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International Journal of Mechanical Engineering and Technology (IJMET)

Volume 8, Issue 5, May 2017, pp. 995–1009, Article ID: IJMET_08_05_104

Available online at http://iaeme.com/Home/issue/IJMET?Volume=8&Issue=5

ISSN Print: 0976-6340 and ISSN Online: 0976-6359

© IAEME Publication Scopus Indexed

NUMERICAL INVESTIGATION ON HEAT

TRANSFER AND FLUID FLOW OF SHELL-SIDE

FOR SHELL AND TUBE HEAT EXCHANGER

WITH HEXAGONAL VENT BAFFLE BY USING

CFD

G. Vijay Teja

M.Tech. Department of Mechanical Engineering, K L University,

Vaddeswaram, Guntur District, AP, India

Dr. K.V. Narasimha Rao

Professor, Department of Mechanical Engineering, K L University,

Vaddeswaram, Guntur District, AP, India

ABSTRACT

Shell and tube heat exchangers with many unconventional baffles are used in

industrial applications to increase the efficiency of the plant. Recently tre-foil hole

baffles are developed for which heat transfer coefficient is higher. In this context,

research is carried out on hexagonal vent baffle and the results are noted. Streamline

flow with different parameters were tested and their effects are noted on heat transfer

coefficient and consistency for different positioned tubes. ANSYS Fluent CFD

commercial package is used with realizable k-ε model. For method standardization,

analytical investigation is carried out to validate the numerical results. The results are

showing that the heat transfer coefficient is higher for hexagonalvent baffle

irrespective of positions of hexagonalvent on different tubes. Pressure drop and wall

temperature are found to befluctuating throughout the length in hexagonal vent and

optimum hydrodynamic performance observed in this configuration.

Keywords: Hexagonal-Vent baffle, shell-and-tube heat exchanger, shell side, heat

transfer enhancement, hydrodynamic performance, turbulence intensity, vorticity,

effectiveness.

Cite this Article: G. Vijay Teja and Dr. K.V. Narasimha Rao. Numerical

Investigation on Heat Transfer and Fluid Flow of Shell-Side for Shell and Tube Heat

Exchanger with Hexagonal Vent Baffle by using CFD. International Journal of

Mechanical Engineering and Technology, 8(5), 2017, pp. 995–1009.

http://iaeme.com/Home/issue/IJMET?Volume=8&Issue=5

Numerical Investigation on Heat Transfer and Fluid Flow of Shell-Side for Shell and Tube Heat Exchanger

with Hexagonal Vent Baffle by using CFD


1. INTRODUCTION

Heat exchangers contribute significantly to many energy conversion processes. Applications

range from food processing industries, nuclear power plants, offshore industries,

pharmaceutical production to aviation industries [1]. Recent developments in other exchanger

geometries have come in various industry applications. However, the shell-and-tube heat

exchanger by far remains the industry choice where reliability and maintainability are vital

[2].Because of their feasibility of desirable design considerations and ranges now-a-days,

CFD is playing very important role for parametric design process [1]. Although it is relatively

simple to adjust the tube side parameters, it is very hard to get the right combination for the

shell side. If possible, an ability to visualize the flow and temperature fields on the shell side

can simplify the assessment of the weaknesses, thus directs the designer to the right direction.

CFD can be very useful to gain that ability. Here the model made in CATIA and CFD

simulation is used to investigate the heat transfer and fluid flow in Shell and tube heat

exchanger with hexagonal vent baffles. Staggered tube bank with triangular pitch layout is

used, which is better for heat transfer and surface area per unit length[3].Wealth of literature

and theories are available to design a heat exchanger according to the requirements. A good

design referred to a heat exchanger with least possible area and pressure drop to fulfil the heat

transfer requirement[4].CFD is the science of predicting fluid flow, heat and mass transfer,

chemical reactions and related phenomena by solving numerically the set of governing

mathematical equations, which is stated in the ANSYS training module [5]. For method

standardization, analytical investigation is carried out by varying other parameters to estimate

performance of design using Kern method [6].

