Heat Exchanger Design CHE 311 Final Project MSU

8/13/2019 Heat Exchanger Design CHE 311 Final Project MSU

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Michigan State University

College of EngineeringDepartment of Chemical Engineering and

Material Science

Final Heat Exchanger Project:

Group: Flummoxed by Flux

CHE 311: Fluid Flow and Heat Transfer

December 6th

2013


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Executive Summary

Here we report the design and optimization process of a shell-and-tube heat exchanger which

effectively heats a stream of acetone flowing at 80,000 kg/hr from 10 C to 30 C using a stream

of DOWTHERM MX (DMX) flowing at 100,000 kg/hr with an inlet temperature of 100 C as

the heating fluid. This report explains the approach used to determine key system parameters,

such as the heat load, the outlet temperature of the heating fluid, and the Log-mean Temperature

Difference (LMTD), as well as the methods used to design a configuration that satisfies these

constraints. We also include several tables comparing the different configurations studied. All

exchangers are in the E-HS-2CN form, in TEMA nomenclature (E single shell pass, HShot

fluid shell side, 2CN2 tube pass countercurrent).

Energy balance calculations across the tube and the shell side yielded an outlet temperature of

81.02 C for DMX and an LMTD of 70.51 C. These values were used to design and compare

several shell-and-tube heat exchangers. The constraints under which the designs were evaluated

were the following: heat tranfer area, fluid velocity (for both sides), pressure drop (for both

sides) and LMTD correction factor, FG. Preliminary shortcut calculations using Aspen V8.2

suggested a 2-2 shell-and-tube configuration as the lowest area configuration (15.23 m2).

However, specific design constraints and optimization resulted in selection of a 1-2 shell-and-

tube configuration with an area of 22.60 m2

In the final heat exchanger configuration, the cold fluid (acetone) flows at 1.39 m/s through the

tube side. The tube side is composed of 190 2-meter long carbon steel tubes with an outer

diameter of and Birmingham Wire number (BWG) 18 arranged in a 0.938triangular pitch.

DMX flows at 1.73 m/s in the 0.440 m diameter steel shell side.

The convective heat transfer coefficients for the shell and the tube sides were calculated to be

1440 W/(m2C) and 2510 W/(m2C), respectively. Fouling was also accounted for both fluids,

using a value for hdiof 5000 W/(m2C) for (hdi-1=0.0002 m2C/W). Calculation of the overall

heat transfer coefficient (U0) gave a coefficient U0=621 W/(m2C). This value is within 6.5% of

the value calculated by Aspen (584.38 W/(m2C)).

A short discussion for future recommendations is also included at the end of the document.


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Design approach

The heat exchanger design approach required preliminary calculation to estimate the outlet

temperature for the hot fluid, DOWTHERM MX (DMX). This temperature was calculated with an

energy balance under the assumption that all heat lost by DMX is transferred to the cold fluid

(acetone), i.e. no energy is lost to the surroundings. The properties of the fluids at several

temperatures were exported from Aspen V8.2 and verified using external sources [1,2]. An Excel

spreadsheet containing both fluids properties at different temperatures was used to interpolate to the

outlet temperature (Table 1). The outlet temperature was used as an iterative basis to find the

temperature for which the average heat capacity of the heating medium and the heat capacity

required by the heat balance converged. This temperature was found to be 81.02 C. Obtaining the

outlet temperature for the heating fluid allowed calculating the LMTD. The LMTD was equated to

the heat load and used to calculate the product of the overall heat transfer coefficient (U 0) and totalheat transfer area (A0).

After the set of preliminary calculations, the parameters were used as inputs for in Aspen to

obtain an estimate of the overall heat transfer coefficient as well as the area. This process was

repeated for the following shell-and-tube configurations: 1-2, 2-2, 2-4. Double pipe heat exchanger

configurations were not considered as they would require either large areas or very high velocities to

meet the large flow rates in the present system.

