct or (seeFig . 10-15)

k = thermal conduct iv it y, B tu/[(hr sq ft F )/ft ]

L = len gt h of tube, ft

L MTD = log mea n t emp er a t u r e d i ff er en ce , F

(see Fig . 9-3)

N = number o f rows of tubes in direct ion of flow

NP = number of tube passes

NR = modif ied Reynolds number, (in lb/(sq ft s cp

N t = n u mb er of t ub es

P = pr es su re d rop, ps iP F = f a n t ot a l pr es su r e, inch es of w a t e r

a = density of a ir , lb/cu f tw = density of wa ter, lb/cu f tP = t em per a tur e r a tio (s ee Fig . 10-8)

Q = h ea t t ra n sf er r ed , B t u /hrd = fouling resistance (fouling factor), (hr ft

2 F/B

r f = fluid film resista nce (reciprocal of film coefficie

rmb = metal res ista nce referred to outs ide bare surfac

rmx = metal res is tance referred to outs ide extendedsu

R = t em per a tur e r a tio (s eeFig . 10-8)

S = s peci fi c g ra v it y (w a t er = 1.0)

t = t em per a tu re a ir sid e, F

T = t em per a t ur e t u be s id e, F

U = overa l l hea t t ra ns fer coef ficient , B tu/(h ft2F

W = m a ss fl ow, lb/h r

Y = correct ion factor, psi/ft (see Fig . 10-14)

= v iscos i ty, cp

w = viscosity a t average tube wa ll temperature, cp = viscosity gradient correct ionSubscripts:

a = a ir side

b = ba r e t ube su rfa ce ba sis

s = shell side

t = tube side

x = ex te nd ed t ub e s ur fa ce b a sis

1 = in let

2 = out let

FIG. 10-1

Nomenclature

10-1
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charged upwa rd a t approximately 212 t imes t he velocityof inta ke, or a bout 1500 ft/min.

Less effect of sun, rain, and hail, since 60%of the facearea of the section is covered.

Increased capa city in the event of fan fa ilure, since thenat ural dra f t s ta ck effect is much great er with induced

dra f t .

Disadvantages of induced draft are:

Higher h orsepower since the fan is locat ed in the hot a ir.

Effluent a ir tempera ture should be limited to 200F, to pre-vent potential damage to fan blades, bearings, V-belts, orother mechanical components in the hot air str eam .

The fan d rive component s ar e less a ccessible for ma int e-na nce, which ma y ha ve to be done in the hot a ir gener-at ed by natura l convection.

For inlet process fluids above 350F, forced dra ft designshould be used; otherwise, fan failure could subject t hefan bla des and bea rings to excessive tempera tures.

Advantages of forced draft are:

Slightly lower horsepower since the fan is in cold air.(Horsepower varies directly as the absolute tempera-ture.)

Bet ter a ccessibility of mechanical components for ma in-tenance. Easily adaptable for warm air recirculation for cold cli-

mates .

The disadvantages of forced draft are:

P oor dist ribution of air over the section.

Gr eat ly increased possibility of hot air recirculation, dueto low discharge velocity from t he sections and absenceof stack.

Low nat ural dra f t capability on fan fa ilure due to smallsta ck effect.

Total exposure of tubes to sun, ra in, an d ha il.

The horizont a l section is the most commonly u sed a ir cooledsection, and generally the most economical. For a fluid withfreezing potential, the tubes should be sloped at least 18 in .per foot t o the outlet hea der. Since in most ca ses there w ill beno problem associat ed wit h freezing, and it is more costly todesign a sloped unit , most coolers a re designed wit h level sec-tions.

Vertical sections a re sometimes used w hen ma ximum dra in-age a nd hea d a re required, such a s for condensing services.

Angled sections, like verti cal sections, a re used for conden s-ing services, allowing positive dra inage. F requently, angle sec-tions are sloped thirty degrees (30) from the horizontal.A-fra mes a re usually sloped sixty degr ees (60) from the h ori-zonta l . SeeFig . 10-4.

Forced draft

DriverDriveassembly

FanFan

ringSupportingstructure

Air plenumchamber

Tube section

Headers

Nozzles

Induced draft

Fan Fan ringAir plenumchamber

Headers

NozzlesDrive

assemblyDriver

TubeSection

FIG. 10-2

Typical Side Elevations of Air Coolers

Baywidth

Baywidth

Unit width

Unit width

Tubelength Tubelength

Tubelength

Tubelength

Two-fan bay with2 tube bundles

Two two-fan bays with6 tube bundles

One-fan bay with3 tube bundles

Two one-fan bays with4 tube bundles

FIG. 10-3

Typical Plan Views of Air Coolers

Non-freeze

Dividedrear header

Tube

bundle

Hot air

Hotair

Exhaust

stream

Cool

air

FIG. 10-4

Angled Section Layout

10-2


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Fa n sizes ra nge from 3 ft t o 28 ft dia meter. However, 14 ftto 16 ft diameter is the largest diameter normally used. Fandrivers ma y be electric motors, steam t urbines, hydra ulic mo-tors, or ga s-ga soline engines. A speed reducer, such as a V-beltdrive or reduction gea r box, is necessary to ma tch th e driveroutput speed to the r elatively slow speed of the a xial flow fan.Fa n tip speeds a re norma lly 12,000 ft/min or less. Gen era lpractice is to use V-belt drives up to about 30 bhp and geardrives at higher power. Individua l driver size is usua lly lim-

ited t o 50 hp.