2. LITERATURE REVIEW

Over the years, significant research and development efforts devoted to better understand the

shell-side geometry. New geometries are been introduced for performance enhancement and

improve reliability. The pioneering works published in the Trans. Institute of Chemical

Engineers during May 1990, on helical baffles paved the way to a major shift from a

conventional understanding of baffles in a shell-and-tube heat exchanger [2]. Helical baffles

serve as guide vanes for shell-side flow as compared to creating flow channels with

conventional segmented baffles. In the past decade, heat transfer has extended the

understanding of the helical baffle geometry through extensive testing and development

[2].Helical geometry gave better results than convectional baffles. More recently, many

researchers started working in this field of redesigning of baffles modeling.

Shell-and-tube heat exchangers with trefoil-hole baffles are new type heat transfer devices

and are widely used in nuclear power plants due to their special advantages, with the fluid

flowing longitudinally on the shell side. However, very little related literature is available. In

order to obtain an understanding of the underlying mechanism of shell-side thermal

augmentation, a CFD model including inlet and outlet nozzles was proposed in the study

[7].Based on the RNG k-Ɛ model, numerical investigations on shell-side fluid flow and heat

transfer are conducted by using commercial CFD software FLUENT14.0. The results show

that the fluid is fully developed after the first trefoil-hole baffle. The heat transfer coefficient

and pressure drop vary periodically along the axial direction. Fluid velocity increases

gradually and the jet flow forms in the region near baffles. The secondary flow is also

produced on the two sides of baffles when the fluid flows through trefoil-hole baffle. The jet

flow and secondary flow can decrease the thickness of boundary layer and then enhance the

heat transfer [7].

G. Vijay Teja and Dr. K.V. Narasimha Rao


In another paper, three-dimensional CFD simulations using the commercial software

ANSYS 15.0-FLUENT, have been performed to study and compare the shell-side flow

distribution, heat transfer coefficient and the pressure drop between the recently developed

trefoil-hole, helical baffles and the conventional segmental baffles, at low shell side flow rates

[8]. In this numerical comparison, the whole heat exchangers consisting of the shell, tubes,

baffles and nozzles are modeled; the numerical model predicts the thermo-hydraulic

performance with a considerably good accuracy, by comparing with experimental data for

single segmental baffles. The model is then used to compute and compare the thermo-

hydraulic performance for the same heat exchanger with trefoil-hole and helical baffles. The

results show that the use of helical baffles results in higher thermo-hydraulic performance

while trefoil-hole baffles has a higher heat transfer performance with large pressure drop

compared to segmental baffles where thermo-hydraulic performance is high in helical baffle

[8]. Hence, there is need of additional modifications in the baffle changes and comparing

them each other.

In this present work, a new baffle hole geometry is designed and studied. The results are

obtained by using ANSYS 15.0 and analytical investigation is carried to standardizing the

analysis.

3. GEOMETRICAL MODEL AND MESHING

The shell side design of a shell-and-tube heat exchanger; in particular, the baffle spacing, and

shell diameter dependencies of the heat transfer coefficient and the pressure drop are

investigated by numerically modeling a small heat exchanger. The flow and temperature

fields inside the shell are resolved using a commercial CFD package by varying the mesh size

as in Figure1to obtain optimization mesh size. Set of CFD simulations have been performed

for a single shell and single tube pass heat exchanger with a variable number of baffles and

turbulent flow. The results are observed to be sensitive to the turbulence model selection. The

best turbulence model among the ones considered is determined by comparing the CFD

results of heat transfer coefficient, outlet temperature and pressure drop with the Realizable k-

ε model method results. The effect of baffle spacing to shell diameter ratio on the heat

exchanger performance is investigated by varying flow rate and optimization is obtained.

Figure 1 Cold fluid outlet Temperature for different mesh densities

The entry and exit points for tubes are exactly starting and ending with length of the heat

exchanger respectively, ends for shell along the length are chopped off at designed length, i.e.

no D-ends or to roid ends are used for simulation purpose. Geometrical details are given in

250

260

270

280

290

300

310

320

330

CO

LD F

LUID

OU

TLET

TEM

PER

ATU

RE

NUMBER OF ELEMENTS

EXPERIMENTAL

C.F.D




Table 1.The baffle plate thickness is kept at 4mm.Shell side fluid must flow aligned to the hot

fluid tubes to accomplish this condition. Cross-section of baffles has to change in such a way

that the turbulence should be kept uniform in baffle contact point on tubes. From this concept,

hexagonal shape of baffles are modeled. Number of elements used is 2595539; number of

nodes is 657982. Maximum Skewness obtained is 0.86 which is lesser than the acceptable

limit of the overall skewness of 0.9[9].The standard deviation is found to be 0.1308,which is

negligible.