Decreasing heat exchanger area typically results in reducing cost [3]. Therefore, the shell-and-tube configurations were preliminarily ranked in terms of their area. Given the importance of

reducing cost, the heat exchangers efficiency (from most economical to least economical) ranked as

follow: 2-2, 2-4, and then 1-2. The 2-2 shell-and-tube heat exchanger configuration was initially used

for the design. However, given the lack of rigorous LMTD correction data for this configuration, the

calculations were uncertain. In addition, the pressure drops across both sides of the 2-2 heat

exchanger were much larger than the 0.5 bar allowable pressure drop, so the configuration was

ultimately discarded. The next tested configuration was a 2-4 shell-and-tube heat exchanger. This

type of heat exchanger is typically very effective [4], however, the pressure drop under the design

conditions was again too large. The most appropriate configuration was obtained using a1-2 heat

exchanger. Despite the relative large area of this configuration (as compared to the previously

proposed), the parameters used in this configuration met the specific requirements. This

configuration was further optimized to minimize the area and optimize the velocity of the fluids. A

comparison of all the different configurations tried can be seen in Table 5 in the Appendix.


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The three main constraints that must be maintained within their appropriate ranges are area,

velocity, and pressure drop. These constraints are controlled by optimizing parameters such as the

tube length. Qualitatively, velocity and pressure drop on the tube side increase with tube length while

area decreases. On the shell side, decreasing the baffle spacing or increasing the number of passes

increases pressure drop while decreasing area. As a result, all of these factors must be balanced in

order to keep the system economically feasible with both a reasonable area and a reasonable pressure

drop. A large pressure drop, for example, could require higher inlet pressure or incorporating a pump

downstream. Other parameters such as tube diameter, pitch, and bundle spacing (between tube

bundle and shell) similarly affect area and pressure. Using iterative calculations in Excel, a set of

geometric parameters was determined that fall within optimal ranges (pressure drop less than 0.5 bar,

tube velocity within 1-2 m/s, shell velocity within 1-2 m/s, [7]) while minimizing required area. For

the preliminary Excel calculation, an initial guess for Uowas selected based on Aspen shortcut

calculations. This initial Uo was used to set an initial heat transfer Aoand thereby calculate the

number of tubes (with a fixed length and tube diameter). An output Uo was calculated based on

film coefficients determined from Nusselt number correlations. For each iteration, the initial

guess of Uowas replaced with the calculated value until the two values converged. During these

Excel calculations, it was also determined that pressurizing the acetone stream to 2 bar would be

preferred as it would reduce the percent pressure drop. Fouling was selected based on typical values

for organic heat transfer fluids [3].

After these parameters were selected, the design was simulated with Aspen Detailed

Calculations. Initial values of baffle cut, nozzle diameters, and other geometric parameters were

determined using common values [3,4]. Within the detailed simulation, several iterations were made

to improve pressure drop and reduce area. These iterations included changes of nozzle diameters (to

determine appropriate shell-side pressure drops), length, and Birmingham wire gauge (without

changing heat transfer area this results in a of the decrease pressure drop due to the large cross

sectional area). After the iterations determined appropriate values, these values were inputted into

excel to illustrate the differences between the Aspen and excel calculations. The simulation was thenexported to Aspen Exchanger Design and Rating (EDR) for economic optimization. Aspen EDR

compared several designs, choosing one that minimized cost to $19,744 (see Appendix C). For

specific geometry parameters see Table 6, Appendix D and the Aspen output in Appendix C.


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Results

Table 1. Component Properties of DMX and Acetone [1,2]

Table 2. Summary of Key Heat Transfer and Flow Properties

Heat Transfer and

Flow Properties

Excel

Calculations

Aspen

Simulation

%

Mismatch

Aspen EDR

Design

% Mismatch

from simulation

Required Heat

Transfer Ao(m2)

