Cfd Simulations of Thermal Performance of Bare and Finned ...ijens.org/Vol_18_I_01/182201-7373-IJMME-IJENS.pdf · compounding and the complex supercharging system. Often, turbocharging

International Journal of Mechanical & Mechatronics Engineering IJMME-IJENS Vol:18 No:06 57

182201-7373-IJMME-IJENS © February 2018 IJENS I J E N S

Cfd Simulations of Thermal Performance of Bare and

Finned Tube Heat Exchangers Used as a

Surpercharging Air Cooler for Marine Engine

JIHANI Youssef *, ADHAM Adam, MABSATE El Mostafa Mechanical and Energy Engineering Research Team: Modeling and Experimentation [ERG(2M)]

Mohammadia School of Engineers, Mohammed V University

Rabat, Morocco

* Corresponding author: [email protected]

Abstract— In this article, a numerical study of the hydrodynamic

and thermal behavior of a supercharging air cooler of a marine

engine with bare bank tube configuration is presented. The CFD

simulation of the air cooler with its actual configuration (bare

tube bank configuration) is implemented with relatively small

dimensions compared to the actual dimensions while keeping the

same ratio of the diameter and pitches tubes of the validation

models. A comparison of the efficiency of this cooler with another

of the same dimension and using plain finned tube configuration

towards heat exchange coefficient and pressure drop is then

carried out. The resolution of the conservation equations is

ensured by the ANSYS Fluent simulation code. The results

obtained show that the heat exchange coefficient and the pressure

variation are greater for plain finned tube configuration and the

efficiency index goes up with appending the fins.

Index Term- component; CFD; Fin-tube heat exchanger;

configuration of heat exchanger; Airside thermal characteristics;

I. INTRODUCTION

From its first use dating back to the beginning of the

20th century, the turbocharging remains the main solution

adapted to thermal engines for the recovery of a large part of

the energy of the fuel and to increase the efficiency and engine

performance. This device took several different types like the

turbocharging by exhaust gases driven by the engine,

turbocharging by free turbine, hyperbar charging, volumetric

compounding and the complex supercharging system. Often,

turbocharging by exhaust gases, which has become almost

generalize in high-power engines, and can be mounted on

marine propulsion engines or power generation, industrial

engines, is used.

The supercharging consists in increasing the engine

air supply. This allows to increase the Bmep (Brake mean

effective pressure) and it is the best compromise between

power and efficiency, on the semi-rapid marine engines type,

supercharging is realized by means of two turbochargers in

parallel driven by the exhaust gases and switched sequentially.

to low load of engine one of the two turbochargers is isolated

by two valves respectively placed on the inlet and exhaust

manifolds, the other turbocharger fed by the exhaust gases of

the two cylinders banks and guarantee the supercharging of

the engine, and during load increase, the automatic opening of

the valves activates the turbocharger previously isolated and

to operate the two turbochargers at the same time.

Such a system has the advantage of improving considerably

the air supply engine at partial loads and thus increase the

available power while reducing exhaust temperatures, fumes

and consumption.

At the admission, air is filtered and then compressed in

turbochargers, where it undergoes an important increase of the

temperature, it is therefore necessary to have a device for

cooling, the air cooler that’s.

By analyzing the energy balance of this type of engine (Fig.

01), it can be seen that a large part of the energy supplied by

the fuel is lost (54%) and almost 14% in the supercharger air

cooler. The improvement of this component is therefore

mandatory.

Air cooler is a kind of heat exchanger. It transfers the heat flow

carried by the supercharging air to the cold refrigeration water

through an exchange wall, without direct contact between the

air and the water. Its shell often takes a rectangular shape, and

it is filled by tubes in which the water passes, and in the space

between the tubes and the shell flows the air. It is often

mounted on the one of the air inlet caissons, it consists mainly

of:

- a bundle made up of expanded tubes and cooling

blades

- a water box blanking off the bundle and equipped

with two flanges permitting the inlet and outlet of

water

- a water box situated opposite to the box, which closes

the water system of the bundle.

In order to improve the performance of the thermal engines, it

is necessary to take into account the importance of the control

of the thermal organs and to oversight the flow of heat crossing

them. Then examination of the thermodynamic cycles

performed, like many studies. In this paper, we investigate the

thermal and hydraulic characteristics of an air cooler with bare

tube bank configuration mounted on the marine engine 12-

M26.2 and compare these features with another air cooler of

the same dimensions using plain finned tube configuration.

un faisceau assemblé comportant des tubes et des ailettes de

refroidissement a bundle (111) made up of expanded tubes

and

mailto:[email protected]



The study is made through the 3D modeling of AIRSIDE and

TUBESIDE with the simulation code ANSYS Fluent.

