process heat transfer hof master

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A Workbook Collection of Dale Gulley's Heat Exchanger Tips

This workbook contains a compilation of edited, and formatted valuable and practical "Tips" on Heat

Exchangers that have been published and offered to the engineering public by Mr. ale !ulley, an experienc

and recognied authority on Heat Exchanger esign, #abrication, and $onsulting.

#or some years , % have endeavored to collect as many of ale&s valued advice and "tips" as % possibly could

'y doing so, % have gained further insight and knowledge by reading and applying his tips and know(how.

ale is not only an outstanding and recognied heat transfer expert, but he has been a contributing and posit

member of The Tubular Exchanger Manufacturers& )ssociation for many years, advocating the useful and

positive efforts this organatn has done for the engineerng profession world(wide.

%n the past *+ years % have arrived at many conclusions and results in dealing with the design, specification,

fabrication, and operation of heat exchangers that are identical with ale&s Tips. My experience also coincid

with that of a lot of my past and present engineering colleagues. My field experience has proven ale&s

advice and Tips to be not only credible ( but also valuable in applying heat transfer to process operations.

% have put my effort into this compilation in order to make use of this valuable engineering know(how as a

Through out this compilation, my personal notes on some of the Tips can be seen off the printed area of the

worksheet and to the right(hand side. % have used this method to record my own experience related to the to

and to add empirical support and reinforcement to what ale describes.

lease note that % have used the following spreadsheet and workbook techni-ues to assist in employing the

ideas and recommendations expounded by ale

/ The bulk of the Tips are organied in the same manner as they are found in ale&s 0ebsite. %

have made use of Exel&s Hyperlink feature to facilitate the -uick and accurate access to any of

the topics that are listed and grouped in the Table of $ontents. 1nce you locate a sub2ect or

topic that you want to read or persue in the Table of $ontents, all you have to do is click on the

sub2ect and the hyperlink will take you directly to the selected Tip.

/ % have made every effort to convert ale&s original presentation of recommended calculations and

e-uations to a format that allows the reader to immediately employ his3her basic data to make the

indicated calculation using Excel&s basic spreadsheet feature. The reader can type in the basic data

numbers. This allows the reader to do several "what(if" calculations -uickly to get an idea of the

perceived effect on the heat exchanger.

/ The various groups of the Tips sub2ect matter are also hyper(linked and a reader can go directly to

one of the groups of Tips directly from the Table of $ontents.

These Tips are compiled and freely distributed with ale !ulley&s permission and approval. % would ask all

engineers who are helped and assisted by this contribution to call or email ale with thanks and gratitude fo

contribution to heat exchange. ale is active in heat exchange design, software, and process engineering

basis for Experienced Based earning when dealing with heat exchangers.

in the !E"W(filled cells and the resultant calculated answer will be generated in B"D #ED



out of Tulsa, 1klahoma. 4eedless to say, his organiation can be of great help in a heat exchanger applicati

)rt Montemayor ( +* )pril 5+66



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735835+66 ( )dded 's factor the second term in the denominator for the e-uation $hris Hasl

for the seal bar calculation on sheet "$alculations". $heresour

835935+66 ( )dded three new tips from !ulleyassociates.com with permission. $hris Hasl

Boiling $heresour

Calculations

Construction

6636+35+66 ( )dded five new tips from !ulleyassociates.com with permission $hris Hasl

Boiling $heresour

Construction

Misc.

73*35+65 ( )dded five new tips from !ulleyassociates.com with permission $hris Hasl

Boiling $heresour

Calculations

Tube Bundle Vibration

Misc.

Estimate ( critical heat flux for propane chillers.

Estimate ( optimum flow velocity for gas inside tubes.

:ongitudinal baffle heat conduction cures.

;ettle <eboilers ( =upports or 'affles

esign Temperatures of $arbon =teel and :ow )lloy Tub

esign Temperatures of 4onferrous Tubes and Tubesheet

#ouling factors for water>hr(ft5(#3'tu?

#ouling #actors for :i-uid Hydrocarbons>hr(ft5(#3'tu?

@ertical Thermosyphon($alculate ressure rop at The

@ertical Thermosyphon(esign for a =maller :i-uid reh

Estimate ( Hydrocarbon !as Heat Transfer $oefficient in

'est esign #eature to revent 'undle @ibration

@iscous #low ( Bse More ressure rop Than Bsual



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es and Tubesheets

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utlet 4ole

eat Aone

=hell =ide



5C66 E. DD l., Tulsa, 1; D768 F .1. 'ox D++5C*, Tulsa, 1;. D76D+(+5C*

hone >C69? D77(+6++

D.

67.

)ir $oolers

6. )ir flow accessories ( don&t overlook these when calculating fan H

5. 'ox header design ( limit of process temperature change

. $onnecting bundles of existing coolers for a new service

7. #an drive changes that increase capacity of existing cooler

*. #an drive noise ( suggestions on how to reduce

8. Maximum motor H for a fan

D. Maximum tube wall temperature for wrap(on fins

9. 1ptimum number of tube rows

C. 1verall heat transfer rate estimate for hydrocarbons

6+. 0hen do bare tubes become more efficient than fin tubesG

66. 0hen To limit number of tube passes in air coolers

65. 0hen to use wind coolers

'oiling

6. )void mist flow boiling inside tubes

5. ;ettle reboiler ( li-uid carryover problem solutions

. ;ettle reboiler ( shell nole arrangement problem

7. ;ettle reboiler ( shell vapor outlet nole location

*. ;ettle reboiler ( siing shell vapor space

8. ;ettle reboiler ( undersied shell effects

Estimate ( pool boiling heat transfer coefficient for hydrocarbons

9. :arge boiling temperature difference problems

C. :owest limit of boiling temperature difference

6+. @ertical thermosyphon ( choking two phase flow with small outlet no

66. @ertical thermosyphon ( minimum recirculation rate

65. @ertical thermosyphon ( check for li-uid preheat one

6. @ertical thermosyphon ( who sets recirculation rate

@ertical Thermosyphon($alculate ressure rop at The 1utlet 4ole



6*.

5.

57.

@ertical Thermosyphon(esign for a =maller :i-uid reheat Aone

$alculations

6. 0hat diameter to use to start design of a coil

5. Estimate ( gas heat transfer coefficient inside tubes

. Estimate ( hydrocarbon heat transfer coefficient in tubes

7. Estimate ( latent heat of hydrocarbons

*. Estimate ( li-uid thermal conductivity of light hydrocarbons

8. Estimate ( overall heat transfer coefficient in shell tube

D. Estimate ( tube length that lowers tube pressure drop

9. How to calculate excess surface and overdesign surface

C. Bse superficial velocities to calculate best heat transfer flow pattern

6+. :3 e-uation for heat Transfer coefficient inside tubing

66. :MT correction factor charts for TEM) ! and I type shells65. :ow :MT correction factor for divided flow

6. 0hat is the lowest :MT correction to use in shell tube

67. Minimum flow area for shell side inlet nole

6*. How to calculate performance of heat exchangers with plugged tubes

68. How to increase heat transfer for low <eynolds numbers

6D. $alculate when to use seal bars on a bundle to increase heat transfer

69. $alculate = T bundle diameter from number of tubes6C. E-uation for calculating tube count in shell tube

5+. $heck for hot tube wall temperature of cooling water

56. =ometimes larger tubes are better than small ones

55. 0eighted MT

Estimate ( optimum flow velocity for gas inside tubes.

Estimate ( Hydrocarbon !as Heat Transfer $oefficient in =hell =ide

$ondensing

6. )void small baffle cuts in = T condensers

5. Estimate ( $ondensing heat transfer coefficient for hydrocarbons insid

. Maximum heat transfer rate inside tubes for total condensation

7. Juick estimate for reflux condenser :MT in air cooler

*. <eflux >;nockback? condenser comments

8. =team condenser types



66.

6D.

D. =ulfur condenser ( design within tube velocity limits

9. 0arning about small temperature pinch points in condensers

C. 0hen to slope single tube pass tubes in condensing service

6+. Aone those condensersK

Estimate ( critical heat flux for propane chillers.

$onstruction

6. 'enefits of using rotated s-uare pitch in shell tube

5. $aution when using a longitudinal baffle in the shell side

. Bsing turbulators for tube side laminar flow

7. iscussion of types of triple segmental baffles in shell tube

*. $heck entrance and exit space for shell noles

8. Horiontal vs vertical baffle cut in shell tube

D. %s expansion 2oint re-uired in the shell of a fixed tube sheetG9. %ncreasing capacity of existing shell tube exchangers

C. :ocating vents on the shell side of vertical exchangers

6+. 1ptimun gasket location for flanges

66. <einforcing rods as tube inserts to increase heat transfer

65. =hell side impingement protection

6. =pecial shell tube heat exchanger type >4T%0?

67. 0hen to consider by(pass strips in shell tube bundle6*. 0hat is too large of temperature change in 5 tube passes G

68. 0hen to rotate s-uare tube pitch in shell tube exchanger

:ongitudinal baffle heat conduction cures.

Heat <ecovery

6. eciding on what fin spacing to use

5. Estimate of nole sie for H<=!. #ace area estimate for H<=! units

7. Maximum exhaust gas temperaure for steel fin tubes

*. 0hen to use bare tubes in waste heat boilers

Materials

6. $ooling water flowing inside +7== B(tubes



*.

ressure rop

6. )llowing for fouling in pressure drop calculations

5. )llowable pressure drop suggestions

. )llowable shell side pressure drop if a multi(leaf>a.k.a. lamaflex? long

7. 'etter baffle window pressure drop e-uation

*. esigning for better use of tube pressure drop

8. Effect of 6st tube rows on shell nole pressure drop

D. ressure drop on kettle side

9. <educing high shell side pressure drop in fixed tube sheet exchangers

C. Bse impingement rods instead of plate to lower shell press. drop

6+. 0hat design pressure drop to use for heavy li-uids inside tubes

66. Maximum velocity inside tubes

65. $alculate shell nole pressure drop

6. %mprove shell side pressure drop calculations

Tube 'undle @ibration

6. #eatures of a new = T bundle that replaces bundle that vibrated

5. @ibration cure when designing shell tube bundles

. $onditions likely to cause shell tube bundle vibration

7. $ures for vibration in existing bundle

'est esign #eature to revent 'undle @ibration

Miscellaneous

6. )llocation of streams in shell tube

5. )rticles published by ale !ulley

. )void these fluids when using lowfin tubing

7. 'est heat transfer flow pattern*. $heck li-uid thermal conductivity at high reduced temperatures

8. $heck piping connections when there is under(performance

D. Evaluating an exchanger for a new service

9. $heck heat release curve data for skipping over dewpoints and bubble

C. 0hen will exchangers with low(fins be more economical than exchan

6+. roblems with excess heat exchanger surface

66. urchasing warning for shell tube exchangers



57.

65. 0hat is the minimum velocity inside tubing for slurriesG

6. =uggestions for low(fins and potential = T bundle vibration

67. $hoose shell tube or multi(tube heat exchangers

6*. Thermal design problem with shell side long baffle

68. Trouble shooting article in Hydrocarbon rocessing

6D. Bnder(surfaced =T -uote

69. 0hen to add shell in =eries6C. 0hen to consider a long baffle in the shell

5+. 0hich stream goes inside the tubes of gas3gas exchangersG

56. 0eighted MT

55. 0hy did performance decline in a TEM) type #,! or H type shellG

5. Aone those condensers

@iscous #low ( Bse More ressure rop Than Bsual



$ote%

Air flo& accessories don't o(erlook lou(ers and screens &hen calculating fan H) March

)ir static pressure loss is used to calculate the horsepower re-uired for fans used in process air coolers. $harts and

e-uations in the literature are usually for the tube bundle only. #re-uently, air coolers have accessories like louvers

and fan guards. They may also have hail, bug, or lint screens. on&t overlook the accessory pressure drop because

they can increase the static pressure as much as 5*L.

Box header design li*it of process te*perature change March, 6CC9

%n the design of an )ir cooled heat exchanger, avoid imposing too large a temperature change in the box headers.

Too much temperature drop between the inlet and outlet tube passes can cause leakage where the tubes meet the

design. This allows a hot top section to slide past a cooler bottom section.

Connecting Bundles of Existing Coolers for a ne& +er(ice )pril, 6CC9

0hen re(using air cooled exchangers in a new service, don&t overlook connecting the bundles in a series(parallel

arrangement. 4ew air coolers nearly always have the bundle connected in parallel. )rrange the bundles for more

series type flow to increase the tube side velocity and get higher heat transfer rates. #or example, an air cooler with

six bundles could be arranged with four bundles in parallel, connected to two bundles in series. The two series

bundles would handle the coldest part of the heat load where higher velocity is needed the most.

,ncrease Capacity of Existing Air Cooler &ith -an Dri(e Changes 1ctober, 6CCD

%f you need to increase the capacity of an air cooler, don&t 2unk it for a new one until you have exhausted the

possibilities on changing the fan and the fan motor. The least expensive change is to increase the fan blade angle if it

will not overload the motor. 'ut check to make sure the blade angle is not already at the maximum. The next best

change in terms of cost is to increase the fan speed by changing the drive ratio between the fan and the motor. %f

these changes are not enough you could increase the motor sie or change the fan for one with more blades.

+uggestions to #educe -an Dri(e $oise

The most effective solution is to reduce the fan speed by changing the drive ratio between the fan and the motor.

1ther suggestions are to reduce the fan blade angle or change to a fan with more blades.

.axi*u* .otor H) for a -an

)dding more H to a fan will only work up to a point. The fan efficiency reaches a peak. Then increasing the H

will produce no more air. )n estimate for this H is

.ax H) / 01 2 345 6-an Dia* 7489 / 5:45 H

#an iam 14;; feet

This is for fan diameters greater than .* ft.

%nput data into !E"W cells and receive output in B"D #ED

tubesheet. %f the temperature change of the tube side stream is over approximately 7++ o#, then use a split header



Te*perature i*it of Wrap"n -ins for Aircoolers Iune, 5+++

)bove a certain temperature, it will be too hot for wrap(on fins. ue to thermal expansion, the aluminum fins will

lose good contact with the tubing. %n this case an integral type fin tube should be used. The summer time air outlet

temperature is a very rough approximation. To be more exact, the tube wall temperature needs to be calculated for

the hottest tube row. Then

58<

0here

temperature of tube wall

Ta air outlet temperature =;;

temperature inside tube 533

thermal resistance of air ;40=

clean overall heat transfer coefficient 148

5++ N >799 ( 5++? x +.65 x D.*

)s you can see, the problem is more severe at high heat transfer rates. 4ot even the aircooled manufacturers agree

exactly what this maximum tube wall temperature should be. The )=ME code for allowable stress of aluminum has

wrap(on fins.

"pti*un $u*ber of Tube #o&s

The optimum number of tube rows is a function of the maximum acceptable temperature rise of the air side. There

are three limitations and the smallest air rise of the three should be used. The limitations are

6 :imit the :MT correction factor to a minimum of +.C for one tube pass ( maximum air

outlet temperature to be the same as the process side outlet temperature.

5

T&all

/ Ta 2 6Th0 Ta9 x #o x >c / o#

Twall

o#

Th6 o#

< o hr(ft5(o#3'tu

B$ 'tu3hr(ft5(o#

Example =team is condensing at 799 o#. )ssume that the B$ is D.* and <

o is +.65.

%f the air outlet temperature is 5++ o#, then

Twall

Twall 7*Co

#

a maximum temperature of 7++ o#. % believe this is the upper limit. Then the above example is operating too hot for

Minimum temperature difference at the hot end to be 9 to 6+ o#.

Maximum air outlet temperature to be ++ o# if tension wound fins are used.



Hydrocarbon > Esti*ate 6AirCoolers9 #ebruary, 5++5

%n the preliminary design or checking of process air(coolers you need an estimate of the overall heat transfer

coefficient >B?. )n estimate that is based on fin surface can be made from the following

-luid in Tube side "(erall Heat Transfer Coefficient

:i-uids <t +.68* x =-rt >avg. tube viscosity? N +.67*

B 63<t

0here viscosity is less than c

!ases <t +.5C x =-rt >6++31? N +.67*

B 63<t

0here 1 is the operating pressure in =%)

When do Bare Tubes beco*e .ore Efficient Than -in Tubes?

%f the inside heat transfer coefficient beomes too low, fin tubes can become inefficient. This can be the case in

investigate both bare and fin tubes.

When To i*it Tube )asses in an Aircooler 4ovember, 6CCC

#or tube side streams that have a high heat transfer coefficient, it is probably not advantageous to use more thantwo tube passes. This would be for condensing streams like ammonia and steam. This could also be true for high

thermal conductivity li-uid streams if the :MT is high. The velocity on these type of streams will have a minor

effect on the overall heat transfer coefficient in the typical aircooler. The ma2or thermal resistance is the air side

heat transfer coefficient.

Air Cooler >sing Wind ecember, 5+++

0here cooling water is not available and the outlet temperature is not critical, an air cooler can be built that

depends only on the wind for cooling. %t will have the best performance when the tubes have high fins and the

tubes are perpendicular to the wind direction. %n areas where the wind does not have a prevailing direction,

arrange the tubes in a bird cage type pattern. Then there is cooling no matter which way the wind blows. %f there

is a prevailing wind direction, use an air cooler bundle that sets on a stand that faces the wind.

heavy oil coolers. %f it is expected that the heat transfer coefficient is below approximately 5+ 'tu3hr(ft5(o#,



+



.ist -lo& Boiling ,nside Tubes 4ovember, 5++6

This is a flow pattern to avoid in heat transfer. The mist flow region is dependent upon velocity, L vapor an

stratification effects. %n this type of flow the tube wall is mostly dry and the li-uid droplets are carried along

core. Therefore the heat transfer is much lower because the much higher thermal conductivity of the li-uid is

little contact with the tube wall. The higher the L vaporiation, the lower the velocity needs to be to avoid m

#or example in a vertical tube where the vaporiation is *+ L and the vapor density is 6.+ lb3cu ft, the velocit

to be below approximately 9+ ft3sec. %f the vaporiation is D* L, the maximum velocity is approximately +

This comes from the #air e-uation. %n a horiontal tube where there can be stratification, these maximum ve

much lower. %f the mist flow region cannot be avoided, then twisted tape turbulators can be used to increase

transfer. They will throw the li-uid in the vapor core toward the tube wall.

@ettle #eboiler ocation of apor "utlet $oles

0hen it is necessary to have dry vapor leaving the kettle side, the location of the noles is important. The in

should not be located directly under the vapor outlet. This probably results in some li-uid carryover. 0hen t

a single vapor outlet, it is usually centered over the bundle with the inlet nole located some distance away.

have been cases where someone other than the thermal designer changed the location of this vapor nole wi

the thermal designers 1;. %n one case the vapor outlet was moved to the back of the kettle resulting in appre

li-uid carryover

@ettle #eboiler )roble* +hell $ole Arrange*ent

=ometimes you see kettle reboilers where the inlet nole is directly under the outlet vapor nole. This arra

creates extra turbulence under the vapor nole which affects the amount of li-uid entrainment in the outlet v

is safer to use the conventional nole arrangement where the inlet is some lateral distance away unless a de

pad is used.

)nother problem with the vertical nole arrangement is when the kettle bundle is relatively long and there is

single pair of noles. Then there is no good flow distribution. The boiling ones near the ends of the bundl

have lower fluid circulation rates and lower heat transfer.

@ettle #eboiler ocation of apor "utlet $oles 1ctober, 5+++

0hen it is necessary to have dry vapor leaving the kettle side, the location of the noles is important. The i

nole should not be located directly under the vapor outlet. This probably results in some li-uid carryover.

there is a single vapor outlet, it is usually centered over the bundle with the inlet nole located some distanc

There have been cases where someone other than the thermal designer changed the location of this vapor no

without the thermal designers 1;. %n one case the vapor outlet was moved to the back of the kettle resulting

appreciable li-uid carryover

+iing the apor +pace in @ettle #eboilers Iune, 6CC9

The sie of the kettle is determined by several factors. 1ne factor is to provide enough space to slow the vap

velocity down enough for nearly all the li-uid droplets to fall back down by gravity to the boiling surface. T

amount of entrainment separation to design for depends on the nature of the vapor destination. ) distillation

with a large disengaging space, low tower efficiency and high reflux rate does not re-uire as much kettle vap

space as normal. 4ormally, the vapor outlet is centered over the bundle. Then the vapor comes from two dif

directions as it approaches the outlet nole. 1nly in rare cases are these two vapor streams e-ual in -uantity

simplification that has been extensively used is to assume the highest vapor flow is 8+L of the total. 1ne cas



where this would cause an undersied vapor space is when there is a much larger temperature difference at o

of the kettle then the other. The minimum height of the vapor space is typically 9 inches. %t is higher for hig

flux kettles.

@ettle #eboiler Effect of >ndersied @ettle Dia*eter Iuly, 6CCD

0hat effect will an undersied kettle diameter haveG The effect will be a decrease in the boiling coefficient.

coefficient depends on a nucleate boiling component and a two(phase component that depends on the recircu

rate. )n undersied kettle will not have enough space at the sides of the bundle for good recirculation. )not

effect is high entrainment or even a two(phase mixture going back to the tower.