Table 1 Geometrical parameter of hexagonal vent

Shell Diameter 150 mm

Shell Inlet Diameter 52.5 mm

Shell Outlet Diameter 52.5 mm

Tube Internal Diameter 15.798 mm

Tube External Diameter 19.1 mm

Number of tubes, baffles 7,6

Distribution Rotated Circular Type

Baffle Pitch 86 mm

Tube Pitch 42 mm

Baffle Type Rotated Circular Type

Baffle Thickness 4 mm

Vent circle radius. 13.8242 mm

parallel face distance 24.053 mm

Figure 2 Geometrical view of hexagonal vent



4. GOVERNING EQUATIONS

The fluid flow assumed study-state turbulent model with incompressible fluid. The shell side

fluid flows only through the hexagonal vents and it is assumed that three are no other

leakages. Finite volume method is adopted to solve the model equations like Continuity

equation (1) momentum equation (2) and energy equation (3). The equations are given below

[10, 11]: 𝜕𝑢𝑖

𝜕𝑥𝑖= 0 (1)

𝜕𝑢𝑖𝑢𝑗

𝜕𝑥𝑖= −

1

𝜌

𝜕𝑝

𝜕𝑥𝑖+

𝜕

𝜕𝑥𝑖{(𝑣 + 𝑣𝑡𝑢𝑟𝑏) (

𝜕𝑢𝑖

𝜕𝑥𝑖+

𝜕𝑢𝑗

𝜕𝑥𝑖)} (2)

𝜕𝑢𝑖𝑇

𝜕𝑥𝑖= 𝜌

𝜕

𝜕𝑥𝑖{(

𝑣

𝑝𝑟+

𝑣

𝑝𝑟𝑡𝑢𝑟𝑏)

𝜕𝑇

𝜕𝑥𝑖} (3)

To understand exact performance of flow, realizable k-𝜖 model is chosen, experimental strong adverse gradient of pressure and recirculation [11], turbulent kinetic energyk (4) and

dissipation 𝜀 (5) whichhave effect on boundary layer transport equations given below:

𝜕𝑢𝑖𝑘

𝜕𝑥𝑖=

𝜕

𝜕𝑥𝑖((𝑣 +

𝑣𝑡

𝜎𝑘)

𝜕𝑘

𝜕𝑥𝑖) + Г − 𝜀 (4)

𝜕𝑢𝑖𝜀

𝜕𝑥𝑖=

𝜕

𝜕𝑥𝑖((𝑣 +

𝑣𝑡

𝜎𝑘)

𝜕𝜀

𝜕𝑥𝑖) + 𝑐1Г𝜀 − 𝑐2

𝜀2

𝑘+√𝑣𝜀 (5)

Where Г(6) represents k generation from mean velocity gradients given as:

Г = −𝑢𝑖𝑢𝑗̅̅ ̅̅ ̅𝜕𝑢𝑖

𝜕𝑥𝑖= 𝑣𝑡𝑢𝑟𝑏(

𝜕𝑢𝑖

𝜕𝑥𝑗+

𝜕𝑢𝑗

𝜕𝑥𝑖)

𝜕𝑢𝑖

𝜕𝑥𝑖 (6)

The turbulent kinetic viscosity is:

Figure 4 geometry view after generating in ANSYS

Figure 5 geometry view after mesh




𝑣𝑡𝑢𝑟𝑏 = 𝑐µ𝑘2

𝜀 (7)

Realizable 𝑘 − 𝜀 model considered is varying from standard RNG𝑘 − 𝜀 model due to functionality of 𝐶µ is no more considered as constant [8].𝐶µ depends upon mean strain, rotation rates, angular velocity and turbulence field and empirical constants of Realizable 𝑘 −𝜀 model [12] are given below:

C1=max [0.43,µ/(µt+5)]; C2=1.9; 𝜎k =1.0;𝜎𝜀=1.2

In addition, the second order upwind scheme has been adopted for the momentum, energy,

turbulence and its dissipation rate. All the convergence residuals are considered very less. The

main two variations of this realizable model are: new eddy-viscosity formula involving a

variable Cµ originally proposed by Reynolds and new model equation for dissipation based on

the dynamic equation of the mean-square vorticity fluctuation[12]

5. BOUNDARY CONDITIONS AND NUMERICAL METHODOLOGY

Data reduction is very important to eliminate unwanted values and calculations in any

engineering experiment. The calculation of experimental values for CFD simulation is carried

out at mass flow rate of 1kg/s to ensure proper turbulence. Details are given in Table5.