Two fa n ba ys a re popular, since th is provides a degree ofsafety a gainst fan or driver failure and a lso a met hod of controlby fan st aging . Fan coverage is the rat io of the projected areaof the fan to the face of the section served by the fan. Goodpra ctice is to keep this ra tio above 0.40 w henever possible be-cause higher ra tios improve air distr ibution across the face ofthe tube section. Face area is the plan a rea of the heat t ra nsfersurface ava ilable to air flow at the fa ce of the section.

The heat -tra nsfer device is the t ube section, which is a n a s-sembly of side fra mes, tube supports, hea ders, and fin tubes.Aluminum fins are normally applied to the tubes to providean extended surface on the a ir side, in order to compensate forthe relatively low heat transfer coefficient of the air to thetube. Fin construction types a re tension-wr apped, embedded,extruded, and welded.

Tension-w ra pped is probably t he most common fin t ype usedbecause of economics. Tension w ra pped t ubing is common forcontinuous service with temperat ures below 400F. Extrudedfin is a mechanical bond between a n inner t ube exposed to theprocess and an outer tube or sleeve (usually a luminum) whichis extrud ed into a high fin. Embedded fin is an a luminum orsteel fin grooved into the ba se tube. Embedded fins a re usedin cyclic and high t emperatur e services. Other ty pes of finnedtubes ava ilable are soldered, edge wra pped, and serra ted ten-sion wrapped. Coolers are regularly manufactured in tubelengths from 6 ft to 50 ft a nd in ba y widt hs from 4 ft to 30 ft .Use of longer tubes usua lly results in a less costly design com-

pared t o using shorter t ubes.

Ba se tube diameters are 58 in. to 112 in. OD with fins from12 in. to 1 in. high , spaced from 7 t o 11 per inch, providing a nextended finned surfa ce of 12 to 25 times t he outside surfaceof the base tubing. Tubes are usually a rra nged on trian gularpitch with t he fin tips of adjacent t ubes touching or separa tedby from 116 in. to 14 in. Mat ching of the tube section to the fa nsystem and the heat t ran sfer requirements usually results inthe section having depth of 3 to 8 rows of fin tubes, wit h 4 rowsthe most typical.

A 1-in. OD t ube is the most popular d iamet er, an d th e mostcommon fins are 12 in. or 58 in. high. The da ta presented inFig. 10-11are for 1 in. OD t ubes with 12 in. h igh fin s, 9 fins/in.(designated as 12 x 9) and 58 in. hi gh fi ns, 10 fins/in. (desig -na ted as 58 x 10).

Common materials of construction for headers are fireboxqu a lit y ca rbon ste el, ASTM SA-515-70, SA-516-70. Tubes a regen era lly AS TM S A-214 (ER W), SA-179 (SML S ), ca rbon st eel.Louvers are generally carbon steel, or aluminum w ith carbonsteel construction being th e most genera l an d most economi-cal . Fins are normally a luminum. Both s ta inless and brassalloys have their applications but are more expensive thancarbon steel.

HEADER DESIGNPlug header construction uses a welded box which allow

part ial a ccess to tubes by mea ns of shoulder plugs opposite ttubes. Plug headers are normally used as they are cheaptha n the a lterna te cover plate design. Cover plate header costruction allows tota l access to head er, tube sheet, and tubeThis design is used in hig h fouling, low pressur e service.

Fi g. 10-5 shows typical designs for both plug header ancover plate header.

AIR-SIDE CONTROL

Air-cooled exchangers a re sized to operate a t wa rm (sum

mer) a ir temperatur es. Sea sonal varia tion of the air temperture ca n result in over-cooling which ma y be undesira ble. Owa y to control the a mount of cooling is by varying th e amouof air flowing through the tube section. This can be accomplished by using multiple motors, 2-speed drives, variabspeed motors, louvers on the face of the tube section, or vaable pitch fan s.

Sta ging of fans or fan speeds ma y be adequa te for systemwh ich do not requ ire precise control of process tempera tur epressure. Louvers will provide a full ra nge of air qua ntity cotrol. They ma y be operat ed ma nua lly, or automa tically ope

169

103

1

1311

18

6

17

35 12

4 15

17

1814

Cover plate header

16 93

110

5

2

8

6

3

16 Plug header

13

11

12

414

7

15

FIG. 10-5

Typical Construction of Tube Section with Plug and Cove

Plate Headers

1. Tube sheet 7. St i f fener 13. Tube keeper

2 . P lu g shee t 8. P lu g 14. Ve nt

3. Top and bottom plates 9. Nozzle 15. Drain

4. End plate 10. Side frame 16. Instrument connection

5. Tube 11. Tube spacer 17. Cover plate

6. Pass par t i t ion 12. Tube supportcross-member

1 8. G a s k et

10-3


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at ed by a pneumat ic or electric motor controlled from a remote

temperat ure or pressure controller in th e process strea m. Lou-

vers used with constant speed fans do not reduce fan power

requirements.

Auto-var iable-pitch fans are norma lly provided with pneu-

ma tically operat ed blade pitch adjustment w hich may be con-

trolled from a r emote sensor. Blade pitch is a djusted to provide

the required a mount of air flow t o mainta in the process tem-

peratu re or pressure at the cooler. The required blade a ngledecreases as ambient air temperature drops and this con-

serves fan power. Hydra ulic variable speed drives reduce fan

speed when less air flow is required a nd can a lso conserve fan

power.