Esti*ate )ool Boiling Heat Transfer Coefficient for Hydrocarbons

'oil h =<=8

0here

Ot >tube wall temperature ( li-uid temperatur 8;

t

arge Boiling Te*perature Differences March, 6CCC

:arge temperature differences in heat exchangers where li-uid is vaporied are a warning flag. 0hen the te

differences reach a certain value, the cooler li-uid can no longer reach the heating surface because of a vapor

This is called film boiling. %n this condition, the heat transfer deteriorates because of the lower thermal cond

of the vapor. %f a design analysis shows that the temperature difference is close to causing film boiling, the v

should be started with the boiling side full of relatively cooler li-uid. This way, you don&t start flashing the li

The li-uid is slowly heated up to a more stable condition. %f the vaporier is steam heated, the steam pressure

should be reduced which will reduce the temperature difference. 0ith steam heating, take a close look at the

o&er i*it of Boiling -il* Te*perature Difference #ebruary, 6CCD

) reboiler or chiller is best designed so that it doesn&t have the lower heat transfer mode of natural convectio

dividing line between natural convection and boiling depends on the type of tubing used. %f steel bare tubes a

Choking a ertical Ther*osyphon ecember, 6CCC

$hoking down on the channel outlet nole and piping reduces the circulation rate through a heat exchanger.

the tubeside heat transfer rate depends on velocity, the heat transfer is lower at reduced recirculation rates. )

of thumb says that the inside flow area of the channel outlet nole and piping should be the same as the flow

inside the tubing. The =hell 1il $ompany, in an experimental study, showed that a ratio of +.D in nole flow

area3tube flow area reduced the heat flux by 6+L. ) ratio of +.7 cut the heat flux almost in half.

)n approximate e-uation for the amount of heat flux reduction is

8747=

55 >Ot?6.5* 'tu3>hr?>ft5?>o#?

o#

temperature, o#

if the :MT is over C+ o#. This is close to the critical temperature difference where film boiling will start.

the lower limit of temperature difference between the tube wall and the boiling fluid is approximately * o#.

designed hydrocarbon chillers down to the temperature difference of 5 o# using low(finned tubes. =pecial en

tube surfaces can be used for even lower temperature differences than 5 o#.

<eduction .+8P (6.8P5 ( +.7



0here P area ratio ;45;

.ini*u* #ecirculation #ate in Ther*osyphon #eboilers

0hen does a recirculation rate become too low >high L vaporiation?G 0hen this happens, the tube wall is n

longer wet and the heat transfer diminishes. The guidelines in the literature show the lowest permissible reci

rates give from 5* to 7+L vaporiation for hydrocarbons. %t has been observed that this threshold is when th

outlet two(phase density >volume basis? is below 6.+ lb3cu(ft. 4early all thermosyphons have outlet densities

this value.

ertical Ther*osyphon Check for iuid )reheat Fone #ebruary, 5++6

0hen designing vertical thermosyphon reboilers with boiling at low operating process fluid pressures, check

presence of a li-uid preheat one. 'ack pressure raises the boiling point at the interface of li-uid preheat on

subcooled boiling. This boiling point rise creates a li-uid one with relatively low heat transfer and it reduce

temperature driving force >MT?. %f the operating pressure is below approximately 5* =%), there should be

li-uid preheat one. The lower the operating pressure, the more likely there is li-uid preheat. %f there is no li

preheat, there may be an input error.

ertical Ther*osyphon #ecirculation #ate ecember, 6CCD

%n the design of vertical thermosyphons, the recirculation rate should be set by the process engineer if thereanything unusual about the connecting piping. The recirculation rate is especially sensitive to the sie and co

of the outlet piping. %f the recirculation rate is left for the thermal designer to set, they will have to make pipi

assumptions that may be violated later in the actual installation.

Esti*ate Critical Heat -lux -or )ropane Chillers

) simple e-uation is presented for a kettle reboiler. %t is conservative for very small bundles.

The crital heat flux depends on the geometry of the bundle. The following estimate is based on 37 inch tubes

%t is actually good for any tube diameter with a tube pitch3tube diameter ratio of 6.5* and triangular tube pitc

) boiling temperature of (+ #. is assumed for the propane.

$H# 5*++ 0=355 'tu3h ft5

$H# crital heat flux in 'tu3>hr?>ft?5

s shell bundle diameter in inches 76

Example

0hat is the critical heat flux for a 76 inch diameter bundleG

$H# 5*++

$H# 65,9*+

@ettle #eboilers +upport or Baffles?

s>+.5*?

>76? +.5*

#or kettle reboilers use segmental baffles instead of full supports if shell fouling factor is greater Than +.++5>



ertical Ther*osyphonCalculate )ressure Drop at The "utlet $ole

) rule of thumb is that the pressure drop at the outlet nole should not be greater than +L of the total static

There is another tip in this boiling section about choking the flow with a small outlet nole. The inside flow

outlet nole should be the same or greater than the total flow area inside the tubing. #or a channel with a sid

the pressure drop is composed of a turning loss and a contraction loss The following e-uations calculate the p

drop at the outlet. The pressure drop for expansion into the channel is not included here but is with the tube p

+.77**D8 >used for pressure dro

>%f ;tr less than +.7+, use +.7

;c +.* >6 ( >4o3s?5? +.59+DD

;T ;tr N ;c +.D5878

+.68C77

0here

s Top channel % >inches? 67.9

;tr pressure loss coefficient for turning loss +.77**D8 >calculated?

;c pressure loss coefficient for contraction into nole +.59+DD

;T total pressure loss coefficient +.D5878

4o 1utlet nole % >inches? C.9 @n velocity thru nole >ft3sec? 65+

Qtp two(phase density >lb3ft? +.6*

On pressure drop thru channel and outlet nole >si? ;401

ertical Ther*osyphonDesign for a +*aller iuid )reheat Fone

)t low operating pressures there will be a sensible heat li-uid one with relatively low heat transfer. This is c

fact that a small pressure change will cause a large increase in the boiling point. There has been a case where

tube length was in the sub(cooled phase. 0hat can you change that will decrease the sie of the li-uid prehea

increase the overall heat transferG

1ne answer is to evaluate the piping system above the top tubesheet. %n order to make an evaluation check th

at the outlet. There is on this section of the website e-uations to calculate the pressure drop of a nole that is

the top channel. Most vertical thermosyphons have the outlet nole at right angles to the top channel. There

change of enlarging the outlet nole that would be the cure. 'ut there needs to be a check to make sure the n

connecting piping are not so large that there is li-uid slip. %f enlarging the right angle nole and piping is no

there are other configerations that will use less outlet pressure drop. 4ext the pressure drop of using a ' type

a long radius ell could be tried. %f this doesn&t do it, try a mitered channel design.

)nother solution to the problem is to investigate inserts such as swisted tape, wire matrix , or helically coiled

;tr RRR6RRRRRR

s+.

On ;T +.+++6+9 x @n5 x Qtp



n a vapor

in very

ist flow.

needs

t3sec.

ocities are

he heat

)rt&s 4ote

et nole % agree. % have also found that locating the inlet li-uid nole directly under the vapor outlet is no

ere is %n )mine ';B reboilers % found that locating the inlet rich amine li-uid as close to the B(tube bun

There gave the best, consistant results in obtaining good solution stripping. This gives the heating medi

hout

ciable

gement

apor. %t

ister

a

will

let

hen

away.

le

in

r

e

ower

r

erent

)

e



e end

heat

) boiling

ation

er

perature

film.

ctivity

porier

uid.

design

. The

re used,

=ince

rule

area

e have

anced



o

culation

above

for the

e and

the

a

-uid

ill befiguration

g

on 6*368 inch pitch.

.

r(ft2-F/Btu)



head.

area of the

outlet

ressure

essure drop.

p calc?

+?

used by the

C+L of the

one and

pressure drop

at right angle to

ay be a simple

ole and

the answer then

channel with



ood.

e tubesheet



$ote%

What Coil Dia*eter to >se to +tart Design 1ctober, 5++5

0hen starting to design a coil or other single continuous tube heat exchanger, the diameter is unknown. )

of this is an economier in a heat recovery system. %n this case it is desirable to have a single flow path rat

using parallel paths where headers are re-uired. The following gives guidelines for li-uids on a diameter s

+ie >nit Capacity flo& rate

6S tube ,+++(*,+++ 3 tube 3 hr

6 US pipe *,+++(6+,+++ 3 tube 3 hr

6 VS pipe 6+,+++(6D,+++ 3 pipe 3 hr

5S pipe 6D,+++(*,+++ 3 pipe 3 hr

S pipe *,+++(D+,+++ 3 pipe 3 hr

7S pipe D+,+++(6+,+++ 3 pipe 3 hr

Esti*ate Gas Heat Transfer #ate for Hydrocarbons #ebruary, 6CC9

%f you need to estimate a gas heat transfer rate or see if a program is getting a reasonable gas rate, use the f

h / 18 !enerally more accur

1r,

h :: !enerally understated

1perating pressure 0;;4; sia.

0 0=74;; lb3tube3hr

Esti*ate Hydrocarbon Heat Transfer Coefficient ,n Tubes

Bse the following e-uation to estimate the heat transfer coefficient when li-uid is flowing inside 37 inch t

Hio / 08; srt6a(g4 (iscosity9 / 31

0here

@iscosity 74; c.

Esti*ate atent Heat of Hydrocarbons

)n e-uation from the 'ureau of =tandards Miscellaneous ublication 4o. CD can be used when the =pecifi

is greater than +.8D and less than +.C7. %t is

%nput data into !E"W cells and receive output in B"D #ED

18 x 6"p4 pressure0;;90= / 'tu3hr(ft5(o#

6.70+.9 'tu3hr(ft5(o#

This is for inside the tubes. The rate will be lower for the shell side or if there is more than one exchanger.

'tu3#t5(hr(o#

This is limited to a maximum viscosity of 7 c



at heat / 6000 ;4;<T9+G:; / 007 'tu3lb

0here

:at heat The fluid&s :atent Heat in 'tu3lb

0;;

;4<;;; >+.8DW=!W+.C7?

#or hydrocarbons below a =pecific !ravity of +.8D and pressures below *+ psia, use

:at heat 6D5 ( +.6C* T

iuid Ther*al Conducti(ity for ight Hydrocarbons Iuly, 6CCC

Xou can make an estimate for the li-uid thermal conductivity of light hydrocarbons if you know their speci

%t is good for propane and heavier.

@ / ;4;5=1

=pecific Heat ;41;;; 'ut3lb Y#

Esti*ate "(erall Heat Transfer #ate 6>9 in + T ecember, 5++6

%n the preliminary design of shell and tube heat exchangers, you need an estimate of the overall heat transfe

>B?. rocess simulator programs give you a B) from which you can estimate the surface if you have a B v

)n estimate for a hydrocarbon B value can be made from the following

#t /

0here,

)vg. tube viscosity =4; c #ouling ;4;;;8 > / 0#t /

)vg. shell viscosity 74; c

limit on the shell viscosity. This is also limited to bare tube surface with no internal turbulation devices.

Esti*ated Tube ength That o&ers Tube )ressure Drop =eptember, 5++6

0hen the calculated tube side pressure drop exceeds the allowable, there are several design options. 1ne o

to design with shorter tubing when the number of tube passes is one. To estimate the new tube length, use

following e-uation

$e& g / 084< feet

0here

:g Existing tube length =; feet

)llowable Tube pressure drop 04;; psi

$alculated Tube pressure drop =4;; psi

The final tube length needs to be slightly longer than calculated because the calculated surface will be large

a lower tube velocity that gives a lower heat transfer.

T The fluid temperature in o#

=!8+ The fluid&s =pecific gravity Z 8+o

#

;4;=8 6specific heat9048 /

-ouling 2 +rt6a(g4 tube (iscosity908; 266a(g4 shell (iscosity9;4=1905; /

#ouling is the total for both sides. The above is limited to a *axi*u* (iscosity of 7 c) for the tube side.

g 6Allo&ed pCalc4 p907 /

)llow.∆ p

$alc. ∆ p



Ho& to Calculate Excess +urface and "(erdesign +urface

Excess surface / 70483

0here

actual heat transfer surface =8;4;surface calculated from design overall heat transfer coefficient 0<;4;

>se +uperficial (elocities to Calculate Best Heat Transfer -lo& )attern

The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin li-uid fil

little vapor in contact with the heat transfer surface. How do you tell if the flow is annularG %t will be whe

superficial gas velocity is above the following value

%f the superficial li-uid velocity is below +.+ ft3s

5047 ft3sec.

where

@: the superficial li-uid velocity ;4=8

%f the superficial li-uid velocity is above +.+ ft3s

814= ft3sec.

where@: the superficial li-uid velocity 04;;

D Euation -or Heat Transfer Coefficient ,nside Tubing

#or <eynolds numbers below 6+,+++ there is an :3 effect on the heat transfer coefficient inside tubing. %f

full tube length for :, you may be too conservative. There will be turbulation at the tube entrance before la

is fully developed. The turbulent length needs to be subtracted from the full tube length. Bse the followin

sies 6.+ inch or less.

00 feet

0here

: variable to use in :3 expression, ft

Tube :ength length of tube, ft =; feet

;4:8; inches

<e <eynolds number 8;;;

0;; x 6Aactual

IAcalculated

9 Acalculated /

Aactual

/

)calculated

To calculate over(design surface use the clean overall heat transfer coefficient for )calculated

.

g.ax

/ 1= I 053 20;; = /

ft3sec. 6less than ;47; fts9

g.ax

/ =340 2 =3 2 040= = /

ft3sec. 6*ore than ;47; fts9

: Tube :ength ( +.++5D i <e

i tube %.., in



.TD Correction -actor Charts for TE.A G and J +hell Types

There are :MT correction factor charts in TEM) for a single type ! shell and two in series of type I shel

#or charts of more shells in series, refer to the enclosed ale !ulley(generated charts in this 0orkbook.

Di(ided -lo& .TD Correction 4ovember, 6CC8

=omething to watch out for is the :MT correction for ivided #low =hell Tube Exchangers. ivided

>shell type I? does not have the same correction as the usual flow pattern >shell type E?. 0e have seen seve

instances lately where a thermal design program made this correction factor mistake. True, there is very litt

difference at correction factors above +.C+. However, there is a difference at lower values. #or example

+helltype -lo&

E-ual outlet temperatures =hell type "E" +.9+*

=hell type "I" +.DD*

+.D8*

=hell type "I" +.8*

$ontact us if you do not have :MT correction factor charts for divided flow. TEM) has one chart for a s

shell but it gives high values for the above examples and it is hard to read in this range. <efer to the enclos

!ulley charts for up to 7 shells in series that are found in this 0orkbook.

o&est i*it of .TD Correction -actor

0hat is the lowest :MT correction factor to be usedG Here is what several literature sources say

Heat Exchanger Design Handbook 6HEDH9%

[# should be kept above +.D* to +.9+S

)erry's Che*ical Engineers' Handbook%

[@alues of # less than +.9+ >+.D* at the very lowest? are generally

unacceptable because the exchanger configuration chosen is inefficient ...S

%n over *+ years of experience, a correction factor of +.D* is the lowest we have seen a thermal designer us

)lthough there was one case where an operating shell(and(tube heat exchanger reflected a lower :MT co

factor than +.D*. )nother way of looking at the correction factor is to never use a temperature cross of mor

* degrees # in a single multi(tube pass shell.

.ini*u* -lo& Area -or +hell +ide ,nlet $ole

#or single phase li-uids and no impingement plate

Minimum area >#low >lb3hr? x +.+7? 3 >9.D =-.<oot>Q?? ;4=:0

=hell =ide #low <ate =;;; lb3hr

Correction -n

$old outlet * o# higher than hot outlet =hell type "E"

in5



=hell =ide fluid density :=48;

#or boiling li-uids and no impingement plate

Minimum area >#low >lb3hr? x +.+7? 3 >55.8 =-.<oot>Q?? ;4587

Ho& to Calculate the )erfor*ance of Heat Exchangers &ith )lugged Tubes

6. Bsing the actual overall heat transfer coefficient >B?, calculate the heat transfer resistances that excludes

tube side resistance

. $alculate new B

7. $alculate a new heat load from new surface and a new B

Ho& to ,ncrease Heat Transfer for o& #eynolds $u*bers =eptember, 6CCC

%f pressure drop is available and if the tube side <eynolds number is less than *,+++ and more than 6,+++, y

can probably increase the heat transfer considerably by increasing the number of tube passes and using sho

This will not only increase the tube velocity but there will be a lower :3 correction. 'oth of these factorsincrease the heat transfer.

Calculate When to >se +eal Bars on a Bundle to ,ncrease Heat Transfer

1ne of the fluid by(pass streams that lower the shell(side heat transfer is the stream that flows around the b

To evaluate, calculate a heat transfer variable named #='. %t is the ratio of the by(pass to the cross flow a

The by(pass area is normally

#=' 's >s ( 1T:? ;4;85

's>s(1T:?N's>1T:(o?>(o?3

0here

s inside diameter of shell =74;; in.

1T: The 1uter Tube :imit, or outer diameter of the tube bundle ==418 in.

's 'affle =pacing 034;; in.

o tube 1 04;; in.

tube spacing 04=8 in. Typical value is 6.5* x tube 1

lb3ft

in5

< other

63B (63hio

5. $alculate new hio and new surface using usable number of tubes

Bnew

63>63hio N <

other ?



Calculate Tube Bundle Dia*eter Ianuary, 5+++

#ollowing are e-uations for one tube pass bundle diameter when the tube count is known or desired

#or tubes with + eg. itch

= 6.+*5 x pitch x =J<T>count? N tube 1.. 014735 inches

#or tubes with C+ eg. itch

= 6.6 x pitch x =J<T>count? N tube 1.. 034:01 inches

0here

Tube 1 ;418; inches

= 'undle diameter

$ount =8; 4umber of tubes

itch Tube spacing 04;;; inches

Tube Count Calculation for + T )ugust, 5++5

%f you don&t have a tube count table for a shell and tube exchanger, the tube count can be calculated. The fo

e-uation is good for any sie tube on any tube pitch. %t is primarily for situations where there is not a need

allowance for bundle entrance and exit area.

$ount 71= Tubes on

0here 5=3 Tubes on

o Tube 1.. ;418; inches

# 04;; for s-uare pitch# 0408 for triangle pitch

4pl 4umber of tube pass lanes >6 for two pass? =

:w Tube pass lane width >typical is +.85* inches? ;4:=8 inches

Tube pitch 04;;; inches

T$ >'undle diameter ( tube 1..? ==4=8; inches

#or tube pass lane width for s-uare rotated tube pitch use >6.767 \ o?. The decrease in the number of tu

due to bundle entrance and exit area could be allowed for by using a larger :w.

Tube Wall Te*perature for Cooling Water Ianuary, 6CCC

0hen designing heat exchangers where hot process streams are cooled with cooling water, check the tube

temperature. Hewitt says that where calcium carbonate may deposit, heat transfer surface temperatures abo

before it enters the water(cooled exchanger.

%f #=' is more that ;408, then seal bars are needed.

# ]+.D9*7 x T$5 ( >:w N o ( ? >T$ x 4pl?^ 3 5

67+ o# should be avoided. $orrosion effects should also be considered at hot tube wall temperatures. )s a

rule of thumb, make this check if the inlet process temperature is above 5++ o# for light hydrocarbon li-uid

++ ( 7++ o# for heavy hydrocarbons. $onsider using )ir coolers to bring the process fluid temperature do



+o*eti*es arger Tubes are Better 1ctober, 6CC9

There is an exception to the rule that a shell and tube heat exchanger service using 37 inch tubes will be ch

than one using 6(inch tubes. This is when the tubeside has a much lower heat transfer coefficient than the

of the tubes and the following conditions are present

The flow will be in laminar flow if two >5? tube passes are used.

%f four >7? tube passes are used, the tubes in the 37 inch selection will have to be significantly shorter than

in order to meet pressure drop. 1n the other hand, the 6(inch tube design uses the full allowable tube lengt

Weighted .TD

%f there is more than a slight curvature in the heat release curve, things get more complicated. Then a step(

method using local temperatures and local heat transfer coefficients are used to calculate the heat exchange

The -uestion is what do you report as the MT and the correction factorG There is a reference in TEM) in

temperature relations section T(.5 that refers to a weighted MT. The article mentioned was published byale !ulley in the Iune 6C88 issue of Hydrocarbon rocessing. The article shows how to calculate a weig

MT and its correction factor if one is re-uired.

Esti*ate "pti*u* -lo& elocity for Gas ,nside Tubes

=ince the design of heat exchangers is a trial and error solution, a good starting point is desired.

Bsually the design starts with an estimated overall heat transfer coefficient. %f you don&t know a good starti

coefficient the e-uations presented here give this starting point with simple e-uations.

%n the design of heat exchangers using up the maximum allowable pressure drops gives the highest heat tra

The e-uations below estimates the tube velocity>0?for a gas that will meet the maximum allowable pressu

#rom 0 you can calculate the tube count or heat transfer coefficient. #or a given tube length the following

tube velocity for turbulent flow. !ases will be in turbulent flow more than CCL of the time. %f your calculat

what the following e-uation calculates, you need more tube travel where tube travel is in the form of numb

length>s? for countercurrent flow. These e-uations can be used for two phase flow as long as the two phase

#or 37 inch tubes with +.+8 tube wall

#or 6.+ inch tubes with +.+8 tube wall

0here

: total tube lengths in ft.

>)dd ]9 x tube % in inches^ ft for turning losses for each tube pass?

0 lb3hr3tube

O allowable pressure drop inside tubes in psi >deduct 6*L for nole pressure drops?

Q density in lb3cu.ft.