-- 22.42 -- 21.33 -4.8

Actual Heat TransferAo(m

2)22.64 22.74 0.44% 24.51 7.7

% Over Design -- 1.45 -- 14.5 100

Uo(W/m C) 621 584 -6.28% 615 5.2

q (kW) 970 919 -5.53% 919 0.0

LMTD (K) 70.51 70.99 0.68% 70.03 1.4

FG 0.98 0.99 0.82% 0.99 0.0

DMX Outlet Tb(C) 81.02 81.99 1.18% 81.99 0.0

hi(W/m C) 2512 2894* 13.2% 2225.8 -23.1

hdi(W/m C) 5000 5000 -- 5000 --

ho(W/m C) 1436 901* -59.4% 884 -1.8

hdo(W/m C) 5000 5000 -- 5000 --

pt(bar) 0.07 0.05 -30.3% 0.13 132.8

pshell(bar) 0.54 0.13 -324.4% 0.08 -37.3

*Found using Aspen Exchanger Design and Rating (EDR), but we were unable to reproduce

exact geometry, explaining the high % errors. Detailed simulations in Aspen Plus did not

provide specific film coefficients.

ComponentProperties

DOWTHERM MX[1] Acetone[2]

Evaluated at T (C): 40 100Interpolated at:

90.5120

kg/m3 948.2 905.2 912.0 786

k W/mK 0.121 0.114 0.115 0.17

Cp J/kgK 1675 1870 1839.2 2182

103kg/ms 9.44 2.09 3.25 0.36

MW (g/mol) -- -- 58.08 238

Pr (Cp/k) -- -- 4.62 51.97


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Sources of Error

Between the hand calculations and the Aspen outputs, there is consistently an error of

approximately 5% for the overall heat transfer coefficient, Uo, and the heat load, q. It is

suspected that the material properties are the main source of error. When evaluating theproperties of DMX at its arithmetic average temperature (90.1C), a basic interpolation method

was used based on two data points at 40C and 100C. With such a wide range, interpolation can

become inaccurate, particularly if the properties do not vary linearly in this range.

The pressure drop calculation for the shell side also had a consistently high error of about 300%

(see Table 2) between Excel calculation and Aspen calculations. It is suspected that the empirical

relationships between geometry and pressure drop used in Aspen are more refined then those

used in the Excel calculations (See Appendix B). For example, Aspen takes into account the

affect of spacing between the baffles and the shell, whereas the relationship used for Excel

calculations is only dependent on equivalent diameter, pitch, and baffle spacing.

Recommendations for Design and Further Optimization

Based upon analysis from Aspen Plus, Aspen EDR, and preliminary Excel calculations, we

recommend that the heat exchanger follow the geometry parameters specified in Table 6,

Appendix D. This design offers the lowest heat transfer area of 22.6 m2while still satisfactorily

meeting heat load requirements. The economically optimized design suggested by Aspen EDR isa possible option as well, though has a higher heat transfer area of 24.5 m2. If a factor of safety is

desired, taking into account leaking and additional fouling that may develop over years of use,

then the Aspen EDR design may be preferred as it is overdesigned by 14.5%. However if volume

and total sizing is a limiting factor, then the design in Table 6 would be preferred.

It should be noted that both designs adequately meet all requirements but are not entirely

optimized. A further economic optimization route would involve increasing tube length to 2.44

m to match industry standards and reduce waste in cut materials. Increasing the length wouldalso decrease number of tubes and required shell diameter, also cutting costs. We recommend

that more detailed calculations be made in Aspen Plus in combination with Aspen EDR in an

effort to decrease total heat transfer area. The current design was more conservative than perhaps

necessary due to the high pressure drop calculated by hand for the shell side. Recognizing that

the Aspen outputs show a lower pressure drop than the hand calculations, this design could be


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made more efficient by decreasing the heat transfer area. One potential route for improvement

could be an increased number of baffles in order to increase cross flow and the overal heat

transfer coefficient.

References

1. Dow Chemical. Dowtherm MX Heat Transfer Fluid Product Technical Data. The Dow

Chemical Company, 1999.

2. Engineering ToolBox. 5th December 2013 .

3. Sinnott, R. K. Chemical Engineering Design. Ed. 4th. Vol. 6. Elsevier, 2005.

4. Macabe, W., J. Smithm and P. Harriott. Unit Operations of Chemical Engineering. McGraw-

Hill, 2001.