A design considerat ion which might be required for sat isfac-

tory process fluid cont rol is co-current flow. In extrem e cases

of high pour point fluids, no a mount of air side contr ol wouldallow satisfactory cooling and prevent freezing. Co-current

flow ha s the coldest a ir cool the hottest process fluid, while the

hott est a ir cools th e coolest process fluid. This is d one in order

to maint ain a high tube wa ll temperatur e. This gives a much

poorer LMTD, but for highly viscous fluids is often the only

wa y t o prevent freezing or unaccepta ble pressure drops. With

a ir coolers, th e most common met hod of accomplishing co-cur-rent flow is t o have the inlet nozzle on the bottom of the hea derwith the pass a rrangement upwa rds . This tota lly reverses the

standa rd design, and ma y cause a problem with dr ainage dur-

ing shut-downs. In ad dition, air side control is necessary wit h

co-current designs.

WARM AIR RECIRCULATION

Extreme va riat ion in air tempera ture, such as encounteredin northern climat es, may require special a ir recirculation fea-tures. These are needed to provide control of process streamtemperat ures, and to prevent freezing of liquid stream s. War mair r ecirculation varies from a sta nda rd cooler wit h one revers-ing fan t o a tota lly enclosed system of aut omatic louvers andfans. These two widely used systems a re termed interna l re-

circulation a nd externa l recirculat ion.

A typical layout for internal recirculation is shown in Fig.10-6. During low am bient operation, the manua l fan continuesto force air th rough the inlet ha lf of the section. The a uto-vari-able fan operates in a reversing mode, an d dra ws hot a ir fromthe upper recirculation cha mber down t hrough the outlet endof the section. Because of the lower recirculation skirt , themanual fan mixes some of the hot air brought down by theaut o-varia ble fan w ith cold outside air a nd th e process repeats.The top exha ust louvers are aut omatically a djusted by a tem-peratur e contr oller sensing t he process fluid strea m. As thefluid temperat ure rises, the louvers are opened. During designambient conditions, the louvers are full open and both fansoperate in a s tanda rd forced draf t mode.

A cooler wit h int erna l recirculation is a compromise betw eenno recirculation a nd fully controlled externa l recirculat ion. Itis cheaper tha n full externa l recirculation, and ha s less sta ticpressure loss during maximum ambient temperature condi-tions. A cooler w ith int ernal recirculation is ea sier to erect, andrequires less plot area tha n a n external recirculat ion design.However, this latt er design is more costly tha n a cooler wit hno recirculation, and cannot provide complete freeze protec-

Without recirculation

Auto-variable fan(slight negative pitch)

Manual fan(on)

ExhaustExhaust Exhaust

Automatic louvers Automatic louvers (partially closed)

upper recirculationchamber

Coil

Manual fan(on)

Auto-variable fan(positive pitch)

Lower recirculation skirt

Minimum

Normal airflow

Recirculated airflow

Normal airflow

Lower recirculation

skirt

Upper recirculationchamber

Coil

With recirculation

FIG. 10-6

Internal Recirculation Design

10-4


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tion. Because there is no control over air intake, and fans alonecannot fully mix air, stratified cold air may contact the section.With the fans off, high wind velocity during low ambient condi-tions could cause excessive cold air to rea ch the section.

A typical la yout for externa l recirculat ion is shown inFig . 10-7.Durin g low a mbient t emperat ure conditions, tw o-speed motors on

low speed, or auto-variable fans at low pitch, are normally used.For this design, the sides of the cooler are closed with manuallouvers. Over both ends, a recirculat ion chamber projects beyondthe section headers, and provides a duct for mixing cold outsideair w ith w arm recirculated a ir. As with the internal recirculationdesign, the top exhaust louvers are controlled by the tempera tur eof the process fluid. However, this design provides for control ofthe inlet a ir tempera ture. As the inlet air louver closes, an intern a llouver in t he end duct opens. These adjustm ents a re determinedby a controller which senses air tempera ture a t the fan . Once thesystem reaches equilibrium, it automatically controls processtemperature and prevents excessive cooling. During warmwea ther, the side ma nua l louvers a re opened, wh ile close contr olis ma intained by a djustment of the exhaust louvers.

The externa l recirculation design is preferred for critical controla nd prevention of freezing. Once operationa l, it requires little at -tention. Upon failure of power or air supply, the system closesautomatically to prevent freezing. It can be designed to automat-ically reduce m otor energy use w hen excess cooling is being pro-vided. The ma in dra wba ck for th is type of system is its high cost.Several actuators and control devices are required, along withmore steel and louvers. It is usua lly too large to be shop assembled,and requires more field assembly than an internal system. Be-cause of the need to restrict air intake, this design increases thestatic pressure, causing greater energy use, and 20-25%largermotors tha n a standa rd cooler.

When designing a n externa l recirculation unit , considertion must be given to the plenum depth a nd duct w ork to alloa ir mixing a nd prevent excessive sta tic pressure loss. The lover intake area should be large enough to keep the air flobelow 500 ft/min du ring m axim um design condit ions.

AIR EVAPORATIVE COOLERS

Wet/dry type (air evapora tive coolers) a ir coolers m a y begood economical choice when a close approa ch to the a mbietemperature is required. In these systems, the designer cta ke adva nta ge of the difference between the dry bulb and wbulb temperatur es. There ar e tw o genera l types of air evapra tive cooler combinations used a lthough other combinatioare possible:

Wet air type In this type, the air is humidified bspraying wa ter into the a ir s trea m on the inlet s ide of the acooler. The air st ream ma y then pa ss through a mist elimintor to remove the excess wa ter. The air then pa sses over tfinned tubes a t close to its wet -bulb temperatur e. If the mi

eliminator is not used, the spray should be clean, treatwa ter or the t ube/fin type a nd met allurgy should be compaible with t he wa ter.