0 68++>OQ3:?+.***

0 *++>OQ3:?+.***



: 56 ft #or 37" tubes, 0 05<1 Tube count

O D psig #or 6" tubes, 0 7=15 Tube count

Q 5.88

Mass flow 6C*+++ lb3h

Example

Bse 37 inch tubes and 68 foot tubes. The maximum allowable pressure drop inside the tubes is D psi >after

gas density is 5.88 lb3cu.ft. The tube side flow is 6C*,+++ lb3hr. 0hat should be the starting tube countG

=olution

0 67CD lb3hr3tube

Tube count 6C*,+++367CD 6+

#or a shell(and(tube heat exchanger, calculate the shell diameter when given the tube count here

Esti*ate Hydrocarbon Gas Heat Transfer Coefficient in +hell +ide

%ts difficult to estimate a gas heat transfer coefficient in the shell side because of the many variables. The fo

Ho 7+.$p>O3: x Q?63 6D 'tu3h ft5 Y#

where

$p specific heat >'tu3lb(#? +.6*

: tube length >ft? 6+

O shell side pressure drop >s 5

>subtract nole losses?

Q density of gas >lb3ft? +.+9*

lb3ft

0 68++>D x 5.883>68N*??+.***

$alculate = T diameter from number of tubes



example

er than

lection

llowing

te

bing

!ravity



ic heat.

r coefficient

alue.

;4;=;

80

ption is

he

r due to

There is no



and

the

ou use the

inar flow

for tube



s.

low

al

le

ngle

ed ale

.

rrection

e than



$alculate the expected performance of an exchanger that has had to have some tubes p

the 6. Xou know the original overall heat transfer coefficient for the un(plugged exchang

number of tubes plugged.

5. Therefore, you know the original heat transfer area, the original hi and ho, the orig

and the original duty and terminal temperatures.

. Xou want to know what will be the new duty capacity and terminal temperatures w

unit operating with plugged tubes.

)fter a heat exchanger goes into operation it may develope leaks in the tube walls.

The following procedure calculates the new heat load and new overall heat transfer co

1. Using the actual overall heat transfer coefficient (U). calculate the heat

ou

ter tubes.

will3. Calculate new U

4. Calculate new heat load from new surface and new U

ndle.

ea.

Rother = 1/U -1/h

io

2. Calculate new hio and new surface using usable number of tubes

Unew = 1/(1/h

io + R

other)



llowing

or

-uare itch

riangular itch

es

all

ve

rough

and

n



eaper

utside

allowed

.

ise

area.

the

ted

g value for this

sfer for single phase fluids.

e drop.

e-uation gives the optimum

d tubeside velocity is below

r of tube passes or total tube

viscosity is less than +.+6* cp,



07;

:;

nole deduction? and the

llowing will give you a value within 5*L.



ged by doing the following

nd the

tubeside velocity

the

icient.

nsfer resistances that exclude the tubeside resistance



A(oid +*all Baffle Cuts in + T Condensers Iuly, 5++6

There will be a theoretical li-uid level when there is condensation in a heat exchanger. The condensing heat

coefficient decreases as its& li-uid film increases. #or best heat transfer the li-uid level should be low as poss

=mall baffle cuts in a shell and tube exchanger will hold a higher li-uid level than large cuts. Bse a separated

model e-uation system to determine the theoretical li-uid level. Bnless you want subcooling, do not use a ba

that would hold a li-uid level higher than the theoretical one.

Esti*ate Condensing Heat Transfer Coefficient for Hydrocarbons ,nside Tubing

Cond h / 3=3

0here

$ond h %nside condensing heat transfer coefficient

0 $ondensing fluid in tubes 18;4;; lbs3hr3tube

.axi*u* Condensing #ate ,nside Tubes )ugust, 5++6

#ollowing is a close estimate of the maximum heat transfer rate for total condensation. %t is based on the max

Hi 70<7 'tu3hr

0here,

li-uid thermal conductivity of the condensat ;478;

#or example this e-uation yields a maximum heat transfer rate for steam to be ,8++.

Kuick Esti*ate for #eflux Condenser .TD in Aircooler

curve has a hump which will give a :MT higher than one calculated from a straight line condensing plot. )

e-uation that makes a -uick estimate for the :MT is

=tandard :MT x #actor

Then :MT #actor 6.7 (+.++C5 >T (66+?

#eflux 6L@nock backM9 Condenser Iune, 5++6

o not design this like the usual vertically condensing heat exchanger where both gas and li-uid flow in one

%n this type of condenser, the coldest condensate will be in contact with the entering hot vapor >in the bottom

4early everything about this type condenser is different. %t is both difficult to design and difficult to control.

flow patterns, pressure drop and heat transfer calculations are different. 'e sure the heat transfer calculations

654089 W;43 / 'tu3>hr?>ft5?>o#?

condensing rate for the average hydrocarbon to be D*+ 'TB3hr(ft5(#. %t is good for other types of chemical c

18; 6@ li

;4;19;4< /

; li-

'tu3hr(ft(o#

This type of service has steam condensing out from a non(condensable gas which is mostly $15. The conden

%n the case of outlet process temperatures below 6*.* o#,

0here T outlet temperature and air inlet temperature is 6++ o#.



Types of +tea* Condensers

=mall steam condensers use shell(and(tube heat exchangers while large steam condensers use surface conden

lower temperatures, a [PS type shell can be used. ) point is reached where the sie or operating pressure re-

a surface condenser.

+ulfur Condenser Tube elocity i*its

#or good operation of a sulfur condenser the design velocities inside the tubes should be within certain limits

velocity range is between 6.* and 8.+ lb3s- ft(sec. 'elow this range there will be slugging. )bove this range

fogging will occur..

+*all Te*perature )inch )oints in Condensers 4ovember, 6CC9

'e extra careful when condensers are designed with a small pinch point. ) pinch point is the smallest temper

difference on a temperature vs. heat content plot that shows both streams. %f the actual pressure is less than th

process design operating pressure, there can be a significant loss of heat transfer. This is especially true of flu

that have a relative flat vapor pressure plot like ammonia or propane. #or example %f an ammonia condenser

can be a 68L drop in heat transfer.

When to +lope +ingle Tube )ass Tubes in Condensing +er(ice Ianuary, 5++5

)t low vapor velocities, it has been proven that even a slight downward slope of tubes gives a significant inc

heat transfer in the case of tube(side condensation. 'ut this does not mean the larger the slope the higher the

have the tubes sloped is when they are operating near atmospheric pressure and there is one tube pass. )n ex

of this is a sulfur condenser. %t has a low pressure drop usually less than +.* psi. They typically are designeda slope of 639 inch per foot of tubing.

Fone Those Condensers

The heat transfer and pressure drop of a condenser usually should be oned. ) typical heat exchanger that co

6++L of the vapor will go through 5 or different flow pattern ones before the flow becomes a li-uid. Ther

better accuracy if the flow patterns are determined and their individualistic e-uations are used.

) conventional [ES type shell is used when the steam condensing temperature is above approximately 65+ o#

designed for 57D =%) operating pressure and the actual pressure is * =% less and the pinch point is 9 o#, the

transfer. The benefit of sloping stops at an angle of approximately 6+o. ) common case of a condenser needi



fer

.

w

cut

m

ction.

ion?.

oned.

ounds.



.

s

e

ur

e

in

le

ses

r

o



#otated +uare Tube )itch

=ome heat exchanger specifications for shell and tube heat exchangers mention s-uare pitch but do not speci

mention rotated s-uare pitch. Engineers with little thermal design experience who are trying to strictly adher

specifications may re2ect this type of tube pitch. The benefits for this type of tube pitch sometimes get lost be

of this. <otated s-uare pitch gives better mixing of the shell fluid and better heat transfer for the heavier flui

#re-uently the shell sie can be reduced when there will be heavier li-uids on the shell side and the designer

rotated s-uare pitch.

Caution When >sing a ongitudinal Baffle in the +hell +ide

The following are potential problems when considering the use of a longitudinal baffle in a new = T heat e

is lost due to conduction across the long baffle. $heck and make sure this has been taken into con

%f the long baffle is not welded to the shell, the pressure drop across the long baffle is more than D

This will also lose thermal efficiency. The seal on the long baffle should be tested in the shop afte

>sing Turbulators for Tube +ide a*inar -lo&

%f the flow inside the tubes of a heat exchanger is in laminar or viscous flow, take a look at enhancing the hea1ne simple and inexpensive device is the twisted(tape insert. Bsing twisted(tape inserts for laminar flow in n

exchangers results in cost savings and smaller heat exchangers. Twisted(tape inserts can be used in existing h

exchangers to make a significant increase in capacity. The amount of increase in heat exchanged depends on

the increase in pressure drop can be tolerated. %f there is no pressure drop limitation, there can be as much as

increase in capacity.

Here are the recommended guide lines for using twisted tape inserts

6 ressure drop in the tube side without inserts is less than to 7 =%.

5 Minimum fluid viscosity of 5 centipoise unless there is a very low velocity

Bse a minimum tube diameter of *39S for .++6 fouling. Bse a minimum of 6S diameter for

+.++6* fouling. %t is not recommendable to use turbulators in a service that has a fouling

factor greater than +.++6*.

These guidelines for tube diameter are due to fouling being more of a problem with turbulators in small tubes

Triple +eg*ental Baffles 4ovember, 6CCD

There is more than one kind of triple segmental baffles in the shell side of heat exchangers. 'e sure you know

kind if you are checking a design that uses them. There is the kind you see in TEM) where there are three digroups in a set. The total number of baffle pieces is six. There is the kind that is like producing two double s

streams in parallel. There are two groups in a set and a total of five baffle pieces. )nother kind has only thre

in a group and each piece has a different shape.

Entrance and Exit +pace for +hell $oles Ianuary, 5++6

There have been cases where not enough space was under the shell noles. This can be critical for applicati

a horiontal thermosyphon or other pressure drop sensitive applications. $heck the distance from the nole

the nearest tube row or impingement plate. %f there is an impingement plate this distance should be U or mor

6. The largest temperature drop across the long baffle is more than 5*+ o#. Then the thermal efficien

5.



nole %.. %f there is no impingement plate this distance should be 638 or more of the nole %.. %f pressur

not a consideration and TEM) re-uirements are met and vibration is not a problem then the above calculated

could be reduced. This criterion naturally doesn&t apply to shells with distributor belts or where the nole is

the back of a B(tube bundle.

#or information on calculating shell nole pressure drops, refer to [$alculate =hell 4ole ressure ropS i

the calculation Tab of this 0orkbook.

Horiontal (s4 ertical Baffle Cut in + T Exchangers May, 5++6

%n shell and tube heat exchangers it is safer from a thermal design standpoint to use vertical baffle cuts but ho

cuts have an advantage in certain situations. Horiontal cuts are best if the shell side stream is clean and sing

There will be less of the shell side stream bypassing through the tube pass lanes. =ince in a multi(tube pass e

there will be more horiontal tube pass lanes than vertical pass lanes, you need to flow perpendicular to these

lanes for minimum by(passing of the shell stream. This means horiontal cut. 0here you do not want to use

cut is when there is either condensing or where there is the possibility of foreign material being in the flowing

%t is suggested to use a maximum fouling factor of +.++5 for horiontal baffle cut. %t may be possible to use h

cut in certain boiling applications.

,s an Expansion Joint #euired in the +hell? ecember, 6CC9) fixed tube sheet exchanger does not have provision for expansion of the tubing when there is a difference i

temperature between the shell and tubing. 0hen this temperature difference reaches a certain point, an expan

2oint in the shell is re-uired to relieve the stress. %t takes a much lower metal temperature difference when th

metal temperature is hotter than the shell metal temperature to re-uire an expansion 2oint. Typically, an all st

because the maximum allowable tube $ompressive stress has been exceeded. )ccording to the TEM) proce

for evaluating this stress, the compressive stress is a strong function of the unsupported tube span. This is no

twice the baffle spacing.

,ncreasing Capacity of Existing +hell Tube Exchangers March, 6CCD

To increase heat transfer check out using low fins or other special tubing. 0hen an increase in capacity will

excessive pressure drop, you may not have to 2unk the heat exchangers. %nvestigate the relatively inexpensiv

modification of reducing the number of tube passes. 1ther possibilities are arranging multiple exchangers in

ocating ents on the +hell +ide of ertical Exchangers Iuly, 6CC9

roper venting of e-uipment is not always given the consideration it deserves. 1ne place where venting is es

a problem is underneath the tubesheet of a vertical exchanger. The problem is that there will always be a spa

the vent connection to trap gases or vapors. 'esides the poor heat transfer in this region, this can cause corro problems. %t is important to get the vent connection as close to the tubesheet as possible. Bsing multiple con

that are smaller is one solution. )nother solution is to fabricate the upper tubesheet with a small vent tunnel

-lange Gasket ocation May, 6CCC

There is an optimum diameter of the gasket for flanges. %t is when the total 1perating moment of the flange u

pressure is e-ual to the gasket seating moment. #or low(pressure flanges, the diameter should be as close to

circle as possible. #or high(pressure flanges, the diameter should be as close to the flange %.. as possible. %

case, low pressure is considered to be below ++ psi. High pressure is considered to be approximately D*+ p

exchanger can take a maximum of approximately 7+ o# metal temperature difference when the tube side is th

0hen the shell side is the hottest, the maximum is typically 6*+ o#. Bsually if an expansion 2oint is re-uired,



and higher.

>sing #ods for Tube ,nserts to ,ncrease Heat Transfer )ugust, 6CC9

Bse concrete reinforcing rods inserted inside the tubes to increase the heat transfer and tube velocity. %t is a -

and economical solution. This is usually done only in clean services. ) typical case is using 39" rods inside

37" x 67 '0! tube. The tube side heat transfer coefficient is increased by a factor of 6.D. However, you ha

be able to stand the increase in pressure drop. %t goes up by a factor of C.*. )nother example is a 6.+" x 68 '

avg. wall tube where the heat transfer goes up by a factor of 6.6D and the pressure drop by a factor of .*.

+hell +ide ,*pinge*ent )rotection

There may be tube vibration or erosion if the shell(side fluid velocity is above a maximum value. These valu

found in TEM) section <$'(7.86 7.85. %n the eighth edition the maximum values can be found on page

The most common impingement protection is a plate baffle that is slightly above the tube bundle. 'ut this ty

protection has some drawbacks. %t has a relatively higher pressure drop than most other methods and the tub

the first several rows tend to vibrate. 1ther types of impingement protection are

6. late within a nole enlarger 5. =olid rods instead of tubes for the first 5 or rows.

. =nap(on tube protectors on top of the tubes in the first 5 or rows

7. =mall angle iron types setting on top of the tubes in the first 5 or rows

*. @apor belt

+pecial + T Exchanger Type 6$T,W9 =eptember, 6CC9

) shell tube heat exchanger with normal segmental baffles has tubes that miss every other baffle. This can

to long unsupported tube lengths for some applications. ) long tube span has a low natural fre-uency and is

to vibration. 1ne solution is to design a [no tubes in windowS >4T%0? exchanger. This design has no tubes

baffle cut out. 'y using intermediate supports between baffles, the natural fre-uency of the tubes can be rais

considerably to resist vibration.

When to Consider Bypass +trips in + T Bundle

Bse a by(pass strip if tubes are removed under a nole. <emoving tubes leaves an open area where the shel

can flow either over or under the bundle.

$onsider by(pass strips if the bundle to shell clearance is more than 37 inches and the shell fluid is mostly se

heat transfer.

Especially consider by(pass strips if the shell li-uid is a hydrocarbon with an average viscosity greater than 6

centipoise and the tube fluid has a high heat transfer coefficient >example water?. %n this case, a * to 6+L inc

in heat duty can be achieved by installing by(pass strips.

What is too arge a Te*perature Change in = Tube )asses? ecember, 6CC8

0arningK :arge tube side temperature change. ) big difference between the inlet and outlet temperature of t

tube side causes leakage and bypass problems. The worst case is a shell and tube exchanger with two >5? tub

passes where a gasket is used to seal between the passes. ) careful analysis should be made if the temperatur



difference across the pass plate is more than ++ o#. #or a channel type that has a welded in pass plate, make

analysis if the temperature difference is more than 7*+ o#. %f this temperature difference causes an over stres

condition, possible cures are

)dd a unit in series so each unit has a smaller temperature difference_

Bse one tube pass if the penalty isn&t too great_

#or air coolers, use a split headers design.

#otated +uare Tube )itch #ebruary, 6CCC

=ome heat exchanger specifications for shell and tube heat exchangers mention s-uare pitch but do not speci

mention rotated s-uare pitch. Engineers with little thermal design experience who are trying to strictly adher

specifications may re2ect this type of tube pitch. The benefits for this type of tube pitch sometimes get lost be

of this. <otated s-uare pitch gives better mixing of the shell fluid and better heat transfer for the heavier flui

#re-uently the shell sie can be reduced when there will be heavier li-uids on the shell side and the designer

rotated s-uare pitch.

ongitudinal Baffle Heat Conduction Cures

0ith a longitudinal baffle and a long temperature range there can be a problem with heat conduction through

longitudinal baffle. There will be a loss of thermal efficiency due to the heat conduction.The longitudinal baffle can be fabricated in one of two ways.

6. :eaving an small enclosed air gap between two longitudinal baffles.

5. =pray an insulating material like <yton on the longitudinal baffle.

Design Te*peratures of Carbon +teel and o& Alloy Tubes and Tubesheets

Bse the higher of the shell(side and tube(side design temperatures up to 8*+ #.

)t higher design temperatures use the arithmetic average of the 5 design temperatures.

Design Te*peratures of $onferrous Tubes and Tubesheets

0ater in the shell(side

0ater in the tube(side

•

•

•

Bse the arithmetic average of the shell(side and tube(side design temperatures.

Bse the higher of the tube(side design temperature or tube(side outlet temperature N 63 of the :MT.



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Choosing -in +pacing Iune, 5++5

%n waste heat applications, the fin spacing depends not only on the heat transfer but the cleanliness of the exh

%f the gas is fouled from soot or other fine particulates, use a maximum of * fins per inch. #or very dirty gase

fin spacing can be as low as 5 fins per inch. Bsually there will be soot if fuels heavier than diesel fuel are fir

The designer needs to know the source of the waste heat gas so that he can make a decision on what fin spaci

H#+G $ole +ie )pril, 5++5

#or an estimate of the nole sie entering and leaving a H<=! unit use

D / 5457 inches

0here

diameter of nole

#low !as flow 0;;; lbs3hr

This is based on a total of +.9 inches of water

-ace Area for H#+G >nits )pril, 5++6

The starting point in the design of a heat(recovery steam generator >H<=!? is the face area. This will determ

the preliminary duct dimensions and starting face areas of any economiers and superheaters.

-ace area / 6-lo& =8;;9 / ;45;

0here

#low Exhaust !as flow 0;;; lbs3hr

0here face area is in s-uare feet.

.axi*u* Exhaust Gas Te*perature for +teel -in Tubes

Here is an approximation of the maximum exhaust temperature for steel fin tubes when generating steam.

1therwise, the fins would need to be the more expensive 7+C == material. This is based on the typical 5 inch

tubing with 6 inch fins and 8 to D fins3inch.

.axTg / 0;<; ;4=7 Bte*p / 0;=0

0here

MaxTg maximum gas temperature'temp water boiling temperature 7;;

When to >se Bare Tubes in Waste Heat Boilers

;405 x 6flo&90= /

ft5

This is based on using 5 inch 1.. tubing with 6 inch high fins. The tubing is arranged on 7 639 inch triangu

o#

o#.

Bse bare tubes if the bundle is -uite small or the gas temperature is greater than 6,*+ to 6,7++ o#.



st gas.

he

.

to use.

e

..

pitch.



Cooling Water -lo&ing ,nside 7;5++ >tubes Iune, 6CCC

4ormally it is 1; to use +7== when cooling water with low chloride content is flowing inside B(tubes. 'u

some reason the operating pressure drops to saturation there can be corrosion problems. The tube vibration t

results from the flashing of steam amplifies the stress that causes stress corrosion cracking.



f for



Calculating -ouled )ressure Drop )ugust, 6CCC

There are various ways to account for fouling when calculating pressure drop.

1ne way would be to add a small amount to the tube diameter. This has a complex effect that is not linear in

) simpler method is to add 6+L for each +.++6 increase in fouling factor. Then multiply this factor by the cl

pressure drop. Xou would use a pressure drop factor of 6.5 for a fouling factor of +.++5.

Allo&able )ressure Drop +uggestions March, 5++5

%f you are at a loss as to what allowable pressure drop to specify, here are some suggestions

-luid and Condition Allo&able )ressure Drop psi

Gas to *

iuid 9 to 6+

Change of phase

Boiling% +.* to 6.+ psi for greater than 6+ L vapor

6.+ to *.+ psi for less than 6+ L vapor

Condensing operating pressure:ess than atmospheric +.*

)tmospheric to 5* psi 6

5* to *+ psi 5

*+ to 6*+ psi

6*+ psi N .+ to *.+

Allo&able +hell +ide )ressure Drop if a .ultileaf 6a4k4a4 a*aflex9 ong Baffle is >sed

#our thin >+.++9S? stainless strips are normally used to seal the sides of the long baffle. 'ecause of their flex

they are not able to withstand large shell side pressure drops. %t is best to limit the pressure drop to * psi with

D.* psi being the maximum.

Better Baffle Windo& )ressure Drop Euation

) new baffle window pressure drop e-uation has been published in the Iune 5++7 issue of Hydrocarbon roc

The name of the article is [More )ccurate Exchanger =hell(=ide ressure rop $alculationsS. The article ca

found on this page with the sub2ect [Heat Exchanger )rticles ublished by ale !ulleyS. The e-uation impr

accuracy of the shell side pressure drop. <efer to the article for more detail. The e-uation has the following f

fi #riction factor for ideal tube bundle

#or + deg. Triangular, 5.5_

#or C+ deg. =-uare, .87_

#or 7* deg. s-. rotated, 5.5C_

#or 8+ deg. Triangular, 6.DC estimated.