For Hexagonalvent baffles, calculation of shell side Reynolds Number:

Re= 4�̇�𝑠

𝜇𝑠𝜋(𝐷𝑠+𝑛𝑡𝐷0) (8)

Where, the turbulence intensity is calculated using equation (9) which gives the

percentage of the intensity. The present investigation yielded a value of 5.5, which is just

about medium turbulence case [13]:

𝑇𝑢𝑟𝑏𝑢𝑙𝑒𝑛𝑐𝑒 𝐼𝑛𝑡𝑒𝑛𝑠𝑖𝑡𝑦 (𝑇𝐼) = 0.16(𝑅𝑒)−1

8 (9)

Table 2 Turbulence intensity %.

Thermo-physical Properties of hot and cold fluids

Table 3 Fluids Material Properties

Fluid properties Cold side fluid (water at NTP) Hot side fluid (water at 80°C)

Density (kg/m3) 998.2 974

Cp (J/kg-K) 4178 4195.3

Conductivity (W/m-k) 0.6 0.67

Viscosity (Kg/m-s) 0.001003 0.000355

Solid Material Properties

Table 4 Solid Material Properties

Area Re TI %

Hexagonal vent 4487.8979 5.5926

Tube side 15817.8284 4.7777

Solid properties Copper Steel

Density (kg/m3) 8978 8030

Cp (J/kg-K) 381 502.48



Table 5 Boundary conditions

BC Momentum Thermal

Hot Inlet 0.36576 (m/s) 353 (K)

Cold Inlet 0.4673148 (m/s) 300 (K)

Hot Inlet Hydraulic Dia. 15.798 mm

Cold Inlet Hydraulic Dia. 52.5 mm

Outlet Pressure Outlet 300 (K)

Values of momentum, pressure and energy are chosen to be 0.7, 0.7 and

0.6respectively.At different mass flow rates, the flow will affect the heat transfer coefficient

and other properties of the stated problem. The mass flow rate is varied with intervals of 0.1

kg/s. Reynolds number is calculated using Eq. (8) and turbulent intensity-TI is calculated

using Eq. (9). The Nusselt number correlation is used to obtain the heat transfer coefficient

(Eq. 10). Prandtl number for shell-side fluid water at bulk mean temperature of 38.6960C is

4.5147. The Nusselt number correlation is given by [14]:

Nu= c (1.13 pr 0.4) Ren (10)

The values of ‘c’ and ‘n’ are taken from the data book by calculating staggered ratios, St/d

and Sl/d ratio, which isfound to be 2. C is 0.482 and n is 0.556 [14]. Using these values and

formula (Eq. 10), Nu number is calculated and thus heat transfer coefficient can be

determined.

Heat transfer coefficient(h) = 𝑁𝑢.𝑘

𝑑 (11)

Table 6 Data of hexagonal vent with various mass flow rates

For the configuration used in the model, the flow is laminar up to mass flow rate of 0.4

kg/s (Re




rate is used in the simulation. Where the turbulence intensity is above 5%, the flow is

completely developed flow due to jet kind of lay path. The heat transfer coefficients of the

model values are found to be increasing with increase in mass flow rate of shell side fluid.

Figure 6 Mass flow rate vs heat transfer coefficient

6. RESULTS AND DISCUSSION

Figure 7 shows the streamline plots of the aerodynamics of the baffle. The fluid is found to be

having higher velocity at baffle opening creating wake region immediately after the baffle.

Hexagonalvent having turbulence at inlet section and it is becomes more and more aligned

along the length of the heat exchanger.

0

50

100

150

200

250

300

350

400

450

0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1

Hea

t tr

ansf

er c

oef

fici

ent-

h (

w/m

2k)

mass flow rate (kg/s)

Figure 7 Streamline plot

Figure 8 Vector velocity plot



Figure 8 shows the Velocity vector plot to understand the velocity of each vector in the

flow. It is clear that a secondary flow occurs in shell have less flow velocity. The flow over

the tubes is fully developed immediately after every baffle. High velocity region is formed

which is indicated with red colour. Figure 8 shows the velocity after baffle, which is relatively

high. Recirculation zones can be indicated clearly in vector plots. Changes in direction of the

flow because of the disturbance offered by baffles are responsible for this recirculation zone.