5. Briedis, Daina. Course Handouts, CHE 311: Fluid Flow and Heat Transfer. Michigan State

University. Fall 2013.

6. Mukherjee, Rajiv. "Effectively Design Shell-and-Tube Heat Exchangers." Chemical

Engineering Progress (1998).

7. Than, Su Thet Mon, Khin Aung Li and Mi Sandar Mon. "Heat Exchanger Design." World

Academy of Science, Engineering and Technology (2008): 604-611.

8. Faghari, A. and Y. Zhang. Transport Phenomena in Multiphase Systems. Burlington: Elsevier,

2006.


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Appendix A: Design Calculation Examples

Table 3.Sample calculations using Excel.

B C D

2 Properties

evaluated at

T(C)

20 90.5095

3 Component

PropertiesAcetone DMX

4 kg/m3 786 912.00

5 k (W/mK) 0.17 0.115

6

Cp (J/kgK) 2181.818182

1839.155875

7 (kg/m*s) 3.60E-04 3.25E-03

8 (/) (m2/s) 4.580E-07 3.566E-06

9 Pr (Cp/k) 4.62E+00 5.20E+01

10 MW (g/mol) 58.08 238

11 Values Cell references

12 Intial: Ao(m2)

=q/(U0*LMTD*F

G)

22.60 (Heat Load)/(LMTD*D$13*D$14)

13 Initial Guess Uo

(W/m2C)

621 Initial guess retrieved from Aspen

14 Fg 0.98 Retrieved from McCabe (Figure 15.6) [4]

15 Calc: Ao(m2) 22.60 (Heat Load)/(D$17*LMTD*D$14)

16 Cal: Ao(ft2) 243.264 (Heat Load)/(D$17*LMTD*D$14)* 10.7639

17 Calc 1/Uo

(W/m2C)

1.61E-031/D$60+1/D$61+D$24*LN(D$24/D$25)/(2*D$21)+(D$24/D

$25)*(1/D$39)+(D$24/D$25)*(1/D$38)

18 Calc Uo

(W/m2

C)

6.21E+02 1/D$16

19Calc Uo

(BTU/ft2F)

0.030399337

1/D$16/20248

20 Values Cell references

21 Tube Fluid Acetone Acetone


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22 Tube Material Steel Steel

23 kw(W/mC) 52 Retrieved from Engineering Toolbox [2]

24 v (m/s) 1.390 mdot/($C$4*D32)

25 BWG 18 Retrieved from McCabe (Appendix)[4]

26

27do(m) 0.01905

Retrieved from McCabe (Appendix) and converted

(0.75in*0.0254m/in)

28 di(m) 0.0165608 Retrieved from McCabe (Appendix) [4]

29 t (m) 1.24E-03 (D24-D25)/2

30 L (m) 2 2

31 Total Nt 189 D12/D30

32 Np 2 2

33Heat Atube(m

2

) 1.20E-01 PI()*D24*D2734 Flow Atube(m

2) 2.15E-04 PI()*D25^2/4

35 Total Aflow(m2) 2.03E-02 D28*D31/D29

36 Pitch Type TRIANGULAR Selected

37 PT(m) 2.38E-02 D24*1.25

38 C=PT- do (m) 4.76E-03 D34-D24

39 Re 5.03E+04 ($C$4*D25*D22)/$C$7

40 Nu 2.45E+02 0.023*D36^0.8*$C$9^0.4

41 hi(W/m2C) 2.51E+03 D37*$C$5/D25

42 hi(BTU/ft2F) 442 D37*$C$5/D25*0.1761

43 hdi(W/m2C) 5000 Retrieve from Sinnott (Table 12.2)

44 hdi(BTU/ft2F) 5000 5000*0.1761

45 ft 9.00E-03 Retrieved from Table 5.53 (McCabe)

46 pt(Pa) 7.73E+03 ($C$4*D22^2)/(2)*((D40*D27)/D25+4)*D29

47 preturn(Pa) 6.08E+03 4*D29*($C$4*D22^2)/(2)