This type of modeling of heat exchangers whose air is one of

the refrigeration fluids or the object of refrigeration occupies

the interest of several researchers. A. Zhukauskas and R.

ULINSKAS [6] suggested a correlation for heat transfer and

pressure drop for in-line and staggered banks of tubes. They

set the Reynolds number (Re) and Prandtl number (Pr) from 1

to 2106 and from 0.7 to 104, respectively, with a wide range

of relative longitudinal and transverse pitches. Hui Li et al.[10]

the effect of the air inlet angle on the air–side performance of

plate-fin heat exchangers by means of CFD. They concluded

that the heat transfer coefficient and pressure drop expand

with air inlet angle increased, while the pressure drop increase

is much larger. Kang et al.[8] compared the heat transfer and

heat flow characteristics of different configurations of heat

exchangers used for advanced sodium-cooled fast reactor.

L.Gu et al.[2] proposed a numerical study of the airside

thermal characteristics of bare tube bank and finned tube heat

exchangers used in aero-engine cooling. Puterbaugh et al.[9],

Kays et al.[7] published experimental results of the heat

exchange coefficient and the pressure drop of a compact heat

exchanger calculated for some flow velocity. Bhuiyan[1]

analysed thermal and hydraulic performance of finned-tube

heat exchangers as a function of flow. Könözsy[3] proposed

a numerical investigation on various heat exchanger

performances to conclude an optimum configuration for

charge air cooler, oil and water radiators in F1 Sidepods.

J.C.Min [4] gave numerically the airside heat transfer and

pressure drop characteristics of wavy finned tube heat

exchangers. Jang et al. [9] compared the effects of tube layout

on heat transfer characteristics and found that the heat transfer

coefficient of staggered arrangement was more important than

that of in-line arrangement.

Fig. 1. Representation of the energy balance of a marine engine

Fig. 2. Variation of Prandtl number of air with temperature

II. MODEL DESCRIPTION

A. Physical model

The surface heat exchanger in the inlet air system is fresh

water cooled with tubes of corrosion-resistant material. It can

be either one or two stage type. The water passes through the

tubes while the air flows round the tubes. The inlet and outlet

for fresh water to the nozzle header are in one of the end walls.

The other end wall operates as a return header for the water

flow. Both headers have gaskets. The charge air for the engine

flows in at the forward end of the engine and passes through

the tubes to the air receiver.

The compressor output temperature at maximum load is about

200°C. To cool this air out of the compressor it passes through

a refrigerant in which it is cooled by water. The cooler brings

the charge air to a temperature of about 50 - 90 ° C.

The refrigerant bundle (4) is in a molded metal housing, and

has a cooling water inlet manifold (2) and an outlet water box

(3). The manifold (7) contains the output connections with

plugs circuit emptying (5) and (6). Thermometers and

temperature regulators may be fitted to the inlet and outlet on

the waterside. In that case, the temperature regulators balance

the flow between the high and low temperature circuits of the

cooling system. (Fig 3,4).

A washing system is provided on the airside of the heat

exchanger. The washing system consists, in principle, of a

pressure tank, which forces a cleaning fluid through spray

nozzles into the air inlet. This

system is used when the engine is in operation.

Fig. 3. Supercharging air cooler

Fig. 4. Longitudinal section of air cooler

B. Numerical simulation model

The numerical resolution of heat exchange and pressure drop

is conducted with CFD code ANSYS Fluent, the latter allows

to give exact solutions taking into account physical and

realistic phenomena on real geometry through the geometry

conceived with a scaling ratio. After the design of the

geometry with CAD software, we import it to ANSYS Fluent

to define the boundary conditions, the volumes, the zones of

46%

28.1%

14%

6.9%

3.9% 0.6% 0.5%

Representation of the energy balance of a marine engine

puissance effective

Perte par échappement

Réfrigérant d'air de suralimentation

Eau de refroidissement

Huile moteur

Huile de turbosoufflante

Radiation



contacts. It is then meshed by Fluent meshing (the

characteristics of the geometry are displayed in table 01). The

meshed domain is shown in figure 6.

The fluent sub models enabled for the simulation are:

TABLE I

GEOMETRIC FEATURE OF AIR COOLER

a. Governing equations

The numerical solution of the continuity, momentum and

energy equations are given by FLUENT User's Guide [12], the

air flow is steady and in turbulent regime, the thermal radiation

and the gravitational force are neglected:

Continuity equation: 𝜕uj

𝜕xj= 0 (1)

Momentum equation:

i j ji i

j j j j i

u u up u

x x x x x

(2)

Energy equation:

p j eff

j j j

TC U T K

x x x

(3)

The turbulence model used for this study is the standard k-ɛ

realizable. This model of turbulence is often used for industrial

applications.