The fluid flowing inside the shell wets the outer surface of the tubes, which lead to a

temperature change in the surface of the tubes. Figure 9 shows the temperature change is

higher after half of the flow passed and the temperature of the surface is fluctuating

throughout the length at each baffle. So, periodic transfer of heat occurs on the surface of

tubes. Maximum temperature change is observed in hexagonalvent configuration in the last

40% of tube length, where, the temperature change in shell side fluid is somehow uniform

along the length as shown inFigure10.Due to hexagonal vent configuration, temperature is

getting decreased near to the opening of the hole pattern of baffle, it is directing the incoming

flow towards the outer surface of the tubes. This change in temperature is gradually

decreasing and temperature distribution in all tubes is uniform that will prevent generation of

sudden thermal stresses.

Figure 11 shows that hexagonalvent baffles are generating vortex-generated area after

passing through opening of the baffle plate because of sudden expansion. Shell side fluid

temperature is not changing much in hexagonalvent baffles. The vent patterns to create the

disturbance are functioning satisfactorily and the vortex is found to be generating from each

face of each vent as shown in Figure 11. By understanding the Velocity plot (Figure8),

temperature plot (Figure9) and vortex plot (Figure11), it is evident that the hexagonal vent

baffles are yielding satisfactory performance.

Figure 9 Temperature plot on wetted

area of tubes Figure 10 Temperature distribution

in shell side fluid

Figure 11 Temperature Plot on Vortex Core Region




After every passage of baffle, variation takes place in the pressure of the shell side fluid

(Figure12). Pressure drop took place at each baffle in six steps. Even though it is a step-wise

decrease of pressure, just after passing from vent, the pressure is slightly increased for each

baffle as shown in Figure 12. Along the length of the heat exchanger, the pressure drop is not

very high and variation of pressure at inlet and outlet is not much different. Pressure drop in

shell side fluid is showing significant changes after half of the length of the heat exchanger.

The inlet pressure is 15.74 psi, which is dropped to 14.34psi at the outlet resulting in pressure

drop of 1.4 psi, which is very much in the limits.

.

Figure 12 Pressure drop variation along length

The shell side fluid taken at NTP, which is 300 K of water, which is coming out from

shell outlet at a temperature of 323.39 K as shown in Figure13. The raise of temperature for

cold fluid is 23 Kelvin, which is acceptable for this configuration. As can be seen from the

Figure 13, the shell side fluid found to be raising its temperature after travelling half of the

flow length.

The Heat transfer coefficient distribution of wet surface on shell side for each tube is

different which is fluctuating throughout the length, the heat transfer coefficient is different

for each tube of same position (Figures 14 to 20):

100000

102000

104000

-0.1 6E-16 0.1 0.2 0.3 0.4 0.5 0.6

Pre

ssure

(pas

cals

)

length along heat exchanger

hexagonal vent shell side pressure drop

Figure 13 Temperature raise of shell side fluid along length



Figure 14 Tube 1

Figure 15 Tube 2

Figure 16 Tube 3

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

8.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]

distance along flow direction(m)

Figure 14: Tube 1

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure 15: Tube 2

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure16: Tube 3




Figure 17 Tube 4

Figure 18 Tube 5

Figure 19 Tube 6

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure17: Tube 4

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

8.00E+03

9.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure 18: Tube 5

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure 19: Tube 6



Figure 20 Tube 7

7. CONCLUSION

The hexagonal vent baffles model is performing satisfactorily as expected due to increase in

turbulence and residence time of fluid in shell side. Besides, the shell side fluid gets in to

better contact with tube outer surfaces, which also results in higher heat transfer rate. The

streamlines of the flow pass above the tube surfaces. The thermo-hydraulic performance is

moderate with a turbulent intensity of more than 5% ensures that flow will be disturbed. The

velocity vectors show that velocity of flow just after the baffle vent increases compared to

flow before the baffle. The vertex generation occurs from each face of hexagonal vent create a

wake region around the tube that enhances the heat transfer. The temperature difference

attained is acceptable for this configuration, where the temperature of the wetted area on tube

shell-side increased due to decrease of thermal boundary thickness, which is caused due to

flow generation over the tube. The outside surface temperature of the tube wall fluctuates

throughout the flow direction. The heat transfer coefficient of the tube walls will be

fluctuating along flow direction, which is different for each location and for each tube surface.