Wet tube type An a ir eva porat ive cooler may be opeat ed in series with an air cooler if there is a la rge process flutemperat ure change wit h a close approach to the ambient. Tprocess fluid enters a dry finned t ube section and then pa ssinto a wet , plain t ube section (or a ppropriate finned tube setion). The a ir is pulled across the w et tube section and theafter dropping out the excess moisture, passes over the dtube section.

Access door for each bay

Automatic louvers

Handrail

Grating

walkway

Automatic louvers

Bug and lint screenwhen required

Manual louvers

Hingedaccessdoor

Fixed panel inrecirculation compartment

Coil

Coil guard

Manual louvers

Bug and lint screen when required

FIG. 10-7

External Recirculation Design

10-5


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SPECIAL PROBLEMS IN STEAMCONDENSERS

There a re often problems wit h stea m condensers which needspecia l a t tent ion a t the design s tage.

Imploding (collapsing bubbles) or knocking can create vio-lent fluid forces which may damage piping or equipment.These forces are created when a subcooled condensate isdumped into a two-phase condensate header, or when livesteam passes into subcooled condensate. This problem is

avoided by designing the steam system and controls so thatstea m a nd subcooled condensa te do not meet in the system.

Non-condensable ga s sta gna tion can be a problem in th e aircooled stea m condenser a ny tim e there is more tha n one tuberow per pass. The temperat ure of the air increases row by rowfrom bottom to top of the air cooled section. The condensingcapacity of each row w ill therefore vary w ith each tube row inproportion to the T driving force. Since the tubes are con-nected to common headers a nd a re subject to the same pres-sure drop, the vapor flows into the bottom rows from both ends.The non-condensables a re tra pped w ithin t he tube a t th e pointof low est pressur e. The non-condensa bles cont inue to a ccumu-late in all but the top rows until they reach the tube outlet.The system becomes stable w ith t he condensa te runn ing out

of these lower t ube rows by g ra vity. This problem can be elimi-nat ed in several w ays :

B y a ssigning only one tube row per pass. B y connecting the tube rows at the return end w ith 180

return bends a nd eliminat ing the common head er.

AIR COOLER LOCATION

Circulation of hot a ir to the fans of an a ir cooler can grea tlyreduce the cooling capacity of an air cooler. Cooler locationshould take t his into considerat ion.

Single Installations

Avoid loca tin g th e a ir-cooled excha nger too close to buildin gsor structures in the downw ind direction. Hot a ir venting fromthe air cooler is carried by the wind, and after striking theobstruction, some of the hot air recycles to the inlet. An in-duced draft fan with sufficient stack height alleviates thisproblem, but locat ing the a ir cooler a wa y from such obstruc-tions is the best solution.

An a ir cooler w ith forced dra ft fa ns is alw ay s susceptible toair recirculation. If the air cooler is located too close to theground, causing high inlet velocities relative to the exhaustair velocity leaving the cooler, the hot air recirculation canbecome very significant. Forced d ra ft coolers a re preferablylocat ed above pipe lanes rela tively high a bove the ground. In-duced draft coolers are less likely to experience recirculationbecause the exhaust velocities are normally considerablyhigher tha n the inlet velocities.

Banks of Coolers

Coolers arr an ged in a ba nk should be close together or haveair seals between them to prevent recirculation between theunits. Mixing of induced draft an d forced dra ft unit s in closeproximity to each other invites recirculation. Avoid placingcoolers at different elevat ions in the same ban k.

Avoid placing the ba nk of coolers down wind from other hea tgenerating equipment.

Since air can only enter on t he ends of coolers in a bank, t hebank sh ould be locat ed above ground high enough to assure areasonably low inlet velocity.

The prevailing summer wind direction can ha ve a profoundeffect on th e performa nce of the coolers. Normally the ba nkshould be oriented such th at the w ind flows para llel to the longaxis of the ba nk of coolers, and t he items w ith t he closest ap-proach to the ambient temperature should be located on theupwind end of the bank.

These generalizations are helpful in locating coolers. The

use of Computational Fluid Dynamics to study the effect ofwind direction, velocity, obstructions, a nd h eat generat ing ob-jects should be considered to a ssure the best locat ion and ori-entation of air cooled heat exchangers, especially for largeinsta lla t ions .

MULTIPLE SERVICE DISCUSSION

If different services can be placed in the sa me plot a rea w ith-out excessive piping runs, it is usually less expensive to com-bine them on one structure, with each service having asepara te sect ion, but sha ring the same fan and motors . Sepa-rate louvers may be placed on each service to allow inde-pendent control. The cost and space savings makes thismethod common practice in th e a ir cooler industr y.

In designing mult iple service coolers, the service with themost critical pressure drop should be calculated first. This isbecause t he pressure drop on t he critical item might restrictthe ma ximum tube length th at the other services could toler-at e. The burden of forcing more th an one service into a singletube length increases the possibility of design errors. Severa ltria l calculat ions may be needed to obta in an efficient design.