; ressure loss coefficient for velocity head e-uation

$6 $onstant based on the type of tube layout



istortion factor for ideal fluid stream. %t varies with baffle cut. <efer article elsewhere on this si

e-uation. >'affle cuts from 57L to 5CL >fractional? have a distortion factor of 6.+?

ENA.)E%

This is taken from the first experimental case in [) <eappraisal of =hellside #low in Heat Exchangers HT(@

)verage flow of CC+,+++ lb3hr with a density of 85.7 lb3ft is flowing through a 6.5* % nole. The shell %5.5* in. and the 1T: is 55.D* in. The tube 1 is +.D* in. on a tube pitch of +.CD* in. with + degree layo

There are D baffles and 58L baffle cut.

The following are taken from a tip in this section named [%mprove =hell =ide ressure rop $alculationsS

fi +.6+5*

4cw *.C8

=l 66.+

=w 77.7D

$6 for a + degree layout is 5.5 6 since the fractional baffle cut is 58L

;p 6.

!w >CC++++ x +.+7?377.7D 9C+.* >3s-ft(sec?

Ow 65.D9

Designing Better >se of Tube )ressure Drop 1ctober, 6CCC

0hen the calculated pressure drop inside the tubes is under(utilied, the estimated pressure drop with increas

number of tube passes is

0=4; psi

0here

revious ressure drop 048 psi 4)== 4ew number of tube passes 5

1)== 1ld number of tube passes =

This would be a good estimate if advantage is not taken of the increase in heat transfer. =ince the increased n

of tube passes gives a higher velocity and increases the calculated heat transfer coefficient, the number of tub

be used will decrease. The use of fewer tubes increases the new pressure drop. #or a better estimate of the n

pressure drop, add 5*L if the heat transfer is all sensible heat.

4cw

Effective number of tube rows crossed in baffle window

=l Total of leakage areas >in5?

=w 4et flow area in baffle window >in5?

;p +.6+5* > >5.5 x*.C8? (5>66377.7D?5? ?

Ow ;p x +.+++6+9 x !w53Q

Ow 6. x +.+++6+9 x >9C+.*?5

x D385.7

$e& tube ) / ) x 6$)A++")A++97 /

∆



Bsually when shell(and(tube heat exchangers are designed, the tube layout is made so that the shell entrance

approximately e-ual to the shell nole flow area. The average distance to the 6st tube row is n37 where n

inside diameter of the shell nole. %n this case the pressure loss coefficient is 6.+ for the pressure drop calcu

for the shell nole entrance.

%f the shell noles are greater than 5S and some tubes are not omitted from the tube layout, the nole entran

pressure drop can be significantly higher than the normal calculation based on the nole flow area. %n a case

8S shell nole and where no tubes were omitted in a 'EM type heat exchanger, the pressure drop was time

higher than that calculated with 2ust the nole flow area. #or more information, you can refer to the tip [$a

=hell 4ole ressure ropS in this 0orkbook.

@ettle )ressure Drop )pril, 6CCC

Bsually you will see the allowable pressure drop on the specification sheet for the shell side of a kettle reboil

be stated as [nilS. This is close to being true only for the bundle. The inlet and outlet kettle noles will hav

definite pressure drop. %t is best to locate the inlet nole on the side of the kettle and above the bundle. This

keeps the pressure drop down because there are no tubes in the vicinity to provide a restriction.

-ixed Tube +heet Exchanger and High +hell +ide )ressure Drop Iuly, 5+++0hen there is a design problem meeting the allowable shell side pressure drop, reverse the stream sides. =in

a fixed tube sheet exchanger, the unit can be designed with one >6? tube pass. 1ther types of heat exchangers

be designed with a single tube pass but they can have more operating problems. The pressure drop can be fu

reduced by using axial noles that are on the exchanger centerline. This eliminates large turning pressure dr

,*pinge*ent #ods Ianuary, 6CCD

0hen shell pressure drop is critical and impingement protection is re-uired, use rods or tube protectors in top

rows instead of a plate. These create less pressure drop and better distribution than an impingement plate. )

impingement plate causes an abrupt C+ degree turn of the shell stream which causes extra pressure drop.

+pecifying )ressure Drop for Hea(y iuids ,nside Tubes

#re-uently process engineers specify * or 6+ psi for allowable pressure drop inside heat exchanger tubing. #

li-uids that have fouling characteristics, this is usually not enough. There are cases where the fouling exclud

turbulators and using more than the customary tube pressure drop is cost effective. This is especially true if t

relatively higher heat transfer coefficient on the outside of the tubing. The following example illustrates how

pressure drop can have a big effect on the surface calculation. ) propane chiller was cooling a gas treating li

that had an average viscosity of D.* c. The effect on the calculated surface was as follows

* 7,+65

5* 5,6+7

*+ 6,76C

Xou can see that using 5* psi pressure drop reduced the surface by nearly one(half. This would result in a pr

reduction for the heat exchanger of approximately 7+L. This savings offset the cost of the pumping power.

Effect of 0st Tube #o&s on +hell $ole )ressure Drop

)llowable tube pressure drop, psi

Exchanger surfaceft5



.axi*u* elocity ,nside Tubes

)n estimate for maximum tube velocity inside steel tubes

3; srt6density9 / 0;4= ft3sec

0here

maximum fluid velocity

ensity fluid density :=4;; lb3cu ft.

Calculate +hell $ole )ressure Drop

=hell nole pressure drop calculation methods are difficult to find in the open literature. The nole pressur

are difficult to predict accurately. There is a complex flow pattern of a tube matrix, bundle bypassing, and re

'ecause of this, it is possible to have pressure loss coefficients greater than the customary 6.* velocity heads

sharp edge expansion3contraction edges.

%f the bundle entrance area is e-ual to or greater than the inlet nole flow area, use a pressure loss coefficien

%f the bundle exit area is e-ual to or greater than the exit nole area, us a pressure loss coefficient of +.*9. T

indications that it should be larger. The following procedure is for the situation where the nole flow area is

than the entrance or exit area and the bundles do not have an impingement plate. %f there is an impingement p

there will have to be added a turning loss to the calculation below. %f the two shell side noles are not the sa

calculate the inlet pressure drop and take 53 of it and make a separate calculated pressure drop for the outlet

take 63 of it.

+hell Entrance or Exit Area%

6. $alculate the bundle bypass area =b ` x n x h

. $alculate the shell entrance and exit area.>)s?

)s =b N )slot>refer TEM) <!(<$'(7.856 7.855?

7. $alculate ratio of =b to total area #< =b3)s

*. ;n +.8* N5.67 >#< (+.7?

>minimum ;n +.9, maximum 6.9?

>On total of both noles?

where

n 4ole % in.

s =hell % in.

t Tube outside diameter in.

#5 +.D+D for 7* degree pitch, all others use 6.+

h +.*>s(1T:? in.

;n ressure loss coefficient

1T: 1uter tube limit diameter in.

t Tube center to center pitch in.

@s velocity in the entrance3exit area >ft3sec?

*ax

/

@max

5. $alculate the slot area )slot +.D9*7n5 >t (t?3>#5 x t?

8. On ;n x .+++6+9@s5 x density

On Total nole pressure drop >lb3ft5?



ENA.)E

the 1T: is 55.D* in. The tube 1 is +.D* in. on a tube pitch of +.CD* in. with + degree layout.

$alculate =b

h +.*+>5.5*(55.D*? +.7D*

=b ` x 6.5* x +.7D* 69.5

$alculate )slot

$alculate total area )s

)s =b N )slot 69.5 N 5D.*9 7*.96

$alculate #<

#< 69.537*.96 +.7

$alculate ;n;n +.8* N5.67>+.7 (+.7? +.8* >use minimum +.9?

$alculate nole pressure drop

@s >CC++++ x +.+7?3>7*.96 x 85.7? 6.9*

$omment ( Bsing 6.* total pressure loss coefficient and the nole flow area gives only +.56 =%

,*pro(e +hell +ide )ressure Drop Calculations

The shell side pressure drop calculation can be improved by better e-uations for the baffle window and the n

pressure drops. 'oth of these methods can be found elsewhere on this web page.

The baffle window pressure drop in the open literature is a function only of the number of tubes crossed and

velocity in the window. %t does not take into account a friction factor, type of tube pattern or fluid eddies.

0hen there are no tubes removed under the shell noles and the noles are large, using the nole flow are

result in wrong pressure drop calculations.

This is taken from the first experimental case in [) <eappraisal of =hell side #low in Heat Exchangers HT(

5.5* in. and the 1T: is 55.D* in. The effective tube length is 66.D5C ft. The tube 1 is +.D* in. on a tube

of +.CD* in. with + degree layout. There are D baffles and 58L baffle cut

#rom the following the cross flow pressure drop is calculated

's 6D.8 in

fi +.6+5* ( %deal tube bank correlation > I. Taborek?

CC+,+++ lb3hr with a density of 85.7 lb3ft is flowing through a 6.5* in. % nole. The shell % is 5.5* in.

)slot +.D9*7>6.5*5? >+.CD*(+.D*?3>6.++ x .CD*? 5D.*9

On +.9 x +.+++6+9 x 6.9*5 x 85.7 6.+ psi

)verage flow of CC+,+++ lb3hr with a density of 85.7 lb3ft is flowing through a 6.5* % nole. The shell %



4c 6.D*

<b +.*8

<e 7+,57C

<l +.86*

Oc 8.76 psi

Oshell Oc N Ow N On

#rom other tips Ow 65.D9

On 6.+Oshell 8.76 N65.D9 N6.+ 5+.5 psi

Experimental 5+. psi



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-eatures of a $e& + T bundle to #eplace Bundle That ibrated

6. %f possible, design for lower cross flow velocity with special baffles.

5. Make sure that impingement plate is very secure.

. Bse a tube3baffle clearance of 6387.

7. Bse thicker baffles.

*. Bse closer baffle3shell clearance.

8. Bse thicker tubes.

D. %f tubes are low fins, have the tubing bare where it goes through the baffles.

ibration Cure When Designing +hell Tube Bundles May, 5+++

The cure depends upon whether it is flow induced or acoustical type vibration. 'oth types can be cured by u

a lower cross flow velocity across the bundle. To do this, use double or triple segmental baffles. This not on

lowers the velocity but the closer resulting baffle spacing increases the natural fre-uency of the bundle. )not

possibility is to use a [4o Tubes in 'affle 0indowS design. Then you can use as many baffle supports as nec

with very little effect on shell pressure drop.

%f the vibration is the acoustical type, use either + degree triangular pitch or s-uare rotated pitch. The forme

)nother cure is to use a de(resonating baffle. %n a few cases, putting the problem stream inside the tubes wou

Conditions ikely to Cause +hell Tube Bundle ibration May, 6CCD

'undle vibration can cause leaks due to tubes being cut at the baffle holes or tubes being loosened at the tube

There are services that are more likely to cause bundle vibration than others are. The most likely service to c

vibration is a single(phase gas operating at a pressure of 6++ to ++ =%. This is especially true if the baffle s

is greater than 69 inches and single segmental type. )nother service that sometimes causes bundle vibration

in the shell side. 0ater has a relatively higher momentum than other most fluids. Therefore, if extra precaut

bundle design are not taken, a vibration problem can develop later when the exchanger goes into operation.

Cures for ibration in Existing Bundle =eptember, 6CCD

Most flow(induced vibration occurs with the tubes that pass through the baffle window of the inlet one. The

unsupported lengths in the end ones are normally longer than those in the rest of the bundle. #or 37 inch tu

unsupported length can be 7 to * feet. The cure for removable bundles, where the vibration is not severe, is t

the bundle. This can be done by inserting metal slats or rods between the tubes under the noles. 4ormally

only needs to be done with the first few tube rows. )nother solution is to add a shell nole opposite the inle

to cut the inlet fluid velocity in half. #or non(removable bundles, this is the best solution. )dding a distribut

on the shell would be a very good solution but it is expensive.

%f a B(tube bundle has a vibration problem in the bend area, metal slates or rods can be inserted between the %f a slight decrease in heat transfer is not a problem, encircle the B(bends with a band or heavy wire and s-ue

the tubes together.

Best Design -eature to )re(ent Bundle ibration

%n designing a shell(and(tube heat exchanger, use a +o triangular tube pitch if possible. This will lower the v

fre-uency which is a direct function of something called a =trouhal number. The =trouhal number is a consta

of the vortex shedding fre-uency, shell side velocity and tube 1.



The +Y triangular tube pitch has a significantly lower =trouhal number than other tube pitch types. Bsing 'a

for 37 inch tubes on +o triangular tube pitch the =trouhal number is +.56. 'ut for 8+o rotated triangular tub

=trouhal number is +.96.



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Allocation of +trea*s in +hell Tube May, 6CC9

#or those exchangers that need countercurrent flow, the stream with the highest pressure drop is usually best

the tube side. This is true unless the design pressure is so high for the shell side that there would be material

High pressure drop instead of high design pressure is opposite of conventional thinking. %f there are gas stre

both sides with mol. weights about the same and a small temperature difference, put the stream in the tubes w

highest value of the following

>33hr?>3hr?3op. pressure

1therwise, calculate the little more difficult term @el x @el x ensity term for each side and put the stream w

highest value in the tubes.

Heat Exchanger Articles )ublished by Dale Gulley

$opies of the articles are available in .pdf format

A(oid These -luids When >sing o& fin Tubing

0hen a fluid has a high surface tension, the fluid doesn&t readily flow from the gap between the tube fins. Th

resistance and lowers the heat transfer. The types of fluids that are to be avoided are those whose surface ten

above + to 7+ dynes3cm. This includes such fluids as condensing steam, a-ueous solutions with a high L of

amines and glycols.

>se +uperficial (elocities to Calculate Best Heat Transfer -lo& )attern

The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin li-uid film

vapor in contact with the heat transfer surface. How do you tell if the flow is annularG %t will be when the su

gas velocity is above the following value

g.ax / 037=4;

where@: the superficial li-uid velocity 84; ft3sec.

Check iuid Ther*al Conducti(ity at High #educed Te*peratures 4ovember, 5+++

There have been instances where process simulators have given results where the li-uid thermal conductivity

nearly the same as the vapor thermal conductivity when the reduced temperature was still significantly lower

critical temperature. Examine carefully the li-uid thermal conductivity when its reduced temperature is abov

approximately +.D+. Xou may be able to 2ustify a higher conductivity value and thus a higher heat transfer co

by using an independent and reliable correlation for the calculation.

6. "More )ccurate Exchanger =hell(and(Tube ressure rop $alculations", Hydrocarbon rocessing, Iune 5

5. "Troubleshooting =hell(and(Tube Heat Exchangers", Hydrocarbon rocessing, =eptember 6CC8

. "$omputers help esign Tubesheets", The 1il !as Iournal, May 5+,6CD7

7. "$omputer rograms aid esign 0ork", The 1il !as Iournal, Ian. 6,6C8C

*. "How to $alculate 0eighted MT&s", etroleum <efiner, Iuly 6C88

8. "How to #igure True Temperature ifference in =hell(and(Tube Exchangers", The 1il !as Iournal, =ep 6C87

D. "Make This $orrection #actor $hart to #ind ivided #low Exchanger MT", etro3$hem Engineer, Iuly

9. "Bse $omputers to =elect Exchangers", etroleum <efiner, Iuly 6C8+

1= 053 20;; = /



Check )iping Connections When There is >nderperfor*ance Iuly, 5++5

0hen a heat exchanger is installed and it is not achieving the desired heat duty, the first thing to check is the

%s the piping connected to the correct sidesG %t may be piped(up backwards. The worst case is when the shel

has a viscosity more than approximately c and there is no extra heat transfer enhancement inside the tubin

could cause the fluid, when piped to the tube side, to be in laminar flow with its low heat transfer coefficient.

E(aluating a +hell Tube Exchanger -or a $e& +er(ice =eptember, 5+++

The best information to have for a shell and tube heat exchanger is a specification sheet and a full set of draw

%f both are not available, it is better to have the drawings. This is because they are more accurate on the mech

details and they have tube layout details and seal bar information that the specification sheet does not have. 0

most often missing on older heat exchangers are the bundle drawings. %n this case, you need the original spe

sheet. Then you can use its data and simulate the shell side heat transfer and pressure drop by running a therm

design rogram to get a baffle configuration. Then this is used with the new process data to evaluate the new

This procedure will not be as accurate as having the exact baffling but it is the best you can do if this is all yo

to work with.

Check Heat #elease Cur(es for +kipping "(er De&points Bubblepoints Ianuary, #re-uently process engineers specify tabular heat release data that skips over dew points and bubble points.

increments of heat load or temperatures are used, chances are that the dew points and bubble points will be m

%t is important that the heat content at dew points or bubble points be shown.

When Will Exchangers With o&fins be .ore Econo*ical Than Exchangers

With Bare Tubes?

6. %f the shell sie is a least 5 sies smaller >pipe sie?.

5. %f the shell sie is at least 67" 1..

. %f there are fewer exchangers. when using low(fins

7. 0hen >total shell resistance3total tube resistance? is greater than +.7

Excess Heat Exchanger +urface )roble*s =eptember, 5++5

Excess surface does not always mean being safe. %t can lead to control problems, pulsations, or freeing of c

@aporiation services and reboilers can particularly be a problem. rovide a way to control the flow of the he

medium in a new plant. %n an existing installation without control, the boiling temperature difference may be

that there is complete flashing of the li-uid into vapor. Then the li-uid feed rushes in to replace it which resu

pulsations that may give downstream problems. The -uickest solution is to either plug the tubes or put an ori

the outlet vapor line to restrict the flow.

)urchasing +hell Tube Exchangers March, 5++6

%t is to the benefit of purchasers of shell and tube heat exchangers to not insist on applying their design. %f th

exchanger is to be built to TEM) re-uirements, it will void the guarantee. The last line of paragraph !*.5 sa

thermal guarantee shall not be applicable to exchangers where the thermal performance was made by the purc

.ini*u* elocity inside Tubing for +lurries

The minimum velocity for slurries inside tubes for shell(and(tube is 7 ft3sec. This is for a fine material like a



#or slurries there is a special <eynolds number used for calculating the settling velocity. #or more informatio

slurries, refer to chapter $66 in the piping handbook.

+uggestions for o&-ins and )otential + T Bundle ibration May, 5++5

Tube bundles are more likely to vibrate if there is not a close clearance between the tubes and baffles. :ow(f

more susceptible to vibration because of the valleys between the fins. )nother factor that makes them suscep

that some low(fins are manufactured with the fin 1.. smaller than the bare ends. =ome suggestions if the d

software shows that the bundle may vibrate are

+hell Tube or .ultiTube? Iune, 6CCD

%t is best to use Multi(tube >Hairpin? Exchangers instead of =hell Tube when

6. Xou re-uire a small surface >less than 7++ s-uare feet?_

5. There is a temperature cross in the heat transferred in a =hell Tube_. The li-uid flows are less than 6*+,+++ lbs3hr_

7. 4atural gas flows less than 6,5++ P =-. root>oper. pressure?

Ther*al E(aluation of ong Baffles )ugust, 6CCD

The two thermal design problems associated with using two shell passes and a longitudinal baffle in =hell an

heat exchangers are

6 Heat conduction through the baffle. There is a calculation method by 0histler.

%t is a correction applied to the :MT.

5 #luid by(pass around the long baffle. %f possible, use an exchanger type where

the long baffle is seal welded to the shell in order to avoid bypassing of the shell fluid.

This should be done with a full penetration weld. The exchanger types, where the

long baffles can be welded in, are #ixed Tube =heets or B(Tubes. %f B(tubes,

the number of tube passes must be a multiple of four. Then the bundles can be

removed. 1ther designs use multi(leaf long baffles for two shell passes. =ince

these cannot make a perfect seal, the amount of shell fluid bypassing the bundle

must be calculated.

Trouble+hooting Article 1ctober, 6CC8

To find out more about heat exchangers, see ale !ulley&s article in the 6CC8 =eptember issue of Hydrocarbo

rocessing. The title is [Troubleshooting =hell(and(Tube Heat ExchangersS. %t gives helpful information on

diagnosing problems.

>ndersurfaced +T Kuote

0hen a vendors heat exchanger -uote is under(surfaced, the following should be asked

6. =pecify the low(fin tubing be bare where it passes through baffling.

5. =pecify a tight tube hole tolerance.

. urchase tubing that has a fin 1.. the same as the bare ends.



6. )re there seal stripsG %f so, how manyG

5. 0hat tube hole clearance was used in the baffles

When to Add +hells in +eries )ugust, 5+++

Bsually you should design for the least number of shells for an item. However, there are times when it is mo

economical to add a shell in series to the minimum configuration. This will be when there is a relatively low

the shell side and the shell stream has the lowest heat transfer coefficient. This happens when the baffle spac

close to the minimum. The minimum for TEM) is >=hell %..3*?. Then adding a shell in series gives a highe

velocity and a much better heat transfer because of the smaller flow area in the smaller re-uired exchangers.

When to Consider a ong Baffle in the +hell 1ctober, 5++6

The cost curve for a shell and tube heat exchanger decreases with increasing surface. The curve flattens at ab

8,+++ s-uare feet of bare surface. %f the first selection has multiple shells that are not countercurrent flow and

each shell has less than 8,+++ s-uare feet, consider using a long baffle for cost savings. This is especially tru

the exchanger is of a type where the long baffle can be welded to the shell >less likely to bypass fluid?.

Which +trea* Goes ,nside Tubes for GasGas Exchangers?

%n a counter(current flow heat exchanger, the steam with the highest factor as calculated below goes inside th

Xou can also use the following factor if both gases molecular weight and temperature are about the same on

Why Did the )erfor*ance Decline in a TE.A - G or H Type +hell?