48 Total pt(bar) 0.14 (D41+D42)/10^5

49 Total pt(psi) 2.0305 (D41+D42)/10^5*14.5037

50 Pin (bar) 2.026 1.013*2

51 %p drop 6.82% D43/D44


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Appendix B: Equations Used*

q =m H=mCp T; m = VAflow Tlm=

(Thb Tcb ) (Tha Tca )ln (Thb Tcb )

(Tha Tca )

Q = UAFG Tlm Z=

Tha Thb

Tcb Tca; H= Tcb Tca

Tha Tca;FG determined from figure 15.6 [5]

Heat transfer Reynolds #: Re tube =r VtDi ; Reshell=

r VshDe ;De= 4A

flow

LHTperimeter

Flow Reynolds: Reshell=r V

sh

DH ; Triangle Pitch: DH=

3.44PT

2

p do do

1

U0=

1

hi

r0ri+

1

hdi

r0ri+

rk

r0rlm

+1

hd0+

1

h0;hdfouling

Nutube=hidikf

= 0.023Re0.8 Pr0.3

Nushell=hoDekf

= 0.36Re0.55 Pr1/3

ptubes= NpftLtDit

+ 4

Vt22gc

pshell= DsDe

Vsh22

fs(N

b+1); fs= exp(0.576 0.19ln(Resh )); Nb= LtB

1;For 1-2 HX:Ds= Dbundle+ 0.056;Dbundle=do

Nt0.249

1/2.207

Shell Flow AreaAs=DsCB

PT

; Pitch PT=1.25do; C=PT do; B = Baffle Spacing

*All equations and variable definitions used in this report are taken from course handouts (e.g.

Heat Transfer Help Sheet [8]) or can be found in the course textbook, McCabe [5].


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Appendix C: Spreadsheets and Aspen Outputs

Table 4. System parameters

System Properties Acetone DMX

(kg/hr) 80000 100000(kg/s) 22.22 27.78

Ta(C) 10 100

Ta(K) 283.15 373.15

Tb(C) 30 81.019

Tb(K) 303.15 354.17

q (MW) 0.9697 -0.9697

LMTD (K) 70.51

Table 5. Comparison of Preliminary Shortcut Calculations and Finalized Design

Heat Transfer Ao(m ) Uo(W/m C)

Shortcut

Calculation

1-2 HX 18.8 730

2-2 HX 15.2 850

2-4 HX 15.2 850

Hand Calc 1-2 HX 22.6 620

Detailed Calc 1-2 HX 22.7 580

Figure 1. Temperature Profile Plot

Stream Temperatures

TS Bulk Temp. SS Bulk Temp

Distance from End (mm)

0 200 400 600 800 1000 1200 1400 1600 1800 2000

10

20

30

40

50

60

70

80

90

100


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Aspen Simulation Output

*** OVERALL RESULTS ***

DUTY AND AREA:

CALCULATED HEAT DUTY WATT 918904.8519

CALCULATED (REQUIRED) AREA SQM 22.4166

ACTUAL EXCHANGER AREA SQM 22.7420

PER CENT OVER-DESIGN 1.4520

HEAT TRANSFER COEFFICIENT:

AVERAGE COEFFICIENT (DIRTY) WATT/SQM-K 584.3846

AVERAGE COEFFICIENT (CLEAN) WATT/SQM-K 584.3881

UA (DIRTY) J/SEC-K 13099.8882

LOG-MEAN TEMPERATURE DIFFERENCE:

THERMAL EFFECTIVENESS (XI) 0.2221

NUMBER OF TRANSFER UNITS (NTU) 0.2850

LMTD CORRECTION FACTOR 0.9881

LMTD (CORRECTED) K 70.1460

NUMBER OF SHELLS IN SERIES 1

NUMBER OF SHELLS IN PARALLEL 1

STREAM VELOCITIES:

SHELLSIDE MAX. CROSSFLOW VEL. M/SEC 0.7415


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SHELLSIDE MAX. CROSSFLOW REYNOLDS NO. 5046.7091