The temperature change of shell-side fluid took place after travelling half of the length. The

pressure drop of the system is found to be within acceptable limits.

REFERENCES

[1] B.I. Master, K.S. Chunangad, A.J. Boxma, D. Kral, P. Stehlík, Most frequently used heat exchangers from pioneering research to worldwide applications, Heat Transfer Eng. 27

(2006) 4–11.2006/06/01.

[2] B.I.Master, K.S. Chunangad, V. Pushpanathan, Fouling mitigation using helixchanger heat exchangers, in: Engineering Conferences International, 2003.

[3] Digvendra singh1, Narayan Das Pal2, International Journal of Scientific Engineering and Applied Science (IJSEAS) – Volume 2, Issue3,March 2016 ISSN: 2395-3470, Designing

and Performance Evaluation of a Shell and Tube Heat Exchanger using Ansys

(Computational Fluid Dynamics) www.ijseas.com.

[4] P.K. Nag, Heat and Mass Transfer, McGraw Hill education (India) Private Limited, New Delhi

[5] Chapter 1-Introduction to CFD- Introduction to CFX by ANSYS training lectures. http://www.petrodanesh.ir/virtual%20education/mechanics/ansyscfx/lectures/cfx12_01_in

tro_cfd.ppt.

0.00E+00

1.00E+03

2.00E+03

3.00E+03

4.00E+03

5.00E+03

6.00E+03

7.00E+03

-2.00E-01 1.00E-15 2.00E-01 4.00E-01 6.00E-01

Wal

l H

eat

Tra

nsf

er C

oef

fici

ent

[ W

m^-2

K^-1

]


Figure 20: Tube 7




[6] Donald Q. Kern, Process Heat Transfer, 28th reprint 2015, Indian Edition, Mc-GrawHill Education.

[7] Guo-Yan Zhou, Jingmei Xiao, Lingyun Zhu, Juntao Wang, Shan-Tung Tu, The 7th International Conference on Applied Energy – ICAE2015,A numerical study on the shell-

side turbulent heat transfer Enhancement of shell-and-tube heat exchanger with trefoil

hole baffles.

[8] Anas El Maakoul, Azzedine Laknizi, Said Saadeddine, Mustapha El Metoui, Abdelkabir Zaite, Mohamed Meziane, Abdelatif Ben Abdellah, Numerical comparison of shell-side

performance for shell and tube heat exchangers with tre-foil hole, helical and segmental

baffles. http://dx.doi.org/10.1016/j.applthermaleng.2016.08.067.

[9] Arun Kumar Tiwari, Pradyumna Ghosh, Jahar Sarkar, Harshit Dahiya, Jigar Parekh.Numerical investigation of heat transfer and fluid flow in plate heat exchanger

using Nano-fluids. http://dx.doi.org/10.1016/j.ijthermalsci.2014.06.015.

[10] J. F. Yang, M. Zeng, Q. W. Wang, Numerical investigation on combined single shell-pass shell-and-tube heat exchanger with two-layer continuous helical baffles, Int. J. Heat Mass

Transfer, Vol.84 (2015), pp. 103–113.

[11] Fluent Inc., 12.4.3 Realizable - Model, FLUENT User’s Guide, 2006. [12] Fluent Inc., Chapter 10. Modeling Turbulence, FLUENT User’s guide, November 28,

2001

[13] ANSYS Fluent User's Guide, Release 15.0, November 2013, 6.3.2.1.3 turbulent intensity, Equation-6.4.3, p. 258.

[14] Heat and Mass Transfer Data Book, Seventh edition 2010,New Age International (p) limitedpublishing.

[15] Designing and Performance Evaluation of a Shell and Tube Heat Exchanger using Ansys (Computational Fluid Dynamics). B. Peng, Q.W. Wang, C. Zhang, G.N. Xie, L.Q. Luo,

Q.Y. Chen, et al., An experimental study of shell-and-tube heat exchangers with

continuous helical baffles, J. Heat Transfer 129 (2007) 1425–1431.