Has performance declined after the bundle has been pulled and later installed back in the shellG %f the longitu

long baffle is sealed on the sides with leaf seals, they are probably the problem. These thin flexible strips sho positioned so that they form a concave pattern and flex upward. Then, when the shell fluid puts pressure on t

leaves, they will press harder against the sides of the shell. %f there is too much pressure ( or if the bundle is

upside down ( the leaves will flex downward, and the shell fluid will bypass the bundle. )nother possibility

the leaf seals were damaged when the bundle was out of the shell.

-ouling factors for &ater6hrft=-Btu9

#or cooling water when velocity is (9 ft3sec

-ouling -actors for iuid Hydrocarbons6hrft=-Btu9

#actor >flow?5 3 density

#actor >flow?5 3 pressure

+.+++* steam,steam condensate,engine 2acket water

+.++6+ boiler feed water

+.++6* clean water,moutain water,etc.

+.++5+ normal cooling tower water

#ouling +.+5*[email protected]

0here @ ft3sec

+.++6+ %f sp. gravity )t 8+# less than +.9+, lube oil and heating oils



iscous -lo& >se .ore )ressure Drop Than >sual

High viscosity fluids can have a problem achieving the design heat transfer. The fluids are usually petroleum

and have an )% of 5+ or less.

:ow pressure drops can cause maldistribution of the tubeside flow which in turn reduces the heat transfer.

That is why you can see allowable pressure drops 5 or times higher than usual. There is a method by ).$. M

for calculating this minimum allowable pressure drop. )nother thing that can help is to use more tube passes

tubes than normal. )lso the fluid could be placed in the shell side if cleanig isn&t a problem.

+.++5+ %f sp. gravity )t 8+# +.9+ (+.9D

+.+++ %f sp. gravity )t 8+# +.9D (6.++

+.++*+ Heavy fuel oils



n

lems.

on

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ient

,



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are

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igh

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)rt Montemayor =eptember +, 5++*

age 6+9 of 6*8#ile4ame 59*+6C+68.xls

0ork=heet TEM) esignations

TEM) E=%!4)T%14=

-ront End +tationary Head +hell Type #ear End +tationary Head

A Channel and removable cover E One-pass shell Fixed tubesheet; like "A"

Stationary head.

B Bonnet !nteral Cover# - $-pass shell %ith lonitudinal . #ixed tubesheet_ like "'"

ba&&le stationary head.

C Channel interal %ith tubesheet G Split Flo% Shell $ Fixed tubesheet; like "C"

' removable cover. stationary head.

Sho%n( )emovable *ube

Bundle

$ Channel interal %ith tubesheet H +ouble split &lo% ) Outside, packed &loatin head

' removable cover.




age 6+C of 6*8#ile4ame 59*+6C+68.xls


D Special, hih-pressure closure J +ivided shell &lo% + Floatin head %ith backin

device split-rin#

Conventional Front nd eads(

Aor,

B

@ /ettle type o& reboiler T 0ull-throuh &loatin head

Other popular rear end head types employed(

> 1-tube bundle desin

2o )ear ead )e3uired#

W 0acked &loatin tubesheet %ith

lantern rin




age 66+ of 6*8#ile4ame 59*+6C+68.xls





age 666 of 6*8#ile4ame 59*+6C+68.xls


=ome examples of the TEM) designation for Heat Exchangers are shown below

BE. #ront bonnet >%ntergral $over?, with one(ass =hell and a #ixed Tubesheet rear 'onnet

#ixed tubesheet heat exchanger. This is a very popular version as the heads can be removed to clean the insid

of the tubes. The front head piping must be unbolted to allow front head removal_ if this is undesirable, then

this can be avoided by applying a type ) front head. %n that case only the cover needs to be removed. %t is no

possible to mechanically clean the outside surface of the tubes as these are fixed inside the shell. $hemical

cleaning can be used in the shell side. =hown is a version with one shell pass and two tube passes. This is

probably the least expensive of the shell(and(tube designs.

BE. This is the same type of heat exchanger as shown above, except it has only one tube pass

AE. $hannel with <emovable $over, 1ne ass =hell, #ixed Tubesheet 'onnet

This is almost the same type of heat exchanger as the first 'EM. The removable cover allows the inside of th

tubes to be inspected and cleaned without unbolting the piping. However, as can be expected, the tradeoff is

that this convenient feature makes it more expensive.






The maintenance feature of having a removable tube bundles re-uires an exchanger as the following

AE+ $hannel and <emovable $over, 1ne ass =hell, #loating Head with 'acking evice

) floating head heat exchanger is excellent for applications where the difference in temperature between the

hot and cold fluid causes unacceptable stresses in the axial direction, between the shell and tubes. The

floating head can move, i.e. it provides the ability to allow tube expansion in the axial direction.

4ote that the bundle can not be pulled from the front end. #or maintenance both the front and rear end head,

should be selected.

However, it is wise and prudent to be aware of the inherent trade(offs in this design. 4ote that the tube(side

fluid can leak through the internal floating head cover gasket and mix >or contaminate? the shell(side fluid.

%t is very difficult (and sometimes impossible to mitigate or compensate for the internal bolts tightening the

internal bonnet to remain under constant, steady tor-ue. Hot fluid temperatures make the bolts expand and

the result is a reduction in bolt tor-ue and subse-uent leaks through the bonnet gasket. )dditionally, it is a

common and expected occurance for maintenance crews to find the internal bolts badly rusted or corroded to

the point where they have to be burned or sawed off in order to extract the "removable" tube bundle.

The chemical engineer has other options to apply when re-uiring mechanical expansion of a heat exchanger

tube bundle. @arious rear head design also exist that allow for tube bundle expansion. )mong these are the

popular >and inexpensive? "B" tube bundle design. ) "" and "0" rear head design will also contribute this

feature without the haard of internal mixing >or contamination? of the two fluids.

)lso, be aware that any TEM) shell and tube design with a removable tube bundle feature has ( by nature ( a

larger shell diameter > increased cost? due to the need to be able to pull the rear tubesheet the length

of the exchanger&s shell. ) larger diameter shell can sometimes also present problems in a lower <eynolds

number >yielding a lower heat transfer? and internal by(passing of the shell fluid around the baffles >this also

reduces the effective heat transferred. )ll these effects eventually lead to a bigger heat exchanger >more area

and more tubes? in order to do a heat transfer operation.

including the backing device, must be disassembled. %f pulling from the front head is re-uired a type AET






:ongitudinal 'affles ( their application and inherent problems

The employment of longitudinal baffles in heat exchangers ( such as the "#", "!", and "H" shell types ( can

often resolve both heat transfer and fluid flow problems within the shell and tube exchanger used.

Their application can significantly increase the shell(side <eynolds 4umber and lead to more efficient shell(si

heat transfer coefficients with a subse-uent increase in heat transfer. )dditionally, these type of baffles permithe engineer to incorporate counter(flow heat transfer. True counter(current heat transfer is as efficient

a heat transfer configuration as an engineer can obtain. %n some heat recovery applications, this is highly sou

'y splitting the shell(side flow, some applications can actually have a significant reduction in shell(side press

drop. This is especially true in partial vacuum process operations where a minimum of pressure drop can be

tolerated.

However, the application of longitudinal baffles should be always carefully scrutinied and used sparingly. T

are, as would be expected, some very important trade(offs involved in the application of longitudinal baffles.

#irstly, if a longitudinal baffle is a process necessity, the baffle should be seal(welded against the inner shell

wall in order to ensure that there will be no internal, by(pass leakage. This positive step negates the possibiliof having a removable tube bundle. )dditionally, the welding necessity re-uires a minimum shell diameter

and this winds up being applicable only to relatively large streams.

'y the basic need to establish effective shell(side flow around a longitudinal baffle, one has to accept the

obvious fact that a minimum of shell(side clearances can be tolerated. 1nce having said and applied these fa

one then has to also accept that the re-uired, small baffle clearances mean extraordinary fabrication techni-ue

and resultant super(human maintenance efforts to extract a removable tube bundle. %n far too many actual

field cases, it has been found that the removable tube bundle with a longitudinal baffle is a non(practical devi

#ield results have shown that in most cases the tube bundle has resulted in being destroyed in order to remove

This extraordinary and desperate maintenance act labels such a design as non(practical.




age 66* of 6*8#ile4ame 59*+6C+68.xls




)rt Montemayor

age 668 of 6*8

Heat Exchanger Tube Sheet Layout Count TableSource( "Applied 0rocess +esin &or Chemical and 0etrochemical 0lants"; 4ol. 5; p.$6

rnest . 7ud%i; 8ul& 0ublishin Co.; ouston, *9 :<=#

Shell !. +., inches

Tube O. D. & Pitch 8 10 12 1!1"# 1$!1"# 1%!1"# 1!1"# 21!1"# 2!1"# 2$ 2% 2

O n e - 0 a s s

F i x e d

* u b e s 5>6" on :=>:<" *rian. 55 < :?= :5= :5 $6@ 5?@ 5: 6: ==5 <<5 @<5

5>6" on :" *rian. 55 =@ : ::@ :=@ $:@ $@@ 565 6$5 65 =@@ <<@

5>6" on :" S3uare 55 =5 = :?: :5 :5 $5= $@ 5== 6: 6= =@

:" on :-:>6" *rian. := 55 =@ @5 :?5 :55 :<5 $?= $6@ 5?@ 5<: 6$@

:" on :-:>6" S3uare :@ 55 6= <= 5 ::: :5 :@ $:= $== 5?5 5=

* % o - 0 a s s

F i x e d

* u b e s 5>6" on :=>:<" *rian. 5$ = 6 :$6 :<< $$ 5?? 5@? 6=$ =$ <$< @56

5>6" on :" *rian. $ =< ? ::? :=6 $? $<6 5$< 5 6< ==< <6<

5>6" on :" S3uare $< 6 @ 6 :$< :@$ $$$ $? 56< 6? 6< =<?

:" on :-:>6" *rian. :< 5$ =$ <$ $ :$< :<$ $?6 $66 $$ 56< 6:?

:" on :-:>6" S3uare :$ $< 6? =< @< :?< :5< :@$ $: $6 $ 56

1

* u b e s

5>6" on :=>:<" *rian. 56 <6 6 :56 :? $56 5?6 5 6<? == <6

5>6" on :" *rian. $< <? @$ :? := $:$ $@? 55< 6?< 66 =<<

5>6" on :" S3uare :$ 5? =$ @$ :?? :6$ : $6$ 5?6 5<$ 65< =?<

:" on :-:>6" *rian. 99 $< 6$ = 6 :$? :=6 :$ $56 $6 56?

:" on :-:>6" S3uare 99 :$ $$ 5 = @< :?? :56 :? $:6 $=< 5?6

F o u r - 0 a s s

F i x e d

* u b e s 5>6" on :=>:<" *rian. 99 6 6 :? :=6 :< $<< 55$ 6:$ 66 =@< <?

5>6" on :" *rian. 99 66 @$ < :56 :? $5$ $$ 5<? 6$6 =? =<

5>6" on :" S3uare 99 6 @$ :$< :6$ :$ $6$ 5? 5<< 66? =:?

:" on :-:>6" *rian. 99 $6 66 <? @ :?6 :5 :@< $:$ $= 5? 5<

:" on :-:>6" S3uare 99 $6 6? 6 @6 6 ::? :6$ : $:6 $<? 5:?

1

* u b e s

5>6" on :=>:<" *rian. 99 $ =< 6 :$$ :<< $: $< 5@ 65 =56 <$$

5>6" on :" *rian. 99 $? =$ <6 :6< : $=6 5: 5< 6<$ =6$

5>6" on :" S3uare 99 $6 66 <6 ? :5? :@6 $$< $< 56$ 6:6 6$

:" on :-:>6" *rian. 99 $? 5< =? @6 ::? :6$ :@ $: $<< 5$$

:" on :-:>6" S3uare 99 :< 5$ =? << ? :$$ :<< : $5 $<

S i x - 0 a s s

F i x e d

* u b e s 5>6" on :=>:<" *rian. 99 ? ::< :@6 $5? $6 5@$ 66? =5$ <5$

5>6" on :" *rian. 99 << :?6 :=< $?$ $= 5$$ 5 6<6 =6

5>6" on :" S3uare 99 =6 @ ::< := $:$ $<< 5$6 56 6<?

:" on :-:>6" *rian. 99 56 =< $ ::$ :=? :$ $$< $@6 55

:" on :-:>6" S3uare 99 66 << ::< :=6 :6 $$< $<

1

* u b e s

5>6" on :=>:<" *rian. 99 @6 ::? :=< $?< $@$ 5= 6:< =:? =<

5>6" on :" *rian. 99 =< :56 :6 $< 5?? 5<< 66? =:

5>6" on :" S3uare 99 =< ? :: :<? $:? $< 5$$ 5$ 6=:" on :-:>6" *rian. 99 5? 6$ < :?? :5? :< $?< $=$ 5?6

:" on :-:>6" S3uare 99 6$ <? ? ::? :=$ :$ $$6 $<

i " h t - 0 a s s

F i x e d

* u b e s 5>6" on :=>:<" *rian. 99 6 :6? : $= 55$ 5 66 =@<

5>6" on :" *rian. 99 $ :$6 :@? $$6 $< 566 6$$ 6<

5>6" on :" S3uare 99 6 :5$ :@6 $$ $< 5=$ 6:6

:" on :-:>6" *rian. 99 << ? :$? :=6 :? $6? $

:" on :-:>6" S3uare 99 @6 6 :$ :=? :$ $5?

1

* u b e s

5>6" on :=>:<" *rian. 99 < :?$ :6$ :? $=6 56$ 5 6? =@

5>6" on :" *rian. 99 =$ $ :$$ :@? $$< $< 5=? 6$$ 6

5>6" on :" S3uare 99 6 @? :?< :6< :6 $=6 5?< 5@6 65

:" on :-:>6" *rian. 99 $6 5 = ? :: :=6 :? $5 $?

:" on :-:>6" S3uare 99 56 =? @? :6$ :@? $?< $=6

2otes( :# *he above tube counts have an allo%ance made &or *ie )ods.$# *he )adius o& Bend &or the 1-*ube bundles is e3ual to $.=# *ube O.+.#; *he actual number o& 1-tubes is :>$ o& the above &iu



)rt Montemayor March 65, 5++6

<ev +

age 66D of 6*8#ile4ame 59*+6C+68.xls

0ork=heet HP esign

HE'T E(CH')*E+ S,--'+

/

8

1/

128

xchaner eat +uty $0 Btu>hr Overall 1, estimated 100

2umber o& shell passes 1

2umber o& tube passes 2

7o ean *emperature +i&&erence, 7*+ %2

F Factor see belo%# 0.8

Adusted 7*+ %0

eat *rans&er Area calculated $/2

+esin continency &actor 1.2$

Over-desin allo%ance 1.00

eat *rans&er Area re3uired %02

6=? psi, Saturated Steam )e3Dd, $1/ lbs>hr

CE )e3Dd :6 de rise, pm $/# pm

Calculation o actor3

0 or S# ?.:6

) $.5

*erm : ?.<

0x ?.:6

*erm $ :.<?

*erm 5 ?.5 :.6<*erm 6A :5.6=

*erm 6B @.$<

*erm 6 ?.<$

F 0.8

* in, Cold Side t:# oF

* out, Cold Side t$# oF

* in, ot Side *:# oF

* out, ot Side *$# oF

Btu>hr - Ft$ - oF

oF

oF

Ft$

Ft$

G)0-:#>0-:#H :>2#

)I$J:#?.=>)-:#

3, 1, A, ∆*m

%, cp, t

:

E, Cp, *

$

E, Cp, *

:

%, cp, t

$



SHELL & T,4E HE'T E(CH')*E+ SPEC55C'T5O) Sheet 1 o& 1

nlish 1nits#

Corporation 0roect 2o.

: Service Lean -E' Solution Cooler 3uipment 2o.

$ 7ocation 1nit 0.O. 2o.

5 anu&acturer 6 odel 6 &r )e&. 2o. 6 2o. )e3Dd

6 *A SiKe, *ype oriK. 4ert. Connected in Series 0arallel

= Sur&ace>1nit 6 8ross &&. Shells>1nit One Sur&ace>Shell 6 8ross &&.

< 0'!+ 2o. 0lot 0lan 2o. Other )e&. +% 2o.

@ PE+O+-')CE O O)E ,)5T

Fluid Allocation SHELL S5DE T,4E S5DE

Fluid Circulated:? *otal Fluid nterin lb>h

:: 4apor !n>Out# lb>h

:$ 7i3uid lb>h

:5 Steam lb>h

:6 2on-Condensables lb>h

:= Fluid 4aporiKed or Condensed lb>h

:< Steam Condensed lb>h

:@ *emperature LF

: +ensity, Speci&ic 8ravity

: 4iscosity c0

$? 4apor olecular Eeiht

$: Speci&ic eat Btu>lbMLF

$$ *hermal Conductivity Btu>hM&tMLF

$5 7atent eat Btu>lb

$6 Operatin 0ressure, !nlet psi

$= 4elocity ax. in. &ps

$< 0ressure +rop, Clean Allo%.>Calc.# psi

$@ Foulin )esistance

$ eat xchaned Btu>h 7o *+ 1ncorrected# LF 7o *+ Corrected# 6 LF

$ *rans&er )ate, Service 6 *rans&er )ate, Clean 6

5? CO)ST+,CT5O) ')D -'TE+5'LS

5: S77 S!+ *1B S!+ Sketch Bundle, 2oKKle Orientation#

5$ +esin 0ressure psi

55 *est 0ressure psi

56 +esin *emperature LF

5= 2umber o& 0asses per Shell

5< !n

5@ Out

5 !ntermediate

5 *ubes( *ype 2umber 6 O+ 0.%$ in. 1/ BE8 or in. ( in. Av. Eall

6? *ube 7enth in. *ube 0itch 0.%$ in. Flo% 0attern circle one#

6: Shell( !+ 6 in. O+ 6 in. *ube-to-*ubesheet Noint +olle7 an7 Seal el7e7

6$ Ba&&les - Cross( *ype 6 Spacin 6 in. 6 Cut on ( +iam. Area

65 Ba&&les - 7on( 0erm. )emovable Seal *ype( Bypass Seal(

66 !nlet 2oKKle 6 lb>&tMsec Bundle ntrance 6 lb>&tMsec Bundle xit 6 lb>&tMsec

6= xpansion NointP Qes ( 2o *ype( !mpinement 0rotectionP ( Qes 2o

6< P'+T TH9 in. C.'. in. P'+T TH9 in. C.'. in.

6@ *ubes Stainle:: Stl 1/ 4* ;in. Floatin *ubesheet Carbon Steel 6 !!!!

6 Shell Fixed *ubesheet Carbon Steel 6 0.12$

6 Shell Cover *ube Supports Carbon Steel 6 0.12$

=? Channel Cross Ba&&les Carbon Steel 6 0.12$=: Channel Cover 7on Ba&&le Carbon Steel 6 0.12$

=$ Flt ead Cover 8askets Stainle:: Stl !!!!

=5 1ser Spec.(

=6 Code )e3uirements( 'S-E Sec. <555 Para. 1 =12> StampP e: *A Class(

== Eeihts( Shell 6 lb Filled %ith Eater 6 lb Bundle 6 lb

=< +e;ar?:

=@

=

)ev +ate +escription By Chk. Appr. )ev +ate +escription By Chk. Appr.

? For 0urchase

&t$ &t$

&t$MhMLF>Btu.

Btu>&t$MhMLF. Btu>&t$MhMLF

.

ρ@23

-'TE+5'LA -'TE+5'LA

A Stress )elieved ark "S)D# and>or )adioraphed ark D9)D# 0arts

:. !tems marked %ith an asterisk 6# to be completed by 4endor.

ConnectionsSiKe ')atin

<?R ?R6=R5?R

+ e @ .

) o .

-onte;ayor



PL'TE & +'-E HE'T E(CH')*E+ SPEC55C'T5O) Sheet 1 o& 1

nlish 1nits#

Corporation 0roect 2o. 12#$/%

: Service Cooling ater Exchanger 3uipment 2o.

$ 7ocation 1nit 0.O. 2o.

5 anu&acturer 6 odel 6 &r )e&. 2o. 6 2o. )e3Dd One

6 SiKe, *ype 6 ! 6 Frames>1nit One Connected in Single

= Sur&ace>1nit 6 Eecti@e Sur&ace>Frame 6 *ro::

< 0'!+ 2o. 0lot 0lan 2o. Other )e&. +% 2o.

@ PE+O+-')CE O O)E ,)5T

Fluid Allocation O* S!+ CO7+ S!+

Fluid Circulated Cooling ater

:? *otal Fluid nterin lb>h 1$00 20/#8

:: 4apor !n>Out# lb>h !!!! !!!! !!!! !!!!

:$ 7i3uid lb>h 1$00 1$00 20/#8 20/#8

:5 Steam lb>h !!!! !!!! !!!! !!!!

:6 2on-Condensables lb>h !!!! !!!! !!!! !!!!

:= Fluid 4aporiKed or Condensed lb>h !!!! !!!! !!!! !!!!

:< Steam Condensed lb>h !!!! !!!! !!!! !!!!

:@ *emperature LF 2$ 120 0 10$

: +ensity, Speci&ic 8ravity 0.0% 0.2 0.$ 0.2

: 4iscosity c0 0.$# 1.% 0.%/ 0./$

$? 4apor olecular Eeiht !!!! !!!! !!!! !!!!

$: Speci&ic eat Btu>lbMLF 0.8/% 0.8# 1.0 1.0$$ *hermal Conductivity Btu>hM&tMLF 0.1%8 0.1/0 0.$8 0./$

$5 7atent eat Btu>lb !!!! !!!!