After all service plot areas have been estimated, combinethem int o a unit ha ving a ra tio of 2 or 3 to 1 in length to widt h(assuming a two fan cooler). After assuming a tube length,calculate t he most critical service for pressure drop using theassumed number and length of tubes and a s ingle pass . I f thedrop is acceptable or very close, calculate the critical servicecompletely. Once a design for th e most critical ser vice has been

completed, follow the sa me procedure wit h t he next m ost criti-cal service. After t he second or subseq uent s ervices have beenra ted, it is often necessary to lengthen or shorten t he tubes orchange the overall arr an gement. If tubes need to be ad ded forpressure drop reductions in a lready oversurfaced sections, itmight be more cost effective to add a row(s) rat her tha n widenthe entire unit . The fan a nd motor calculat ions ar e the sameas for a single service unit , except tha t th e qua ntity of air usedmust be the sum of a ir required by a ll services.

CONDENSING DISCUSSION

The exam ple given covers cooling problems a nd w ould workwit h stra ight line condensing problems that ha ve the approxi-ma te ra nge of dew point t o bubble point of the fluid. Where

de-superheating or subcooling or where disproportionateam ounts of condensing occur a t certain t emperatur es, as withsteam and non-condensables, calculations for air coolersshould be done by zones. A hea t relea se curve developedfrom entha lpy data will show the qua nt ity of heat t o be diss i-pat ed between va rious temperatur es. The zones to be calcu-lated should be straight line zones; that is, from the inlettemperature of a zone to its outlet, the heat load per degreetemperature is t he same.

After the zones ar e determined, an a pproximat e rate mustbe found for each zone. Do this by taking rates from vapor

10-6


7/16

cooling, condensing, and liquid cooling, then average thesebased on the percent of heat load for that phase within thezone. Next, calculate th e LMTD of each zone. Begin w ith t heoutlet zone using the final design outlet tempera ture a nd theinlet tempera ture of that zone. Continue to calculate the zoneas if it w ere a cooler, except tha t only one pass an d one or tw orows should be assumed, d epending on th e percenta ge of heatload in that zone. In calculating the pressure drop, averageconditions may be used for estima ting.

If t he calculations for zone one (or lat er a succeeding zone)show a la rge number of short t ubes with one pass, as is usuallythe case with steam and non-condensables, recalculate thezone with mu ltiple rows (usually four) an d short tubes ha vingone pass t ha t uses only a percentage of the total pressure dropallowed. The tota l cooler w ill be calculat ed as if each zone werea cooler connected in series to the next one, except that onlytube pressure drops should be calculated for t he middle zones.Thus, each zone must ha ve the sam e number of tubes and trueam bient must be used in calculating t he LMTD. Only the t ubelength ma y va ry, with odd length s for a zone acceptable a s longas overall length is rounded to a sta nda rd tube length.

If the calculations for zone one (and succeeding zones) fitwell into a longer tube length, the LMTD m ust be weighted.

After th e outlet zone has been calculated, calculate zone tw ousing the inlet temperature for i t and its out let t emperat ure,wh ich is t he inlet t emperatur e of zone one. The a mbient usedto find the zone tw o LMTD w ill be the design ambient plus theair r ise from zone one. Continue in this ma nner, alw ay s usingthe previous zones outlet air t emperatu re in calculating t hecurrent zones LMTD. After the cooler size and configurationha ve been determined, the fan an d motor calculat ions will bemade in t he normal manner.

The ultima te pressure drop is the sum of the drops for ea chzone or a pproximat ely the sum of the drop for each pha se usingthe tube length an d pass ar ran gement for each phase. An es-tima ted overall tube side coefficient ma y be calculated by es-tima ting t he coefficient for each phase. Then ta ke a w eightedavera ge based on the percenta ge of heat load for each phase.

The tota l LMTD m ust be th e weighted a verage of the calcu-lat ed zone LMTDs.

THERMAL DESIGN

The basic equation to be satisfied is the same as given inSection 9, Heat E xchangers:

Q = UA CMTD Eq 10-1Normally Q is known , U an d CMTD a re calculat ed, and the

equat ion is solved for A. The a mbient a ir tempera ture t o beused will either be known from ava ilable plan t da ta or can beselected from the summer dry bulb temperat ure dat a given inSection 11, Cooling Towers. The design a mbient a ir temper a-ture is usually considered to be the dry bulb temperat ure tha t

is exceeded less tha n 5 percent of the time in th e area wherethe installation is required.

A complication a rises in ca lculat ing t he LMTD because theair qua nt ity is a variable, and therefore the a ir out let tempera-ture is not known. The procedure given here start s wit h a st epfor approxima ting t he a ir-temperat ure rise. After th e air-out-let tempera ture ha s been determined, the corrected LMTD iscalculated in th e mann er described in the shell and tube sec-tion, except tha t MTD correction factors to be used are fromFigs.10-8 a nd 10-9 which have been developed for the cross-flow situa tion existing in air-cooled exchangers.

Fi g. 10-8 is for one tube pass. It is also used for multiple tupasses if passes are side by side. Fig. 10-9 is for two tupasses and is used if the tube passes are over and under eaother. A MTD correction fact or of 1.0 is used for four or mopasses, if passes are over and under each other. A correctifactor of 1.0 may be used a s an approximation for three passealthough the factor will be slightly lower than 1.0 in somcases.