[16] Y. You, Y. Chen, M. Xie, X. Luo, L. Jiao, S. Huang, Numerical simulation and performance improvement for a small size shell-and-tube heat exchanger with tre-foil hole

baffles, Appl. Therm. Eng. 89 (2015) 220–228. 10/5.

[17] J.F. Yang, M. Zeng, Q.W. Wang, Numerical investigation on combined single shell-pass shell-and-tube heat exchanger with two-layer continuous helical baffles, Int. J. Heat Mass

Transfer 84 (2015) 103–113.

[18] Anil Waman Date, Introduction to Computational Fluid Dynamics, Cambridge University Press, New York (2005).

[19] S. V. Patankar, Numerical Heat Transfer and Fluid Flow, Series in Computational Methods in Mechanics and Thermal Sciences, Hemisphere Publishing Corporation,

Printed and bound in India by Replika Press Private Limited (2016)

[20] M. Necati Ozisik, Basic Heat Transfer, McGraw-Hill, Inc. (1997) [21] R.K. Shah and D.P. Sekulic, Fundamentals of Heat Exchanger Design, John Wiley &

Sons, Inc., New Jersey (2003)

[22] Durgesh J Bhatt, Priyanka M Jhavar, Prakash JSakariya, review paper on analysis of heat transfer in shell and tube type heat exchangersijsrdv2i6249, Volume 2, Issue 6,pp: 446-

453.

[23] M. M. Aslam Bhutta, N. Hayat, M. H. Bashir, A. R. Khan, K. N. Ahmad and S. Khan, “CFD applications in various heat exchangers design: A review”, Applied Thermal

Engineering, Vol.32, pp. 1-12 (2012).

[24] E. Ozden, I. Tari, Shell side CFD analysis of a small shell-and-tube heat exchanger, Energy Conversion & Management, Vol.51, pp.1004-1014 (2010).

[25] Dr. Zena K. Kadhim, Dr. Muna S. Kassim and Adel Y. Abdul Hassan, Effect of (MGO) Nanofluid on Heat Transfer Characteristics For Integral Finned Tube Heat Exchanger,

International Journal of Mechanical Engineering and Technology, 7(2), 2016, pp. 11-24.



[26] Abhishek Kr. Singh and Durg Vijay Rai, The Variation in Physical Properties Affects the Vertical Compressive Strength of The Rudraksha - Bead (Elaeocarpus Ganitrus Roxb).

International Journal of Mechanical Engineering and Technology, 7(3), 2016, pp. 276–284

[27] Abhay Sehgal, Vinay Aggarwal and Satbir S. Sehgal, Computational Analysis of Stepped and Straight Microchannel Heat Sink. International Journal of Mechanical Engineering

and Technology, 8(2), 2017, pp. 256–262.

NOMENCLATURE

Latin Symbols

Across-cross-flow area at the shell centerline, mm2

Ao-heat exchange area based on the external diameter of tube, mm2

B -Baffle spacing, mm

Cp-specific heat capacity, J/(kg. K)

Ci -coefficients in k-ɛ model

Ds-Internal shell diameter, mm

Do-external tube diameter, mm

Dct-outer diameter of central tube, mm

h -Average heat transfer coefficient, W/(m2 K)

k -Turbulent fluctuation kinetic energy, (m2/s2)

L -Tube total effective length, m

ṁ-mass flow rate, (kg/s)

nt-number of tubes,

Ncr -number of tubes in central row

Pt-tube pitch, mm

Pr- Prandtl number

Dp-pressure drop, Pa

Qave-average heat transfer rate, W

Re-Reynolds number

Tin-inlet temperature, K

Tout-outlet temperature, K

∆𝑇𝑚-Logarithmic mean temperature difference, K u -Average velocity, (m/s)

x; y; z -Cartesian coordinate

Greek Symbols

Γ-generalized diffusion coefficient ɛ -Turbulent kinetic energy dissipation rate, (m2/s3)

𝜆-Thermal conductivity, W/(m K) µ- dynamic viscosity, kg/(m.s)

𝑣-Kinematic viscosity, (m2/s) 𝜌-Density, (kg/m3) 𝜎𝑘-Prandtl number for k 𝜎∈-Prandtl number for ∈

Subscripts

in -Inlet

out -Outlet

s-Shell side

t -Tube side

turb- turbulent

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NUMERICAL INVESTIGAT ION ON HEAT TRANSFER AND FLUID F … · 2017. 5. 31. · limit of the overall...

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