$6 Operatin 0ressure, !nlet psi %$ /0

$= 4elocity ( ax. in. &ps 8.0 8.0

$< 0ressure +rop, Clean Allo%.>Calc.# psi 10 6 10 6

$@ Foulin )esistance 0.001 0.00

$ eat xchaned 0%28 Btu>h 7o *+ 1ncorrected# 1$%.0 LF 7o *+ Corrected# 6 LF

$ *rans&er )ate, Service 6 *rans&er )ate, Clean 6

5? CO)ST+,CT5O) ')D -'TE+5'LS

5: Allocation HOT S5DE COLD S5DE Sketch Frame, 2oKKle Orientation#

5$ +esin 0ressure psi 1$0 12$

55 *est 0ressure psi Co7e Co7e

56 +esin *emperature LF 00 005= 2umber o& 0asses per Frame TBo 6

5< Corrosion Allo%ance in. 0.0/2$ )one

5@ !n 1$0 + / 12$

5 Out 1$0 + / 12$

5 !ntermediate !!!! !!!!

6? lb>&tMs

6: !mpinement 0rotectionP e:

6$ 2o. o& 0lates Frame Capacity ax. 2o. o& 0lates#

65 P'+T TH9 in. C.'. in. P'+T TH9 in. C.'. in.

66 0lates Stnle:: Steel1/ 4* ;in. 0.012$ Connections Stnle:: Steel 0.012$

6= 0late 8askets Carbon Steel 6 0.012$ Frame Carbon Steel 0.012$

6< nd Cover Carbon Steel 6 0.012$ Carryin Bar Carbon Steel 0.012$

6@ Carbon Steel 0.012$ Carbon Steel 0.012$

6

6 OSA *ype 0rotective ShroudP e: aterial( Carbon Steel !nsulation( Heat Con:er@ation

=? Cleanin( 0aintin(

=: Code )e3uirements( 'S-E Sec. <555 Para. 1 =12> StampP e:

=$ Client Spec.( Eeihts( mpty Frame 6 lb Filled %ith Eater 6 lb

=5 +e;ar?:

=6

==

)ev +ate +escription By Chk. Appr. )ev +ate +escription By Chk. Appr.

? -+ec-< For !n3uiry ABC +F 9Q

&t$ &t$

&t$MhMLF>Btu

Btu>&t$MhMLF Btu>&t$MhMLF

ρv$, !nlet>Outlet

-'TE+5'LA -'TE+5'LA

A Stress )elieved ark "S)D# and>or )adioraphed ark D9)D# 0arts

:. !tems marked %ith an asterisk 6# to be completed by 4endor.

ConnectionsSiKe ')atin

+ e @ .

) o .

-onte;ayor



)rt Montemayor 1verall Heat Transfer $oefficient1ctober +5, 5++

<ev +

age 65+ of 6*8#ile4ame 59*+6C+68.xls

0ork=heet Typical "B"

Tyical O@erall Heat Tran:er Coeicient:

=ource http33www.the(engineering(page.com3forms3he3typB.html

+hell and Tube Heat Exchangers "(erall L>M

Hot -luid Cold -luid

Heat Exchangers

0ater 0ater 9++ \ 6,*++ 67+ ( 587

1rganic solvents 1rganic =olvents 6++ ( ++ 6D \ *5

:ight oils :ight oils 6++ ( 7++ 6D \ D+

Heavy oils Heavy oils *+ ( ++ C \ *

<educed crude #lashed crude * ( 6*+ 8 \ 58

<egenerated E) #ouled E) 7*+ ( 8*+ DC \ 667

!ases >p atm? !ases >p atm? * ( * 6.+ \ 8

!ases >p 5++ bar? !ases >p 5++ bar? 6++ ( ++ 6D \ *

Coolers

1rganic solvents 0ater 5*+ ( D*+ 77 \ 65:ight oils 0ater *+ ( D++ 85 ( 65

Heavy oils 0ater 8+ ( ++ 66 ( *

<educed crude 0ater D* ( 5++ 6 \ *

!ases >p atm? 0ater * ( * 6.+ \ 8

!ases >p 5++ bar? 0ater 6*+ ( 7++ 58 \ D+

!ases 0ater 5+ ( ++ 7 \ *

1rganic solvents 'rine 6*+ ( *++ 58 \ 99

0ater 'rine 8++ \ 6,5++ 6+8 \ 566

!ases 'rine 6* ( 5*+ ( 77

Heaters

=team 0ater 6,*++ \ 7,+++ 587 ( D++

=team 1rganic solvents *++ \ 6,+++ 99 ( 6D8

=team :ight oils ++ ( C++ * \ 6*C

=team Heavy oils 8+ ( 7*+ 66 \ DC

=team !ases + ( ++ * \ *

Heat Transfer >hot? 1il Heavy oils *+ ( ++ C \ *

Heat Transfer >hot? 1il !ases 5+ ( 5++ 7 ( *

#lue gases =team + ( 6++ * ( 69#lue gases Hydrocarbon vapors + (6++ * ( 69

Condensers

)-ueous vapors 0ater 6,+++ \ 6,*++ 6D8 \ 587

1rganic vapors 0ater D++ \ 6,+++ 65 \ 6D8

<efinery hydrocarbons 0ater 7++ ( **+ D+ ( CD

W*=C Btuhrft=o-



)rt Montemayor 1verall Heat Transfer $oefficient1ctober +5, 5++

<ev +


0ork=heet Typical "B"

@apors with some non condensables 0ater *++ ( D++ 99 \ 65

@acuum condensers 0ater 5++ ( *++ * \ 99

Vaporizers

=team )-ueouos solutions 6,+++ \ 6,*++ 6D8 \ 587

=team :ight organics C++ \ 6,5++ 6*C \ 566

=team Heavy organics 8++ ( C++ 6+8 \ 6*C

Heat Transfer >hot? oil <efinery hydrocarbons 5*+ ( **+ 77 \ CD

Air Cooled Exchangers

)rocess -luid 6tube side9

0ater ++ ( 7*+ * ( DC

:ight organics ++ ( D++ * ( 65

Heavy organics *+ ( 6*+ C ( 58

!ases *+ ( ++ C ( *

$ondensing hydrocarbons ++ ( 8++ * ( 6+8

,**ersed coils

Coil -luid )ool -luid

Natural circulation

=team *++ \ 6,+++ 99 \ 6D8

=team :ight oils 5++ ( ++ * \ *

=team Heavy oils D+ ( 6*+ 65 \ 58

)-ueous solutions 0ater 5++ ( *++ * \ 99

:ight oils 0ater 6++ ( 6*+ 69 \ 58

Agitated =team 9++ \ 6,*++ 67+ \ 587

=team :ight oils ++ ( *++ * \ 99

=team Heavy oils 5++ ( 7++ * \ D+

)-ueous solutions 0ater 7++ ( D++ D+ ( 65

:ight oils 0ater 5++ ( ++ * ( *

Jacketed (essels

Jacket -luid essel -luid

=team *++ ( D++ 99 ( 65

=team :ight organics 5*+ ( *++ 77 ( 99

0ater 5++ ( *++ * ( 99

0ater :ight organics 5++ ( ++ * ( *

)rts 4ote )bove Bs were originally given in metric units and the conversion to good,

old fashioned B= engineering units is based on

ilute a-ueous

ilute a-ueous

ilute a-ueous

solutions

ilute a-ueous

1.0 4tu"hr!t2!o F $./%82/ att:";2!o9



)rt Montemayor * 9ETTLES +EC5+C,L'T5O) COOLE+E!*!#

1ctober 57, 6CCD

age 655 of 6*8 Electronic #ile4ame 59*+6C+68.xls (

!+ T Shell !nternal diameter, in. T $5

*ube external diameter, in. T ?.@=

*ubesD 0itch, in. T ?.5@=

CD T Clearance bet%een tubes T ?.:@=

B T Ba&&le spacin, in. T :=

2 T 2umber o& Shell-side ba&&les T ::

0.#%2

E T Shell-side mass &lo%rate,lb>h T 5$==??

Shell-side unit mass &lo%rate,lb>h-&t$ T /%0#

3uivalent shell diameter, &t T ?.?6=555

Shell-side )eynolds 2umber T /#.

& T 0.00$2

?.=$

$?

FluidDs 4iscosity, lb>&t-h #8.#

k T ?.?<

FluidDs 0randtl 2umber T 22.%

:s T Shell &luidDs speci&ic ravity T ?.5?5@$:

Shell-side pressure drop, psi T 1/.8

O+*

0*

aS Shell-side cross&lo% area, &t$

8S

+e

2)e

Friction &actor &or pressure drop, &t$3in$

c0 FluidDs eat Capacity, Btu>lb-o#

µ FluidDs 4iscosity, c0 T

µD

FluidDs therm. cond., Btu>&t-h-o#

20r

Φs 4iscosity ratio, µ3µ%#?.:6 T

∆

From "0rocess eat *rans&er"; +. /ern; c8ra%-ill; :=?; paes :6@-:6

P F *2:D

:=)G1>"=$.22 x 1010>D

e:

:



uic? & Dirty Tubular Heat Exchanger +ating Sheet

0roect +eactor ar; ater Sy:te; ,gra7e 0roect 2o.

!tem 2o. E!*!(( Service By +ate>*ime :5-ar-@ :=(:

!nput &lo%s, conditions and properties data &or shellside and

tubeside. trans&er coe&&icients to this point i.e., not includin shellside h#(

*ube Side Shell $5=

C Fluid 2ame ar; ater :=

6:,??? Flo% #, lb>h :=,??? *hen the re3uired trans&er A T $,:5 :5? 2umber o& tubes re3uired T =6=

:?$ :?? )eset tubes>pass Step 5#, then no. o& passes T 6

Av. +ensity <$.?= <:. *otal tube count T =6

Av. 4iscosity ?.@$5 ?.=? 8.1

Av. eat Capacity : : Actual e&&ective trans&er area, A T $,$5

eat xchaned =,=?,?:= U, Btu>h =,=?,??? O9I

Av. *hermal Conductivity ?.5<? ?.5<

Foulin )esistance ?.??$ ?.??:= *ube 0itch ?.5@= in.

0randtl 2o. 6.< 5. and estimate shell diameter 0attern *ri

1ncorrected *+ :. Shell !+ &rom *ube Count *ables $@ in.

Corrected *+ :6.? Select Ba&&le Spacin :< in.

2umber o& Ba&&les T :6

?.<??

*ube O+ ?.@=?? in. ?.== in.

lenth can be trial and error#. BE8 :< 5$=,???

*ube !+, d T ?.<$? in. $=,$=

*ube 7enth, 7 T $? &t. Shellside Friction Factor T ?.??:@

?.5?$ 2.%

&&ective trans&er area per tube T 5.$@ ?.6

:,:6?

:=.:

*ubes>pass T :6< ?.? O9

tubes per tube pass. lb>h per tube T $,<5 <:6.

Av. velocity, &ps T <.:: O9I

6?,5$6

*ubeside Friction Factor, & T ?.?:?

:.?: O9I adust tube lenth, number o& tubes per pass, number o& passes, and>o::5.@ ba&&le spacin. )emember to reset shell diameter &rom tube count tab

:,55= re3uired.

Ste 1. Ste #. =tart configuring the exchanger. 'egin with the total calcul

start Btu

On that basis, assumed 1o Btu

&t$

*emp. in, o#

*emp. out, o#ρ, lb3ft5

µ, c *ubeside ∆ >incl. returns? psic

p, u Fo

&t$

k, Btu>hM&tMo#

), &t$FhFo#3'tu Ste $. =elect tube arrangementc

pµ

o#o#

Flo% Area across Bundle, as &t$

Ste 2. %nput tubing 1, '0! and 3uva en ame er, e see a e ass 4elocity, 8

s lb>h

Shellside )eynolds 2o., 2)e

Flo% area per tube, at in.$ Shellside ∆ psi

&t$ Outside *rans&er Factor, h

Outside Film Coe&&icient, ho

a cu a eo

Ste . Estimate the number of Check( di&&erence, 1calc.

vsassum.

clean

*ubeside )eynolds 2o., 2)e

Ste /. $heck tubeside velocity and ∆, shellside ∆. %f too high or

∆ per pass, psi ns e rans er ac or,h

!nside Film Coe&&icient, hi



C'LC,L'T5O) SHEET

Sinature Art ontemayor +ate 18!eb!0# Checked +ate 0ro 2o.

0roect D. . 9ern Proce:: Heat Tran:erJ -c!*raB HillJ 1$0J . 2%# File

Subect HoriKontal n!Proanol Total Con7en:er Sheet 1 o& 10

:

$ A horiKontal, :-$ condenser is re3uired &or condensin pure propyl alcohol emanatin &rom the top o& a distillation

5 column. Side-to-side, $= cut semental ba&&les %ill be used. Basic data is as &ollo%s(

6

= 0ropanol &lo%rate /0000 lb>hr

< 0ropanol vaporsD inlet pressure 1$.0 psi@ 0ropanol vaporsD inlet temperature 2##

Coolin Eater inlet temperature 8$

0ropanol allo%able pressure drop 2.00 psi

:? CES allo%able pressure drop 10.00 psi

:: +irt &actor 0.00

:$ Condenser tubesD lenth 8.00 &eet

:5 *ubesD O+ 0.%$00 inches

:6 *ubesD lenth 8.00 &eet

:= *ubesD aue 1/ BE8

:< *ubesD !+ 0./200 inches

:@ *ubesD pitch 0.%$ *rianular, inches

: Clearance bet%een tubes 0.18%$ inches

: 0ropanol 7atent eat at := psi 28$ Btu>lb

$? 0ropanol olecular Eeiht /0.1

$:

$$

$5

$6

$=

$<

$@

$

$

5?

5:

5$

55

56

5=

5<

5@

5

5

6?

6:

6$

65

66

' 4 C D E * H 5 9 L

o#o#

* e m p e r a t u r e , o

F

=

$66

+istance alon tubes

4aporinlet

Condensateoutlet

Coolin %aterout

Coolin %ater in



C'LC,L'T5O) SHEET



Subect HoriKontal n!Proanol Total Con7en:er Sheet 2 o& 10

6=

6< First, make a heat and material balance to establish the heat load and the coolin %ater re3uired(

6@

6 0ropanol latent heat &or condensation T 1%100000 Btu>hr

6 Coolin %ater terminal temperature T 120

=? Coolin %ater re3uired T #88$%1 lb>hr T %/ pm=: 0ropanol Eater +i&&er.

=$ $66 iher *emperature :$? :$6

=5 $66 7o%er *emperature = :=

=6 ? +i&&erence 5= 5=

==

=< 7o ean *emperature +i&&erence T 7*+ T 1#1

=@

= Since the shell side 0ropanol vapor is essentially isothermal, the exchaner is in true counter&lo%.

=

<? *he caloric temperature o& the hot &luid

<: *he caloric temperature o& the cold &luid

<$ *he averae temperature o& the cold &luid

<5 *he in&luence o& the tube-%all temperature is included in the condensin &ilm coe&&icient.

<6

<= 102.$

<<

<@ xecute a trial calculation(

< a# 100

< Condensin &ilm coe&&icients %ill enerally rane &rom :=? to 5??. Assumin a &ilm coe&&icient o& :,???

@?

@:@$ eat trans&er area T A T 121$

@5 Uuantity o& 5>6" O+ tubes T %%

@6

@= b#

@< lare number o& tubes, makin a $-pass assumption inadvisable.

@@ From the tube counts table, 6 tube passes usin 5>6" O+ tubes on :=>:<" trianular pitch , yields a

@ %// tubes in a 1 inch !+ shell.

@ c#

? Corrected area, A T 120

: 101

$

5

6

=

<

@

' 4 C D E * H 5 9 L

o#

o#

*c

tc

ta

*he mean ta T o# can be used as the caloric temperaure of the cold fluid

Assume that 1+ T Btu>hr-o#(ft$

&or %ater, 1C %ill rane &rom :5? to $5? Btu>hr-oF-&t$.

U>1+ ∆T &t$

Assume that # tube passes are used. *he 3uantity o& %ater is lare, but the condenser %ill have a

count o&

*he corrected 1+ coe&&icient, usin the 5:" shell, is no% calculated(

&t$

Corrected 1+ U>A ∆T Btu>hr-o#(ft$



C'LC,L'T5O) SHEET



Subect HoriKontal n!Proanol Total Con7en:er Sheet o& 10

? Calculation: or :hell :i7e hot lui7 n!Proanol

:

$

5 5 crosses &or the proposed side-to-side &lo%. Since these are the minimum ba&&les that can be used, this should

6 yield the lo%est attainable shell-side pressure drop in this con&iuration.=

< 1. 3. @.:; p.:5#

@ %here,

!+ T Shell inside diameter, inches

CD T Clearance bet%een tubes, inches

:?? B T Ba&&le spacin, inches

:?: *ube pitch, inches

:?$

:?5 ##$

:?6

:?= *he condensate loadin on the horiKontal tubes T 8DD T 8./ lb>hr-linear &t

:?< %here,

:?@ 7 T *ube lenth, &eet

:? *ube 3uantity e&&ective &or condensation

:?

::? 200

:::

::$ 10%$ )e&er to line V :<=#

::5 %here,

::6

::= *he inside %ater# &ilm heat trans&er coe&&icient T 100 From &i. $= #::<

::@ 12$ 3. =.5:; p. #

:: %here,

::

:$?

:$: 18#

:$$

:$5 0.0$ From *able 6#

:$6

:$= 0.80 From *able <#

:$<

:$@ 0./2 c0 From Fi. :6#

:$

:$

:5?

:5:

:5$

:55

' 4 C D E * H 5 9 L

Assume a maximum ba&&le spacin. *his %ill be 5$-:>$", 5:", and 5$-:>$" %hich is e3ual to <" or 2 baffles and

*he shell-side or bundle cross&lo% area T aS >%? >$&? >'?3>

* 677? &t$

0*

*he shell-side mass velocity T 8s 0 3 a

S lb>hr-&t$

E>7W2t$>5 T

2t

Assume the value o& the averae condensin &ilm coe&&icient T hO Btu>hr-&t$-oF

hiO h

i >%31? Btu>hr-&t$-oF

hiO *he inside &ilm %ater# heat trans&er coe&&icient re&ered to the tube O+, Btu>hr-&t $(o#

hi Btu>hr-&t$

(o

#

*ube %all temperature T tE t

a J Gh

O>h

iO J h

O#H *

v - t

a# T oF

*v Averae temperature o& hot &luid vapor#, oF

Shell side &ilm temperature T t& *

v N t

%?35 o#

Shell side &ilm thermal conductivity T k& Btu>hr-&t$(o#3ft

Speci&ic 8ravity o& shell side &ilm T s&

4iscosity o& shell side &ilm T µ&



C'LC,L'T5O) SHEET



Subect HoriKontal n!Proanol Total Con7en:er Sheet # o& 10

:56

:5= 3uation :$.6$; p. $<<#

:5< %here,

:5@ hD T

:5 &ilm coe&&icient absolute viscosity T 1.$00# lb>&t-hr

:5:6? T #.2

:6: T T #.18EG08

:6$ 8DD T Condensate loadin &or horiKontal tubes, lb>hr-&t

:65

:66 Averae shell side condensin &ilm coe&&icient T 1%8

:6=

:6< Calculation: or tube :i7e col7 lui7 ater

:6@

:6 Flo% area o& a 5>6" O+ x :< BE8 tube T 0.020 From condenser tube table#

:6

:=? Flo% area per tube T 0.#02

:=: %here,

:=$ 2umber o& tubes e&&ective &or condensation

:=5

:=6 n T 2umber o& tube passes

:==

:=< 121/$08

:=@

:= Averae %ater velocity in the tube side T 4 T $.#1 &t>sec

:=

:<? :?$.=:<:

:<$ 0.%2 c0 T 1.%# lb>&t-hr

:<5 *ubesD !+ T 0.0$1% &t

:<6

:<= )eynolds 2umber &or pressure drop only# T /0%

:<<

:<@ 100 From Fi. $=#

:<

:< 10%$

:@? %here,

:@:

:@$

:@5 Based on hD T :@$ instead o& the assumed $??, a ne% value o& t% and t& could be obtained to ive a more exact value

:@6 o& hD based on the &luid properties at a value o& t& more nearly correct. o%ever, it is not necessary in this example

:@= because the condensate properties %ill not chane materially.

:@<

:@@

:@

' 4 C D E * H 5 9 L

hD µ& $ 3 k

& 5 ρ

& $ g?:>5 6.* >7 !&&3µ

& ?-:>5

Averae condensin &ilm coe&&icient, Btu>hr-&t$(o#

µ&

k& &ilm coe&&icient thermal conductivity, Btu>hr-&t$

(o

#3ftρ

& &ilm coe&&icient density, lb>&t5

Acceleration o& ravity, &t>hr $

Btu>hr-&t$(o#

in$

2* a&

t 3 677 n &t$

2*

aDt Flo% area per tube, in$

Eater mass velocity in the tube side T 8t % > a

t T

lb>hr-&t$

8t > 5,<??Wρ# T

At the averae %ater temperature, ta, ofo

F(

Eater viscosity T µ T

+ 8t3µ

*ube side %ater heat trans&er &ilm coe&&icient T hi Btu>hr-&t$(o#

hiO h





C'LC,L'T5O) SHEET



Subect HoriKontal n!Proanol Total Con7en:er Sheet $ o& 10

:@

:? Calculation: or :hell :i7e re::ure 7ro

::

:$ *he propanol vapor temperature T 2##

:5 0ropanol vapor viscosity T 0.010 c0 T 0.02#2 lb>&t-hr From &i. :=#

:6:=

:< *he hydraulic radius employed &or correlatin shell-side coe&&icients &or bundles havin ba&&les is not the true hydraulic

:@ radius. *he direction o& &lo% in the shell is partly alon an d partly at riht anles to the lon axes o& the bundleDs tubes.