The procedure for the thermal design of an air cooler cosists of assuming a selection an d th en proving it to be correThe typical overall heat transfer coefficients given in F i10-10are used to approximate t he heat t ransfer area requireThe heat tr an sfer area is converted to a bundle face area u siFig. 10-11 which lists the amount of extended surface avaable per squa re foot of bundle a rea for tw o specific fin t ubon two different tube pitches for 3, 4, 5, and 6 rows. After asuming a t ube length,Fig. 10-11is also used to ascertain thnumber of tubes. Both the t ube side and a ir side mass veloties are now determina ble.

The tu be-side film coefficient is ca lculat ed from F igs.10-a nd 10-13.Fig . 10-17gives the air-side film coefficient bason outside extended surface. Since all resistances must based on the sa me surfa ce, it is necessary to multiply the r

ciproca l of the t ube-side film coefficient a nd t ube-side foulifactor by t he ra tio of the outside surface to inside surface. Thresults in an overall tran sfer rate based on extended surfacdesignated a s U x. The equation for overall heat transfer rais :

1

U x=

1

h t

Ax

Ai

+ rdt

Ax

Ai

+ rmx +

1

h a Eq 10

The basic equation will then yield a heat transfer area extended surface, Ax, and becomes:

Q = (U x)(Ax)CMTDEit her method is valid an d each is used extensively by the

mal design engineers. Fig. 10-10 gives typical overall hetra nsfer coefficients based on both extended surfa ce and ouside bare sur face, so either met hod ma y be used. The extendsurface method ha s been selected for use in th e example wh ifollows . The a ir-film coefficient in Fig . 10-17and the a i r s t a tpressure drop inFi g. 10-18 are only for 1 in. OD tubes wi58 in. high fins, 10 fins per inch on 214 in . t r iangular pitcRefer to B ibliogra phy Nos. 2, 3, a nd 5 for informat ion on othfin configura tions and spa cings.

The minimum fan area is calculated in Step 16 using tbundle face ar ea, number of fans, a nd a minim um fan coveraof 0.40. The calculated a rea is then converted t o a d iametand rounded up to the next avai lable fan size. The air-sista tic pressure is calculat ed fromFig . 10-18 and the fan topressure is estimat ed using gross fan a rea in S tep 20. Fina lfan horsepower is calculated in Step 21 assuming a fan efciency of 70%, and d river horsepower is est ima ted by a ssum

ing a 92%-efficien t speed r educer.Example 10-1 P rocedure for estima ting tr an sfer surfac

plot a rea, a nd horsepower

Required data for hot fluid

Nam e an d pha se: 48 AP I hydrocarbon liquid

P hysical propert ies a t avg t emp = 200F

C p = 0.55B tu/(lb F )

= 0.51 cp

10-7


8/16

FIG. 10-8

MTD Correction Factors (1 Pass Cross Flow, Both Fluids Unmixed)

FIG. 10-9

MTD Correction Factors (2 Pass Cross Flow, Both Fluids Unmixed)

10-8


9/16

k = 0.0766 Bt u/[(hr sq ft F )/ft ]

(From this D at a B ook Section 23)

Hea t load: Q = 15,000,000 B tu/hr

Flow quant ity: Wt = 273,000 lb/hr

Temperature in: T1 = 250F

Tempera tur e out: T2 = 150F

Fouling factor rdt = 0.001 (hr sq ft F)/B tu

Allowable pressure drop: P t = 5 psi

Required data for air

Ambient temperat ure: t 1 = 100F

Eleva t ion : Sea level (seeFig . 10-16for altitudecorrection)

C Pa i r = 0.24 B tu /(lb F )

Basic assumptions

Type: Forced dra ft , 2 fans

Fintube: 1 in. OD with 58 in. high fins

Tube pitch: 2 12 in . t r iangular ()

B undle layout: 3 tube passes, 4 rows of tube30 ft long t ubes

First trial

1. Pick a pproximat e overall t ra nsfer coefficient from F10-10. U x = 4.2

2. Calculate approximate a ir temperature r ise

t a =

U x+110

T1+ T22

t 1

t a =

4.2 + 1.010

250+1502 100

= 52 F

3. Ca lcu la te CMTD

Hydocarbon

Air

250

152_____98

150100_____

50

LMTD = 71.3 F (see Fi g. 9-3)

CM TD = (71.3) (1.00) = 71.3F(3 tube passes a ssumed)

4. Ca lcu la te required sur face

Ax =Q

(U x)(CMTD)

Ax =15,000,000

(4.2)(71.3) = 50,090sq f t

5. Calculate face area using APS F factor fromFig. 10-11

F a =Ax

APSF

F a =50,090

107.2 = 467 sq ft (4 rows a ssumed)

6. Ca lcu la te unit w idth wi th assumed tube length

Service

1 in. Fintube

12 in. by 9 58 in. by 10

U b U x U b U x

1. Wa ter & wa ter solutions

(See note below)

Engine jacket water(rd = 0.001) 110 7.5 130 6.1

Process water(rd = 0.002) 95 6.5 110 5.2

50-50 ethylene glycol-water (rd = 0.001) 90 6.2 105 4.9

50-50 ethylene glycol-water (rd = 0.002) 80 5.5 95 4.4

2. Hy drocar bon liquid coolers

Viscosity, cp,a t a vg . temp.