SHELLSIDE MAX. WINDOW VEL. M/SEC 0.9854

SHELLSIDE MAX. WINDOW REYNOLDS NO. 6706.7435

TUBESIDE MAX. VELOCITY M/SEC 1.3675

TUBESIDE MAX. REYNOLDS NO. 55901.4452

PRESSURE DROP:

SHELLSIDE, BAFFLED FLOW AREA N/SQM 12041.7048

SHELLSIDE, NOZZLE N/SQM 681.3618

SHELLSIDE, TOTAL N/SQM 12723.0666

TUBESIDE, TUBES N/SQM 5012.7961

TUBESIDE, NOZZLE N/SQM 357.5980

TUBESIDE, TOTAL N/SQM 5370.3941


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Appendix D: Mechanical Summary

Table 6. Key Geometric and Design Parameters

Geometric Parameter Dimension Geometric Parameter Dimension

vtube (m/s) 1.4 Pitch Type Triangular

Tube Material Carbon Steel Tube Pitch PT(m) 0.0238

Tube BWG 18 Total Tube Aflow(m ) 0.0203

Tube do(m) 0.01905 Dbundle(m) 0.384

Tube di(m) 0.0166 Shell diameter DS(m) 0.440

Tube Length (m) 2 vshell (m/s) 1.7

Tube Number Nt 190 Baffle Number NB 9

Tube passes Np 2 Shell passes Np 1

Baffle Spacing B (m) 0.2

Figure 2. Tube Layout generated by Aspen EDR

Figure 3. Shell and Baffle Layout generated by Aspen EDR

99.95

m

m

120.57

m

m

-

. .

7.

7.

7.

.

.

. . .

.

.

. 7

.7 7

.7 7

7

.

.

75

75

2969 Overall

339 223 220 1555

223 461 1074

Pulling Length 1460

T1

T2 S1

S2

A

379

379

T2

T1

379

379

S1

S2

416

Views on arrow A


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Appendix E: Validation of Answer and Check of Realistic Results

The specifications of this design are considered to be realistic based upon typical values

published in several heat exchanger design guides [3,6,7]:

Optimal tube-side velocities for acetone are typically between 1.5 and 3 m/s, asdescribed in Table 8.2 in the course handout [5]. The velocity in this design is 1.4 m/s,

which is slightly lower than this suggested range. However, another design guide

suggested a reasonable range of 1-2 m/s [7], and it was found in this case that a slightly

lower tube-side velocity resulted in a more reasonable pressure drop.

Optimal shell-side velocities for benzene are typically between 1.4 and 2.8 m/s [Table

8.2, ref. 5]. As DMX is an alkylated aromatic, it likely has similar behavior and viscosity

as benzene, so this range should serve as a good approximation. The shell side velocity

for this design is 1.7 m/s, which is taken to be reasonable.

As these streams are at relatively low pressure, a reasonable pressure drop for both the

shell and tube side should be below 10 psi or about 0.5 bar. In this design, pressure dropis maintained well below 0.5 bar for both streams (see Table 2).

This tube uses a nominal outer diameter tube of carbon steel. Carbon steel tubes are

inexpensive and the diameter is common and preferred for compact exchanger

designs with low-fouling fluids [7].

The optimum ratio of tube length to shell diameter is typically between 5 and 10 [7]. For

this design the ratio is 4.5. This is less than optimal, but still reasonable.

The Shell diameter is typically about 56 mm larger than the bundle diameter [7], which

is determined by the number and size of tubes. This spacing is specified to both reduce

shell-side pressure drop and to increase the ease of construction and maintenance of theexchanger [6].

The preferred tube lengths are: 1.83 m, 2.44 m, 3.66 m, 4.88 m and 7.32 m [7]. The tube

length selected is 2 m. This is not ideal due to the English basis of many tubes. A future

design optimization for economics would likely take this into account and adjust the

length to 2.44 m. However, the length is reasonable from a sizing perspective.

For the heating requirements of this system, a 1-2 heat exchanger is appropriate. A 2-4

heat exchanger may decrease area requirements, but the complexity of construction and

maintenance are likely economically prohibitive and unnecessary for the relatively low

heat load.

Heat Exchanger Design CHE 311 Final Project MSU

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