: *he &lo% area at riht anles to the lon axes is variable &rom tube ro% to tube ro%. A hydraulic radius based upon

: the &lo% area across any one ro% could not distinuish bet%een s3uare and trianular pitch. !n order to obtain a simple

:? correlation combinin both the siKe and closeness o& the tubes and their type o& pitch, excellent areement is

:: obtained i& the hydraulic radius is calculated alon instead o& across# the lon axes o& the tubes.

:$

:5 6 W &ree area#>%etted area# T

:6 T 0.$$ inches T 0.0#$8 &t From &i. $#

:=

:< Shell-side )eynolds 2umber T 8$1

:@

: Shell-side &riction &actor &or $= cut semental ba&&les T & T 0.001#1 From &i. $#

:

$?? 2umber o& shell-side crosses T 2J:# T

$?:

$?$ Assume that the propanol vapor &ollo%s the ideal as la% at the lo% pressure.

$?5

$?6 0ropanol vapor density T 0.2/

$?=$?< 0ropanol vapor speci&ic ravity T s T 0.00%8

$?@

$? 2.$8 &t

$?

$:? Shell-side pressure drop T 1.2 psi 3. :$.6@; p.$@5#

$::

$:$

$:5 Calculation: or tube :i7e re::ure 7ro

$:6

$:= For the tube side )eynolds 2umber T /0% the correspondin tube-side &riction &actor

$:<

$:@ & T 0.0001 From &i. $<#

$:

$: *ube-side pressure drop T Straiht tube pressure drop J )eturn 7oss pressure drop

$$?

$$: . ps& T 0.02 psi

$$$ 3. @.6=; p. :6#

$$5

' 4 C D E * H 5 9 L

o#

Shell!:i7e eMui@alent 7ia;eter =De>3

+e G6# ?.= W0

* +.98

* ( +.* π d$ 37+^ 3 >+.* π d?

+e !

s 3µ

&t$>in$

E > 4:# *

$>*

:# 0

:>0

$# T lb>&t5

Shell !nside +iameter T +s

:>$# G & W8s$

s >4N6? 3>*.55 6+:?

e s?^

&t$3in$

Straiht tube pressure drop T ∆t & W 8

t$ :n3>*.556+:?

e s Φ

t?



C'LC,L'T5O) SHEET



Subect HoriKontal n!Proanol Total Con7en:er Sheet / o& 10

$$6

$$= %. psi 3. @.6<; p.:6#

$$<

$$@ *otal tube-side pressure drop T %.2 psi

$$ %here,

$$ 7 T tube lenth, &eet$5? n T 2umber o& tube passes

$5:

$5$ D T

$55

$56

$5=

$5<

$5@ 1$2.#

$5

$5

$6?

$6:

$6$ 101 From line :#

$65

$66 0.00

$6= 2ote( !n condensation calculations the omission o& the tube metal resistance may introduce a sini&icant error and

$6< should be checked.#

$6@

$6

$6 1%8 h outside# 10%$

$=? 1$2.#$=: 101

$=$ 0.00

$=5 0.00

$=6 1.2 %.2

$== 2.00 10.00

$=<

$=@ Conclu:ion3

$= *he &irst trial calculated is satis&actory and yields the &ollo%in exchaner(

$= Shell :i7e Tube :i7e

$<? !+ T 1 inche: Uuantity and lenth T %//J 8N ! 0

$<: Ba&&le spacin T 1 inche: =arox. O+, BE8, ' pitch T "#J 1/ 4*J 1$"1/ triangular

$<$ 0asses T 1 0asses T #

$<5

$<6 !t is interestin at this point to compare a vertical condenser %ith this horiKontal model. *he horiKontal and vertical

$<=

$<< number o& tubes in both models is the same. *o this end a vertical condenser %ill be assumed %hich uses the same

$<@ tube count as the above except that the tube lenth may be :$ or :< &t as needed# to account &or the lo%er

$< coe&&icients obtained in the vertical orientation.

' 4 C D E * H 5 9 L

)eturn 7oss pressure drop T ∆r 6Wn>s# 4$35 g&?

Φt *he viscosity ratio µ3µ

%?+.67 in the tubes

Acceleration o& ravity, 5$.$ &t>sec$

Calculation o clean o@erall coeicient ,C

1C >h

io h

o ?3>h

io N h

o? Btu>hr-&t$-oF

a cu a on o r ac or 73

Corrected 1+ Btu>hr-o#(ft$

)d T 1

C - 1

+#>1

C W 1

+# T hr-&t$-oF>Btu

Shellside

Su;;ary o+e:ult:

*ubeside

1C T1

+ T

)d calculated T

)d re3uired T

Calculated ∆

Allo%able ∆

condensin &ilm coe&&icients are both a&&ected by E and 2t, and the best basis fof comparison is otained when the



C'LC,L'T5O) SHEET



Subect <ertical n!Proanol Total Con7en:er Sheet % o& 10

$<

$@? *he vertical condenser to be rated %ill be oriented as seen in the sketch belo%. *he process conditions %ill be

$@: identical to those o& the previous horiKontal model rated. !n order to prevent %ater corrosion in the carbon steel shell,

$@$ the %ater %ill also be introduced in the tube side.

$@5

$@6$@=

$@<

$@@

$@

$@

$?

$:

$$

$5

$6

$=

$<

$@

$

$

$?

$:

$$

$5 *otal heat trans&erred T 1%100000 Btu>hr

$6

$= 7o ean *emperature +i&&erence T 7*+ T 1#1$<

$@

$ 102.$

$

5?? Trial Calculation3

5?:

5?$ a# %0

5?5 *he e3uation &or the condensin &ilm coe&&icient ives reater values &or horiKontal tubes than &or vertical tubes.

5?6

5?=

5?< eat trans&er area T A T 1%$

5?@

5? *he nearest common, available tube lenth usin the same @<< tubes# is(

5?

VVV *ube lenth T 11.$ &eet

5::

5:$

5:5

' 4 C D E * H 5 9 L

o

#

Caloric temperature o& the 0ropanol vapor T *C

Caloric temperature o& the %ater T tC T o#

Assume that the overall dirty heat trans&er coe&&icient, 1+ T Btu>hr-&t$(o#

!t %ill, conse3uently, be necessary to reduce the value o& 1+.

U>1+ ∆T &t$

use 12 &oot lenth tubes #

4aporinlet

Condensateoutlet

Coolin %ater

outCoolin %ater in



C'LC,L'T5O) SHEET



Subect <ertical n!Proanol Total Con7en:er Sheet 8 o& 10

5:6

5:= b# *he same tube layout, usin 5>6" O+ x :< BE8 tubes on :=>:<" trianular pitch and 6 passes %ill also be

5:< used.

5:@

5: c#

5: 180#5$? /%

5$:

5$$

5$5 Calculation: or :hell :i7e hot lui7 n!Proanol

5$6

5$= *ubesD outside diameter, +o T ?.?<$= &t

5$<

5$@ lb>hr-lin. &t 3. :$.5<; p. $<=#

5$

5$ 100

55?

55: 10%$ )e&er to line V :<=#

55$ %here,

555

556 *he inside %ater# &ilm heat trans&er coe&&icient T 100 From &i. $= #

55=

55< 11#.$ 3. =.5:; p. #

55@ %here,

55

55

56? 1%56:

56$ 0.0$ From *able 6#

565

566 0.80 From *able <#

56=

56< 0./$ c0

56@

56 3uation :$.5; p. $<<#

56 %here,

5=? hD T

5=: &ilm coe&&icient absolute viscosity T 1.$% lb>&t-hr

5=$

5=5 T #.2

5=6 T T #.18EG08

5== 8D T Condensate loadin &or vertical tubes, lb>hr-&t

5=<

5=@ Averae shell side condensin &ilm coe&&icient T 10#

5=

' 4 C D E * H 5 9 L

*he corrected 1+ coe&&icient, usin the 5:" shell, is no% calculated(

Corrected area, A T U>1+∆t &t$

Corrected 1+ U>A ∆T Btu>hr-o#(ft$

Condensate loadin &or vertical tubes T E>2t π

o

Assume the value o& the averae condensin &ilm coe&&icient T hO Btu>hr-&t$-oF

hiO h



hi Btu>hr-&t$(o#

*ube %all temperature T tE t

a J Gh

O>h

iO J h

O#H *

v - t

a# T oF

*v Averae temperature o& hot &luid vapor#, oF

Shell side &ilm temperature T t& *v N t%?35 o

#

Shell side &ilm thermal conductivity T k& Btu>hr-&t$(o#3ft

Speci&ic 8ravity o& shell side &ilm T s&

4iscosity o& shell side &ilm T µ& From Fi. :6; also, 6W8D>µ T :,?$=#

hD µ& $ 3 k

& 5 ρ

& $ g?:>5 6.7D >7 !&3µ

& ?-:>5

Averae condensin &ilm coe&&icient, Btu>hr-&t$(o#

µ&

k& &ilm coe&&icient thermal conductivity, Btu>hr-&t$(o#3ft

ρ& &ilm coe&&icient density, lb>&t5

Acceleration o& ravity, &t>hr $

Btu>hr-&t$(o#



C'LC,L'T5O) SHEET



Subect <ertical n!Proanol Total Con7en:er Sheet o& 10

5=

5<? Calculation: or tube :i7e col7 lui7 ater

5<:

5<$ *he tube-side %ater conditions and con&iuration is the same as the horiKontal con&iuration.

5<5

5<6 10%$5<= %here,

5<<

5<@

5<

5< Calculation: or :hell :i7e re::ure 7ro

5@?

5@: !t is necessary to arrane the :$-&oot tube bundle into a minimum number o& bundle crosses, or 2 J :# T =.

5@$ *he spacin bet%een ba&&les %ill be(

5@5 B T 2 inches

5@6

5@= 1.2# 3. @.:; p.:5#

5@< %here,

5@@ !+ T Shell inside diameter, in.

5@ CD T Clearance bet%een tubes, in.

5@ B T Ba&&le spacin, in.

5? *ube pitch, in.

5:

5$ #88%

55

56 *he propanol vapor temperature T 2##

5= 0ropanol vapor viscosity T 0.010 c0 T 0.02#2 lb>&t-hr From &i. :=#5<

5@ 0.0#$8 &t From table in &i. $#

5

5 1/#2

5?

5: Shell-side &riction &actor &or $= cut semental ba&&les T & T 0.001#0 From &i. $#

5$

55 2umber o& shell-side crosses T 2J:# T $

56

5= 0ropanol vapor speci&ic ravity T s T 0.00%8 Same as line $?<#

5<

5@ 2.$8 &t

5

5 Shell-side pressure drop T 2. psi 3. :$.6@; p.$@5#

6??

6?: *his pressure drop prediction is hih, and i& it cannot be compensated &or by elevatin the condenser, it %ill be

6?$ necessary to use the hal&-circle =? cut# support ba&&les as sho%n in xample @-.

6?5

' 4 C D E * H 5 9 L

hiO hi >%31? Btu>hr-&t$

-o

F


Shell-side bundle# cross&lo% area T as !+ W CD W B#>0

* 677? &t$

0*

*he shell-side mass velocity T 8s 0 3 a

S lb>hr-&t$

o#

3uivalent diameter &or pressure drop T +e

Shell-side )eynolds 2umber T )eS

e !

S 3 µ

&t$>in$

Shell !nside +iameter T +s

:>$# G & W8s$

s >4N6? 3>*.55 6+:?

e s?^



C'LC,L'T5O) SHEET



Subect <ertical n!Proanol Total Con7en:er Sheet 10 o& 10

6?6

6?= Calculation: or tube :i7e re::ure 7ro

6?<

6?@ *he basic data is the same as in the horiKontal model example, except &or the tube lenth.

6?

6? $.0 ps& T 0.0 psi6:? 3. @.6=; p. :6#

6::

6:$ %. psi 3. @.6<; p.:6#

6:5

6:6 *otal tube-side pressure drop T %.0 psi

6:=

6:<

6:@

6:

6: $.0

6$?

6$:

6$$

6$5

6$6 /% From line :#

6$=

6$< 0.00#

6$@

6$

6$

65? 10# h outside# 10%$65: $.0

65$ /%

655 0.00#

656 0.00

65= 2. %.0

65< 2.00 10.00

65@

65 Conclu:ion3

65 Shell :i7e Tube :i7e

66? !+ T 1 inche: Uuantity and lenth T %//J 12N ! 0

66: Ba&&le spacin T 2 inche: =arox. O+, BE8, ' pitch T "#J 1/ 4*J 1$"1/ triangular

66$ 0asses T 1 0asses T #

665

666 *his vertical condenser is some%hat secure in per&ormin the speci&ied heat trans&er duty but it exceeds the

66= allo%able pressure drop, althouh not seriously. *he advantae o& horiKontal condensation may be observed

66<

66@ *he vertical unit has an inherent advantae, ho%ever, %hen the condensate is to be subcooled.

66

' 4 C D E * H 5 9 L

Straiht tube pressure drop T ∆t & W 8t

$

:n *.55 6+:?

e s Φt

)eturn 7oss pressure drop T ∆r 6Wn>s# 4$35 g&?

Calculation o clean o@erall coeicient ,C

1C >h

io h

o ?3>h

io N h

o? Btu>hr-&t$-oF

a cu a on o r ac or73

Corrected 1+ Btu>hr-o#(ft$

)d T 1

C - 1

+#>1

C W 1

+# T hr-&t$-oF>Btu

Shellside

Su;;ary o+e:ult:

*ubeside

1C T

1+ T

)d calculated T

)d re3uired T

Calculated ∆

Allo%able ∆

&rom the 1C of 679.* in the horiontal condenser as compared with the C.5 in the vertical unit in identical servi



Some o& this data %as taken &rom Standards o& the *ubular xchaner anu&acturers Associatio

4*

1"2 O. D. Con7en:er tube

Outside !nside%

8 ?.:<=

?.:6

10 ?.:56 ?.6$

11 ?.:$? ?.=:?

12 ?.:? ?.$$ ?.?<$= ?.:5? ?.?@5 ?.6=< ?.=5$

1 ?.?= ?.=<?

1# ?.?5 ?.556 ?.?@< ?.:5? ?.?@6 ?.5@? 0.$8#

1$ ?.?@$ ?.<?<

1/ ?.?<= ?.5@? ?.:?@= ?.:5? ?.?< ?.5?$ :< ?.<$?

1% ?.?= ?.<56

18 ?.?6 ?.6?$ ?.:$< ?.:5? ?.:?=$ ?.$5< : 0./$220 ?.?5= ?.65? ?.:6=$ ?.:5? ?.::$< ?.:@6 $$@ ?.<?

22 ?.?$ ?.666 ?.:=6 ?.:5? ?.::<$ ?.:6: $6:

)OTES3

-aterial actor

Aluminum ?.5=

*itanium ?.=

A.!.S.!. 5?? Series Stainless Steels ?.

A.!.S.!. 6?? Series Stainless Steels :.?$ Aluminum BronKe :.?6

Aluminum Brass :.?<

2ickel-Chrome-!ron :.?@

Admiralty :.?

2ickel :.:5

2ickel-Copper :.:$

Copper and Cupro-2ickels :.:6

Eallthickness

inches

*ube !. +.

inches

*ube &lo%

area in$

Sur&ace area perlinear &oot, &t$

u e %e per linear

&oot, lb o&W

Constant

CWW

*ube !. +.

inches

W *he %eiht o& the condenser tubes is based on lo% carbon steel %ith a density o& ?.$5< lbs>in

WW 7i3uid 4elocity %ithin the tubes T 7bs 0er *ube our# > C W 7i3uid Speci&ic 8ravity# in &eet p



n *A#; @th dition :#; pae :@. 2ote( some o& the tabular *A data contained ))A*A, but this %a

"# O. D. Con7en:er tube 1 O. D. Con7en:er tu

Outside !nside Outside !nside

?.<@? ?.5=$< ?.$<: ?.:@=6

?.@?6 ?.55 ?.$<: ?.:65

?.:$= 0.1/ ?.:$<$ ?.5 $= ?.@5$ ?.6$? ?.$<: ?.::<

?.$?65 ?.:<5 ?.:55= ?.? 5: ?.@<? ?.6=5< ?.$<: ?.:?

?.$$$5 ?.:<5 ?.:55 ?.@6@ 56@ ?.@$ ?.6?5 ?.$<: ?.$?6@

?.$6<5 ?.:<5 ?.:6<< ?.<<= 56 ?.:? ?.=:=5 ?.$<: ?.$:$:

0.2/% 0.1/ 0.1$2 0.$2 #18 0.8# 0.$#/ 0.2/18 0.218

?.$6 ?.:<5 ?.:=@ ?.=$$ 6=? ?.=< ?.=@== ?.$<: ?.$$6:

?.5?: ?.:<5 ?.:<$5 ?.6@< 6@: ?.@? ?.=6= ?.$<: ?.$$@

?.5:=@ ?.:<5 ?.:<<? ?.6$ 6$ ?.6 ?.<:5 ?.$<: ?.$5:6

0. 0.1/ 0.1%0% 0./% $21 ?.?$ ?.<5? ?.$<: ?.$5<:?.5<5$ ?.:<5 ?.:@? ?.$< =<@ ?.5? ?.<@5 ?.$<: ?.$65=

*ube &lo%

area in$

Sur&ace area per linear&oot, &t$

*ube %eihtper linear &oot,

lb o& steel

Constant

CWW

*ube !. +.

inches

*ube &lo%

area in$


. #or other metal materials multiply by the following factors

er sec. Speci&ic ravity o& Eater <? o# 6.++?



s corrected %ith this spreadsheetDs &ormulas.

be 1!1"# O. D. Con7en:er tube

Outside !nside?.? ?.<$$: ?.5$@$ ?.$55? $.?= @?

:.6@5 ==? ?.$? ?.<<6 ?.5$@$ ?.$6? :.:6 :,?5@ :.:@?

:.56 ?.=6 ?.@:6 ?.5$@$ ?.$6 :.@66 :.$??

:.$6: <=< ?.$ ?.@=@6 ?.5$@$ ?.$=@: :.= :,:$ :.$5?

:.:$ @? :.?:? ?.?:$ ?.5$@$ ?.$<66 :.6=? :,$=? :.$<?

:.?5 @6 :.?5? ?.55$ ?.5$@$ ?.$<@ :.56: :,5?= :.$?

?.: ?6 :.?<? ?.$= ?.5$@$ ?.$@@= :.:@5 :,5@@ :.5:?

0.81# 8$2 :.?? ?.:<: ?.5$@$ ?.$$@ :.?= :,66? :.55?

?.@:6 :.::? ?.<@@ ?.5$@$ ?.$?< ?.5 :.5<?

?.<=? $@ :.:$? ?.=$ ?.5$@$ ?.$5$ ?.$6 :,=5@ :.5@?

?.=6 :.:5? :.??$ ?.5$@$ ?.$= ?.@<5 :.5?

?.6 @ :.:=? :.?5@ ?.5$@$ ?.5?:: ?.<6: :,<$< :.6???.5<: :,?<? :.:? :.?5< ?.5$@$ ?.5? ?.6== :,@?<

u e %e per linear

&oot, lb o&

Constant

CWW

*ube !. +.

inches

*ube &lo%

area in$


u e %e per linear

&oot, lb o&

Constant

CWW

*ube !. +.

inches



1!1"2 O. D. Con7en:er tube 2 O. D. Con7en:er tub

Outside !nside Outside !nside

:.?@=: ?.5$@ ?.5?<5 $.5==

:.:5:? ?.5$@ ?.5:6$ $.:<=

:.:$ ?.5$@ ?.5$$? :.@? :,<?

:.$6< ?.5$@ ?.5$ :.@@: :.@<? $.65$ ?.=$5< ?.6<?

:.$< ?.5$@ ?.55=: :.<5= $,?:6 :.@$ $.66: ?.=$5< ?.6<<=

:.56@ ?.5$@ ?.565? :.6$@ :.:? $.=@5? ?.=$5< ?.6@5

:.55 ?.5$@ ?.56$ :.$< $,:? :.56 $.<6:@ ?.=$5< ?.6?:

:.6=$@ ?.5$@ ?.5=<? :.?@?

:.6@6: ?.5$@ ?.5=@ ?.@ $,5??

:.6=@ ?.5$@ ?.5<:5 ?.$6

:.=56 ?.5$@ ?.5<<= ?.@@=

*ube &lo%

area in$


*ube %eihtper linear &oot,

lb o& steel

Constant

CWW

*ube !. +.

inches

*ube &lo%

area in5

Sur&ace area perlinear &oot, &t5



e

$.6:$ 5,@=

$.$?6 5,:

:.5= 6,?:6

:.@?: 6,:$:

u e%eiht per

linear &oot,

Constant

CWW



)rt Montemayor Heat Exchanger Tubesheets

Tubesheet Thickness1ctober +C, 6CC6

<ev +

age 6C of 6*8#ile4ame 59*+6C+68.xls

0ork=heet Tube=heet

*he thickness o& heat exchaner tubesheets is an important consideration in cost-estimatin and selectin

desin alternatives &or process heat systems. Accordin to the *ubular xchaner anu&actureres Assn.