U b U x U b U x

0.2 85 5.9 100 4.7

0.5 75 5.2 90 4.2

1.0 65 4.5 75 3.5

2.5 45 3.1 55 2.6

4.0 30 2.1 35 1.6

6.0 20 1.4 25 1.2

10.0 10 0.7 13 0.6

3. Hyd rocarbon ga s coolers

Pressure,psig

U b U x U b U x

50 30 2.1 35 1.6

100 35 2.4 40 1.9

300 45 3.1 55 2.6

500 55 3.8 65 3.0

750 65 4.5 75 3.5

1000 75 5.2 90 4.2

4. Air an d flue-ga s coolersUse one-ha lf of value given for h ydrocarbon ga s coolers.

5. St eam Condens ers (Atm ospheric pressur e & ab ove)

U b U x U b U x

Pure Steam(rd = 0.0005) 125 8.6 145 6.8

Steam withnon -conden sibles 60 4.1 70 3.3

6. HC condensers

Condensing*Range , F

U b U x U b U x

0 ran ge 85 5.9 100 4.7

10 ra nge 80 5.5 95 4.4

25 ra nge 75 5.2 90 4.2

60 ra nge 65 4.5 75 3.5

100 & over ran ge 60 4.1 70 3.3

7. Other condensers

U b U x U b U x

Ammonia 110 7.6 130 6.1

Fr eon 12 65 4.5 75 3.5

Notes: U b is overall rate based on bare tube area, and U x is overall ratebased on extended surfa ce.

Based on approximate air face mass velocit ies between 2600 and 2800lb /(hr sq ft of face area).

*Condensing range = hydrocarbon inlet temperature to condensing zoneminus hydrocarbon outlet temperature from condensing zone.

FIG. 10-10

Typical Overall Heat-Transfer Coefficients for Air Coolers

10-9
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10/16

Width =F a

L

Widt h =467

30 = 15.57 ft

For simplificat ion round this a nswer to 15.5, thus F a =465 (30-ft-long tu bes a ssumed )

7. Ca lculate number of tubes using APF factor fromFig. 10-11

N t =Ax

(APF )(L)

N t =50,090

(5.58)(30) = 299

8. Ca lculate t ube-side mass velocity from a ssumed numberof passes and reading At fromFig . 9-25 for a 1 in. OD x16 BWG tube

At = 0.5945 sq in.

G t =(144)(Wt)(Np)(3600)(N t)(At)

G t =(0.04)(273,000)(3)

(299)(0.5945) = 184 lb/(f t2 sec)

9. Ca lculate modified Reynolds number

NR =(D i)(G t)

=

(0.87)(184)0.51

= 314

10. Ca lculate tube-side pressure drop using equation fromFig . 10-14and fromFig . 10-15

P t =fYLNp

+ B Np

P t =(0.0024)(14.5)(30)(3)

0.96+(0.25)(3) = 4.0 psi

( is a difficult function to calculate rigorously, seeF ig .10-19)

11. C a lcu la t e tube-side film coefficient using equation fromFig . 10-15 a nd

k

C pk

13

fromFig . 10-13

h t =

J k

C pk

13

D i =

(1900)(0.12)(0.96)0.87

= 252

12. Ca lcu la te a i r quant i t y

Wa =Q

(0.24)(t a)

Wa =15,000,000

(0.24)(52) = 1,200,000 lb/hr

13. Ca lculate a ir face mass velocity

G a =WaF a

= lb /(hr sq ft of face area )

G a =1,200,000

465 = 2,581

14. Read a ir-side film coefficient fromFig . 10-17

h a = 8.515. Ca lculate overall t ra nsfer coefficient

Ax

Ai =

(AR)(D o)D i

Ax

Ai =

(21.4)(1.0)0.87

= 24.6

1

U x =

1

h t

Ax

Ai

+ rdt

Ax

Ai

+ rmx +

1

h a

1

U x =

1

252

(24.6)+(0.001)(24.6)+ 1

8.5

U x = 4.17(rmx is omitted from calculations, since meta l resistanceis small compared to other resista nces)

Second and subsequent trials. If U x calculated inSt ep 15 is equa l or slightly great er tha n U x assum ed in Step 1,and calculated pressure drop in Step 9 is within allowablepressure drop, the solution is acceptable. Proceed to Step 16.Otherw ise, repeat St eps 1-15 as follows:

1. As su me n ew U x between value originally assumed inSt ep 1 an d va lue calculat ed in Step 15.

Fin Height by Fins/inch 12 in. by 9 58 in. by 10APF , sq ft/ft 3.80 5.58

AR, sq ft /sq ft 14.5 21.4

Tube Pitch 2 in. 214 in . 214 in . 238 in . 212 in .

APS F (3 row s) 68.4 60.6 89.1 84.8 80.4

(4 row s) 91.2 80.8 118.8 113.0 107.2

(5 row s) 114.0 101.0 148.5 141.3 134.0

(6 row s) 136.8 121.2 178.2 169.6 160.8

Notes: AP F is t otal externa l ar ea/ft of fintube in sq ft /ft . AR is th e area ra tio of fintube compared to the exteriorarea of 1 in. OD bar e tube wh ich ha s 0.262 sq ft/ft . AP SF is the external a rea in sq ft/sq ft of bundle face area .

FIG. 10-11

Fintube Data for 1-in. OD Tubes

10-10
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11/16

FIG. 10-12

Friction Factor for Fluids Flowing Inside Tubes

10-11


12/16

FIG. 10-13

Physical Property Factor for Hydrocarbon Liquids

10-12


13/16

FIG. 10-14

Pressure Drop for Fluids Flowing Inside Tubes

10-13


14/16

FIG. 10-15

J Factor Correlation to Calculate Inside Film Coefficient, ht

10-14


15/16

2. Adjust t a by increasing t a if calculated U x is higherthan assumed U x, or decreasing t a if calculated U x islower tha n assumed U x.