*A# standards, the tubesheet thickness &or shell-and-tube exchaners is iven by the &ormula(

F T 1.2$

8 T 12 inches

0 T $0 psi

S T 1%$00 psi

T F 1.0/ inches

*A ives precise rules &or determinin the variables F, 8, 0, and S &or exchaner desin. For estimatin

purposes, ho%ever, these terms can be taken as(

* T *ubesheet thickness, inches

F T a &actor

T :.? &or stationary and &loatin-head tubesheets

T :.$= &or 1-tube tubesheets

8 T shell internal diameter, as calculated &rom trans&er sur&ace and tube dimensions, inches

0 T desin pressure, psiS T tubesheetsD material allo%able stress, psi

4alues o& S &or some common materials are sho%n in the &ollo%in table. Eith this table and the other terms,

tubesheet thickness can be calculated in this spreadsheet.

aterial:?? $?? 5?? 6?? =??

SA-=:< 8rade @? 1%$00 1%$00 1%$00 1%$00 1%$00

Stainless Steel -- :@,@?? :<,:?? :=,?? --

:.$=Cr - ?.=o - Si Steel :=,??? :=,??? :=,??? :=,??? :=,???

onel :@,=?? :<,=?? :=,=?? :6,?? :6,@??

SB-:@: 2aval Brass -- :$,=?? :?,=?? $,??? --SB-6?$ Copper 2ickel :$,=?? 10$00 10#00 10#00 10#00

SB-:: Copper <,<?? =,@?? =,??? -- --

From( Che;ical Engineering -agaKine_ lant 4otebook_ May 65, 6CD*

*emperature, o#

T = F G

2

√ P

S



SHELL ')D T,4E HE'T E(CH')*E+ T,4ESHEET L'O,TS =T,4E CO,)TS>

Source( "0rocess eat *rans&er"; +onald U. /ern, c8ra%-ill Book Co. :=?#; pae 6:

5>6" O. +. tubes on :-inch s3uare pitch :" O. +. tubes on :-:>6 inch s3uare pitch :-:>6" O. +. tubes o

8 5$ $< $? $? $: :< :6

10 =$ =$ 6? 5< 5$ 5$ $< $6 :< :$

12 : @< < < <? 6 6= 6? 5 5< 5? $61!1"# @ ? $ @< @? <: =< =$ 6 66 5$ 5?

1$!1"# :5@ :$6 ::< :? :? : @< < < <6 66 6?

1%!1"# :@@ :<< := :=? :6$ ::$ ::$ < ? $ =< =5

1!1"# $$6 $$? $?6 :$ : :5 :5$ :$ :$$ ::< @ @5

21!1"# $@@ $@? $6< $6? $56 :@@ :<< := :=$ :6 < ?

2!1"# 56: 5$6 5? 5?$ $$ $:5 $? :$ :6 :6 :$@ ::$

2$ 6:5 56 5@? 5=< 56< $<? $=$ $5 $$< $$$ :6? :5=

2% 6: 6<? 65$ 6$? 6? 5?? $ $@ $< $<? :<< :<?

2 ==5 =$< 6? 6< 6=< 56: 5$< 5?? $6 $< :5 :

1 <=@ <6? <?? =? =<? 6?< 5 5? 5< 5= $$< $$?

@6 @: < <@< <6 6<= 6<? 65$ 6$? 6:6 $= $=$

$ 6= $6 @? @<< @6 =$$ =: 6 66 6@$ $5 $@

% 56 :6 < << 5 =< =@6 =<$ =66 =5$ 556 5$$

:?6 :?$6 $ < 6 <<= <66 <$6 <:$ <?? 5@? 5<$

)ote3 *hese tube counts can be taken only as an estimate. For accurate tube counts, an actual scaled layout should be done.

/ern does not reveal %here he obtained this in&ormation and he is not speci&ic in ivin details to %hat *A type, orientation, and Outer *ube 7

Conse3uently, the user is advised to scrutiniKe this in&ormation be&ore usin it.

Another estimatin method &or tube counts is &ound in "0etroleum )e&inery nineerin"; 2elson; c8ra%-ill; 0ae =66(

*he number o& heat exchaner tubes can be estimated &rom the e3uation

%here,

C T ?.@= a constant &or S3uare pitch#

0 T the tube spacin, in inches

hell !. +.nches :

*ube0ass$

*ube0ass6

*ube0ass<

*ube0ass

*ube0ass:

*ube0ass$

*ube0ass6

*ube0ass<

*ube0ass

*ube0ass:

*ube0ass$

*ube0ass *

) F C 6 =L"P>2



7 T the Outer *ube 7imit, in inches

*he O*7 is about :-:>$" less than the inside diameter o& the shell in &loatin head exchaners.

!t is about =>" less than the shell inside diameter o& &ixed-head or 1-tube construction.

*ube Spacin T 1.$ inches

Outer *ube 7imit T 1.$ inches

2umber o& *ubes T /1



h s3uare pitch :-:>$" O. +. tubes on :-@> inch s3uare pitch

:< :< :< :< :$ :$$$ $$ $$ $$ :< :<

5= 5: $ $ $= $6 $$

6 66 5 5 56 5$ $

<6 =< =? 6 6= 65 5

$ @ <$ <? =@ =6 =?

:?$ < @ @6 @? << <$

:$5 ::= 6 ? < 6 @

:6< :6? ::$ :? :?$ 6

:@6 :<< :5: :$@ :$? ::< ::$

$?$ :5 :=: :6< :6: :5 :5:

$5 $$< :@< :@? :<6 :<? :=:

$< $= $?$ :< : :$ :@<

5?6 $5 $$6 $$? $:@ $:? $?$

56$ 55< $=$ $6< $<@ $5? $$6

his data applies.

<ube0ass

*ube0ass

:*ube0ass

$*ube0ass

6*ube0ass

<*ube0ass

*ube0ass



SHELL ')D T,4E HE'T E(CH')*E+ T,4ESHEE

Source( "0rocess eat *rans&er"; +onald U. /ern, c8ra

5>6" O. +. tubes on :=>:<-inch trianular pitch 5>6" O. +. tubes on :-inch

8 5< 5$ $< $6 : 5@ 5? $6

10 <$ =< 6@ 6$ 5< <: =$ 6?

12 :? < $ @ $ $ @<

1!1"# :$@ ::6 < ? < :? :?< <

1$!1"# :@? :<? :6? :5< :$ :=: :5 :$$

1%!1"# $5 $$6 :6 : :@ $?5 :< :@

1!1"# 5?: $$ $=$ $66 $56 $<$ $=? $$<

21!1"# 5<: 56$ 5:6 5?< $? 5:< 5?$ $@

2!1"# 66$ 6$? 5< 5@ 5<6 56 5@< 5=$

2$ =5$ =?< 6< 66< 656 6@? 6=$ 6$$

2% <5@ <?$ ==? =5< =$6 == =56 6

2 @$: <$ <6? <$? =6 <5? <?6 ==<

1 6@ $$ @<< @$$ @$? @6= @$ <@

@6 5 @ =$ $< =< 5? @@6

$ ::?$ :?< :??6 = @? 5 $

% :$6? :$?? ::66 ::?6 :?@$ :?@6 :?66 :?:$

:5@@ :55? :$= :$6 :$:$ :$?< ::@< ::$

)ote3 *hese tube counts can be taken only as an estimate. For accurate tube counts, an actual scal

/ern does not reveal %here he obtained this in&ormation and he is not speci&ic in ivin details

As an example o& a discrepancy, re&er to the " shell %ith 5>6" tubes on :=>:<" trianular pitch

Conse3uently, the user is advised to scrutiniKe this in&ormation be&ore usin it.

*rianular pitch should never be used %ith a dirty or &oulin &luid on the shellside o& an exchan

Another estimatin method &or tube counts is &ound in "0etroleum )e&inery nineerin"; 2els

*he number o& heat exchaner tubes can be estimated &rom the e3uation

%here,

C T ?.< a constant &or *rianular pitch#

0 T the tube spacin, in inches

7 T the Outer *ube 7imit, in inches

*he O*7 is about :-:>$" less than the inside diameter o& the shell in &loatin head exchaners.!t is about =>" less than the shell inside diameter o& &ixed-head or 1-tube construction.

*ube Spacin T 1.$ inches

Outer *ube 7imit T 1.$ inches

2umber o& *ubes T %0

Shell !. +.!nches :

*ube0ass$

*ube0ass6

*ube0ass<

*ube0ass

*ube0ass:

*ube0ass$

*ube0ass6

*ube0ass

) F C 6 =L"P>2



L'O,TS =T,4E CO,)TS>

%-ill Book Co. :=?#; pae 6$

trianular pitch :" O. +. tubes on :-:>6 inch trianular pitch :-:>6" O. +. tubes

$6 $: :< :< :6

5< 5$ 5$ $< $6 $? :

@6 @? == =$ 6 6< 6 5$ 5?

$ @6 < << = =6 =? 5 5<

:: ::? : < ? @6 @$ =6 =:

:@$ :<< :5: :: :?< :?6 6 < <<

$:< $:? :<5 :=$ :6? :5< :$ = :

$@$ $<? : : :@? :<6 :<? ::@ ::$

56$ 5$ $6: $5$ $:$ $:$ $?$ :6? :5<

56 5$ $6 $$ $=< $=$ $6$ :@? :<6

6@6 6<6 56 556 5?$ $< $< $?$ :<

=5 =? 5@ 5@< 55 556 5:< $5= $$

<<< <6? 6@$ 6=6 65? 6$6 6?? $@= $@?

@<? @5$ =5 =$$ 6< 6@? 6=6 5:= 5?=

<6 6 <? =$ =<$ =6< =5$ 5=@ 56

< @? <@6 <<6 <5$ <:6 = 6?@ 5?

::?? :?@ @<< @5< @?? < <@$ 66 65<

ed layout should be done.

to %hat *A type, orientation, and Outer *ube 7imits O*7# this data applies.

nd $-passes. An actual layout yields 6 tubes %ith 5>:<" O*7, as compared %ith the listed 5$ tubes.

er. *his con&iuration is impossible to clean mechanically.

n; c8ra%-ill; 0ae =66(

<*ube0ass

*ube0ass

:*ube0ass

$*ube0ass

6*ube0ass

<*ube0ass

*ube0ass

:*ube0ass

$*ube0ass



on :->:< inch trianular pitch :-:>$" O. +. tubes on :-@> inch trianular pitch

:6

$< $$ $? : :6 :6 :$ :$

5$ $ $< $@ $$ : :< :6

6= 6$ 5 5< 56 5$ 5? $@

<$ = =6 6 66 6$ 5 5<

< @ < <: = == =: 6

:?= :?: = @< @$ @? << <:

:5? :$5 ::@ = : < ? @<

:== :=? :6? ::= ::? :?= =

:= :@ :@? :5< :5: :$= :: ::=

$:@ $:$ $?$ :<? :=6 :6@ :6: :5<

$== $6= $5= :6 :@@ :@$ :<= :<?

$@ $ $@= $:= $?< $?? :? :6

55= 5$@ 5:= $6< $5 $5? $$? $:=

5? 5@6 5=@ $@= $< $<? $=$ $6<

6$= 6: 6?@ 5?@ $ $? $6 $@=

6*ube0ass

<*ube0ass

*ube0ass

:*ube0ass

$*ube0ass

6*ube0ass

<*ube0ass

*ube0ass



)rt Montemayor 4ovember +, 6CCD

age 67D of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes

reater than that sho%n &or a &ixed tube sheet desin and smaller &or a pull-throuh &loatin head. !n any case,

tubes that can be &itted into a iven tubesheet depends upon the pass partition pattern, the thickness o& the pas

dividers and exactly %here the drillin pattern is started relative to the dividers and the outer tube limit. Addition

tubes %ill be lost &rom the bundle &or a 1-tube desin because the minimum bendin radius prevents tubes &ro

bein inserted in some, or all, o& the possible drillin positions near the centerline o& the 1-tube pattern. *ubes

%ill also be lost i& an impinement plate is inserted underneath the noKKle. For a no-tubes-in-the-%indo% desi

)u;ber o Tube Pa::e:

1 2 # /

<.$

?.@= ?.5@= *rian. 5 5$ $< $6?.@= :.???? S3uare 5$ $< $? $?

?.@= :.???? *rian. 5@ 5? $6 $6

:.?? :.$=?? S3uare $: :< :< :6

:.?? :.$=?? *rian. $$ : :< :6

.@@

?.@= ?.5@= *rian. <$ =< 6@ 6$

?.@= :.???? S3uare =$ =$ 6? 5<

?.@= :.???? *rian. <: =$ 6 6

:.?? :.$=?? S3uare 5$ 5$ $< $6

:.?? :.$=?? *rian. 5@ 5$ $ $

12.00 :?.@=

?.@= ?.5@= *rian. :? < $

?.@= :.???? S3uare ? @$ < <

?.@= :.???? *rian. ? 6 @$ @?

:.?? :.$=?? S3uare 6 66 6? 5

:.?? :.$=?? *rian. =@ =$ 66 6$

1.2$ :$.??

?.@= ?.5@= *rian. :$@ ::6 < ?

?.@= :.???? S3uare = ? : @@

?.@= :.???? *rian. ::? :?: ?

:.?? :.$=?? S3uare <? =< =: 6<

:.?? :.$=?? *rian. <@ <5 =< =6

1$.2$ :6.??

?.@= ?.5@= *rian. :@? :<? :6? :5<

?.@= :.???? S3uare :5 :5$ ::< ::$

?.@= :.???? *rian. :<5 :=$ :5< :55

:.?? :.$=?? S3uare $ @= @?:.?? :.$=?? *rian. < $ < 6

1%.2$ :<.??

?.@= ?.5@= *rian. $5 $$6 :6 :

?.@= :.???? S3uare : :@ :< :<6

?.@= :.???? *rian. $:: $?: :: :@<

:.?? :.$=?? S3uare ::$ ::? :?$

:.?? :.$=?? *rian. :5? :$6 ::< ::?

1.2$ :.??

?.@= ?.5@= *rian. 5?: $$ $=$ $66

?.@= :.???? S3uare $5< $$6 $:< $?

?.@= :.???? *rian. $@5 $=< $6$ $5<

:.?? :.$=?? S3uare :6 :6$ :5< :$

TOT'L ),-4E+ O T,4ES 5) ') E(CH')*E+ )t3

!& not kno%n by direct count, &ind the tube 3uantity in the tube count table as a &unction o& +otl

, the tube

pitch, p, and the layout. *he shell diameter +i and outer tube limit +

otl iven in the table are those &or a

con@entional :lit!ring loating hea7 7e:ign, &ully tubed out. For a iven shell diameter, the value o& +otl

%ill

the tube count can be reasonably interpolated &rom the *able usin the kno%n or speci&ied +otl

, asumin that

the tube count is proportional to +otl

#$. 'll tube count table: are only aroxi;ate since the actual number o

the actual number o& tubes in the bundle is Fc 4

t. #

c is the fraction of total tubes in crossflow.

Shell !+in.

Outer *ube7imit

+iameter,in.

*ube O+in

*ube0itch, in.

*ube7ayout

8.0%1=Sch. 0>

10.02=Sch. #0>




age 679 of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes

:.?? :.$=?? *rian. :@$ :<$ :=$ :6

21.00 :.$=

?.@= ?.5@= *rian. 5<: 56$ 5:6 5?<

?.@= :.???? S3uare $@< $<6 $6< $6?

?.@= :.???? *rian. 5: 5? $@ $<

:.?? :.$=?? S3uare :@? :< :=@ :=?

:.?? :.$=?? *rian. : : :@? :<6

2.2$ $:.=?

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

2$.00 $5.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

2%.00 $=.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

2.00 $@.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

1.00 $.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

.00 5:.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

$.00 55.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

%.00 5=.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

.00 5@.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

#2.00 6?.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare




age 67C of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes

:.?? :.$=?? *rian.

##.00 6$.$=

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

#8.00 6<.??

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

$2.00 =?.??

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

$/.00 =6.??

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.

/0.00 =.??

?.@= ?.5@= *rian.

?.@= :.???? S3uare

?.@= :.???? *rian.

:.?? :.$=?? S3uare

:.?? :.$=?? *rian.




age 6*+ of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes

s

al

,

8

:

5<

<?

<

5<

6?

<

@?

@6

66

=?

:$

:?

::?

<6@$

:@

:6$

:<<

$

6

$56

:

$:?

::<

e

&




age 6*6 of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes

:$

$?

$56

$<?

:6

:<?




age 6*5 of 6*8 #ile4ame 59*+6C+68.xls0ork=heet Total Tubes



)rt Montemayor 4ovember +, 6CC

Tube OD in. Tube Pitch in. Layout

?.<$= ?.:$= ?.@?6 ?.6?<

?.@=? ?.5@= ?.:6 ?.6<

?.@=? :.???? :.??? :.???

?.@=? :.???? ?.@?@ ?.@?@

?.@=? :.???? ?.<< ?.=??

:.??? :.$=?? :.$=? :.$=?

:.??? :.$=?? ?.6 ?.6

:.??? :.$=?? :.?$ ?.<$=

Tube Pitch Tye:3

)ote3 loB arroB: are eren7icular to the bale cut e7ge

P in. P

n in.

T,4E P5TCH P'+'LLEL TO LO PP ')D )O+-'L TO LO P

)

*hese 3uantities are needed only &or the purpose o& estimatin other parameters. !& a detailed dra%in o&the exchaner is available, or i& the exchaner itsel& can be conveniently examined, it is better to obtain

these other parameters by direct count or calculation. *he 3uantities are described by Fiure =.$-: and read&rom *able !4 &or the most common tube layouts.

5?o *rianular <?o )otated *rianular

Flo%

Flo%

S3uare)otated S3uare



)rt Montemayor Heat Exchanger Temperatures)ugust 56, 5++7

<ev +

age 6*7 of 6*8#ile4ame 59*+6C+68.xls

0orksheet Ht Exchanger Temperatures

Source( Chemical nineerin aaKine; 0lant 2otebook Section; 1nkno%n date

N. *. 0etrosky; 4ulcan aterical Co. Eichita, /ansas

Direct Calculation o Exchanger Exit Te;erature:

!n speci&yin heat exchaner sevices &or process desin, it is &re3uently necessary to arive at optimum con

throuh trial and error. o%ever, the determination o& each set o& condtions %ithin this trial-and-error also i

calculation o& interrelated variables, such as inlet and outlet temperatures and area; and this can result in

trial-and-error calculations %ithin the trial-and -error &or the optimum. !t is, thus, convenient to be able to c

exchaner outlet conditions directly, based on kno%n or assumed values o& inlet temperatures, speci&ic he

&lo%in 3uantities, overall trans&er rate, and sur&ace. Such a direct calculation is developed as &ollo%s and

in the sketch.

2omenclature(

3 T eat duty, Btu>hr or kcal>hr T 1000000

Constant or averae speci&ic heat on the shell side, Btu>lb or kcal>k T 0.$000

Constant or averae speci&ic heat on the tube side, Btu>lb or kcal>k T 1.0000

E T Fluid mass &lo% rate in shell side, lb>hr or k>hr T 100000

% T Fluid mass &lo% rate in tube side, lb>hr or k>hr T #$000

1 T 12$

A T 00.0

2$0

8$

Subscript denotin inlet conditions

Subscript denotin outlet conditions

From the derived e3uations, let( T $0000

B T #$000

C T 0.200###

*here&ore,

1$2

Cp

cp

Overall heat trans&er coe&&icient, Btu>hr-&t$(o# or kcal3hr(m$(o$

*otal exchaner heat trans&er area, &t$ or m$ T

*: Shell-side &luid temperature, o# or o$

t: *ube-side &luid temperature, o# or o$

∆Tm 7o mean temperature di&&erence, o# or o$

:

$

*$ t

: >6 ( $? ( T

:>' ( A?3>A ( '$? o# or o$

3, 1, A, ∆*m

%, cp, t

:

E, Cp, *

$

E, Cp, *

:

%, cp, t

$

C =eUA( 1Z

−1

B )




<ev +

age 6** of 6*8#ile4ame 59*+6C+68.xls

0orksheet Ht Exchanger Temperatures

3uations and their derivations(

dtions

nvolves 7et(

B T

lculate T

ts,

sho%n Combinin both above e3uations,

*he heat trans&erred to the tube-side &luid T 3 T %# cp? >t

$ ( t

:?

*he heat trans&erred to the shell-side &luid T 3 T E# Cp? >T

: (T

$?

%# cp?

E# Cp?

The heat transferred is also=UA ΔT m=

UA [ ( T

1−t

2)−(T 2−t

1)

ln( T

1−t

2 )( T

2−t

1 ) ]

t 2=(

Z B )(T 1−T

2 )+t 1

Z (T 1−

T 2)=

UA

[ ( T

1−t

2 )−( T 2−t

1)

ln (T 1− t 2 )(T

2− t

1 ) ]ln[T

1−(Z

B )(T 1−T

2)−t 1

(T 2−t

1) ]=UA [T 1−( Z

B )(T 1−T

2)+t 1−(T

2−

Z (T 1−T

2) ln [

T 1−T

1 (Z

B )+T 2 (Z

B )−t 1

(T 2−t 1) ]=UA [

T 1−T

1 (Z

B )+T 2 (Z

B )+t 1−T 2+Z (T 1−T 2 )

ln [T 1 (1−Z

B ) +T 2 ( Z B )−t 1

(T 2−t 1 ) ]=UA [ T 1− (Z B )(T 1−T

2)−T 2

Z (T 1−T 2) ]=UA

(T 1−T 2)−( Z B )Z (T 1−T

ln [T 1 (1−Z

B ) +T 2 ( Z B )−t 1

(T 2−t 1 ) ]=UA [ 1− (Z B )

Z ]=UA ( 1Z −




<ev +

−(T 2−t

1)t 2 )

t 1 ) ]

+t 1−(T

2−t

1)

) ]

process heat transfer hof master

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