3.-15. Recalculat e values in Steps 3-15 changing a ssumed

number of passes in Steps 3 and 8, and tube length inSt ep 6, if necessar y to obtain a pressure drop as calcu-lat ed in St ep 9 as high a s possible wit hout exceeding theallowable.

16. Ca lcu la te minimum fan a rea .

Fan a rea /f a n = F A P F =(0.40)(F a)(No. fan s)

FAPF =(0.40)(465)

2 = 93 ft 2(2 fa ns a ssumed)

17. Fan diameter = [4 (FAP F)/]0.5 = [4 (93)/3.1416]0.5

= 11 ft (round ed up)18. Calculate a ir s ta t ic pressure drop using Fp from F

10-18 an d D R at avg a ir temp from Fig . 10-16.

Ta , avg =100F + 152F

2 = 126 F

P a =(Fp)(N)(D R)

P a =(0.10)(4)

0.90 = 0.44 i n c h e s of wa ter

19. Calculate actual a ir volume using D R of a ir a t fa n inle

t 1 = 100F

250

200

150

100

50

0

-50

1000.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4

Density ratio, Dg, dimensionless

Temperature,

F

Reference state dry air at 70Fand sea level, 14.7 psia

ElevationFt.

8,0007,000

6,0005,000

4,0003,000

2,0001,000 0

FIG. 10-16

Air-Density Ratio Chart

FIG. 10-17Air Film Coefficient

FIG. 10-18

Air Static-Pressure Drop

Correction factor

when =

w

0.14

(SeeFig. 10-1

CorrectionFactor,

1. Hydrocarbon vapor ; s team; wa t er 1.0

2. Hydrocarbon liquids (18 to 48 API ), MEA/DE Asolutions

0.96

3. Water/glycol solutions; heat tran sfer fluids 0.92

4. Lube oils; heavy petroleum fractions (10 to 18 API) 0.85

When N r


16/16

ACFM =Wa

(DR)(60)(0.0749)

ACFM =1,200,000

(0.94)(60)(0.0749) = 284,000 Tota l

or 142,000 / Fa n20. Approximate fan tota l pressure using D R of a ir a t fa n and

fan a rea .

P F = P a+

ACFM

4005

D 24

2

(D R)

Where: 4005 =2 g w(3600)a 12 a t 70 F

P F = 0.44+

142,000

(4005)(0.785)(112)

2

(0.94)

= 0.57 inches of wa ter

21. Approximat e bra ke horsepower per fa n, using 70%fa nefficiency.

bh p =(ACF M/fa n)(P F )

(6356)(0.70)

Where th e conversion fa ctor

6356 =

33,000 ftlbmin hp

12 in.

ft

ft 3

62.3 lb

Note: 62.3 is the weight of one cubic foot of wa ter a t 60F.

bh p =(142,000)(0.57)(6356)(0.70)

= 18.2

Actual fan motor needed for 92%efficient speed reducer is18.2/0.92 = 19.8 hp. F or th is a pplication, 20 hp d rivers w ould

probably be selected.

Solution:

(15.5 ft ) (30 ft ) = 465 sq ft

(465 sq ft) (AP SF ) = extended surface ar ea

(465) (107.2) = 49,848 sq ft

Therefore, one unit having 49,848 sq ft of extended surface,tw o 11 ft diameter fa ns, and tw o 20 hp fan d rivers, is required.

MAINTENANCE AND INSPECTION

Attention to the design of the air cooler, and the choice ofma teria ls, is essential to provide low ma intena nce operation.Major factors to be considered are atmospheric corrosion, cli-matic conditions, and temperature cycling of fluid beingcooled.

Scheduled preventive maintenance and inspection is thekey t o trouble-free a ir cooler opera tion. A check of a ll fa ns forvibrat ion should be ma de regularly. At t he first sign of unduevibration on a unit , th e unit should be shut down a t th e earliestopportunit y for thorough examina tion of all moving part s. Asemi-a nnua l inspection an d ma intena nce program should:

Check and replace worn or cracked belts. Inspect fan blades for deflection and for cracks near

hubs.

Grease a ll bearings. Cha nge oil in gear drives. Check the inside of tube section for accumulation of

grease, dirt , bugs, leaves, etc., an d schedule cleaning be-fore tubes become packed with such debris.

BIBLIOGRAPHY

1. A.P.I . Sta ndard 661, Air Cooled Heat Exchangers for General

Refinery S ervices.

2. Briggs, D. E., Young, E. H., Convection Heat Transfer and P res-

sur e Drop of Air Flow ing Across Tria ngula r P itch of Tubes,

Chemical E ngineering P rogress Sym posium S eries, Volume 59,

No. 41, 1963.

3. Cook, E. M., Air Cooled Heat Exchangers, Chemical Engineer-

ing, Ma y 25, 1964, p. 137; J uly 6, 1964, p. 131; and August 3,

1964, p. 97.

4. Ga rdner, K. A., Efficiency of Extended Surfaces, Trans ASME,Volume 67, 1945, pp. 621-631.

5. Robinson, K. K., Briggs, D. E., P ressure Drop of Air Flowing

Across Triangu lar P itch B anks of Finned Tubes, Chemical E n-

gineering P rogress S ymposium Series, Volume 62, No. 64, 1966.

6. Rubin, Frank L., Winterizing Air Cooled Heat Exchangers, Hy-

drocar bon P rocessing, October 1980, pp. 147-149.

10-16

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