The equations of the model are:

Transport equation for turbulent kinetic energy:

( ) 1j t k

j j k j

u k Pk k

t x x x

(4)

Transport equation for the turbulent energy dissipation: 2

1 2

( ) 1j t k

j j j

u PC C

t x x x k k

(5)

Where Pk is the producing term of turbulent kinetic energy

generated by mean velocity gradient with the empirical values

of k–ε model, C1, C2 and Cϒ are empirical constants:

1 20.09, 1.44, 1.92, 1.0, 1.3kC C C

Heat conduction equation in the solid domain:

2 2 2

2 2 20

T T T

x y z

(6)

b. Boundary conditions

The boundary conditions considered in the simulation are:

velocity inlet for flow velocities and temperatures of fluids,

and pressure outlet for fluid pressure at the outlet, convective

heat exchange is affected on the tube walls. Table 02 displays

the different values taken for each boundary condition.

TABLE II

BOUNDARY CONDITIONS

c. Geometry

The design of the geometry with the real dimensions given by

the plans of the constructor is carried out by the software Catia

with PART DESIGN and ASSEMBLED PART and then the

reduction to the scale is realized by GENERATIVE SHAPE

DESIGN, as seen in figure 5.

Fig. 5. Air cooler with fine tube bank configuration (a) and plain finned tube

configuration (b)

d. Mesh generation

At first, the meshing was executed by ANSYS Meshing on the

geometry designed with bare tube bank configuration with the

real dimensions of the air cooler object of the study (Figure 6

shows the entire domain). By using tetrahedral mesh and

refinement by the inflation, the meshing developed more than

14 million elements. To reduce the calculation time we scaled

the geometry by using a similar reduced, while maintaining

the same ratio of the geometrical parameters (length, width,

diameter of tube, pitch)

With the new geometry, we executed several simulations

which converge in 2671810 elements. In fact the pressure drop

physical surface temperature velocity pressure

Air inlet 480 K 5-10-15-20 m/s 𝜕𝑃

𝜕𝑥=0

Air outlet 𝜕𝑇

𝜕𝑥=0

𝜕𝑢

𝜕𝑥=0 _

Water inlet 290 k 0,6 𝜕𝑃

𝜕𝑦=0

Water outlet 𝜕𝑇

𝜕𝑦=0

𝜕𝑢(𝑤)

𝜕𝑦=0 atmospheric

Tube inner wall coupling - -

Tube outer wall coupling - -

Fin surface coupling - -

Symmetry 𝜕𝑇

𝜕𝑥 ,

𝜕𝑇

𝜕𝑦=0

𝜕𝑢

𝜕𝑥 ,

𝜕𝑢

𝜕𝑦=0

𝜕𝑃

𝜕𝑥 ,

𝜕𝑃

𝜕𝑦=0

Item Value

Nomber of tubes 24

Tube diameter (mm) 15

Length AIR COOLER (mm) 1600

Width AIR COOLER (mm) 400

Heigth AIR COOLER (mm) 1200

Material cooper

(a)

(b)



increases by 4% when moving from 1709709 to 26718109

elements and decreases by 11.6% when reducing the mesh size

from 3986256 to 26718109 elements, as shown in table 3a.

On the other hand, we modified the new geometry by adding

simple fins to conceive a plain finned tube configuration and

executed the same configuration of meshing which converged

in 3547628 elements. Table 3b shows the details of the

modification.

TABLE III

GRID INDEPENDENCE RESULTS FOR (a) BARE TUBE BANK CONFIGURATION (b) PLAIN FINNED TUBE CONFIGURATION

(a)

cell number ΔP(Pa)

784753 1278

1709709 1574

2671810 1640

3986256 1827

4424278 2456 (b)

cell number ΔP(Pa)

1124563 2045

3547628 2165

4256721 3537

4899123 5879

5638993 4389

Fig. 6. Meshed domain

e. Data reduction

Nusselt number:

The dimensionless Nusselt number is given by:

Nu=ℎ𝑑

𝜆

Where h is the convective heat transfer coefficient of the

flow, d is the diameter of tubes , λ is the thermal

conductivity of the fluid.

The correlation of Zukauskas defined Nusselt number by:

Nu= c Rem Prn (𝑃𝑟

𝑃𝑟𝑤)0,25

Where c and m in a definite region of Re, and n is relative to

the effect of fluid physical properties on heat transfer.

Reynolds number:

The Reynolds number for the tube side is determined by:

Re=𝑣𝑑

𝜗

Where ϑ kinematic viscosity.

Heat transfer rate

The coefficient of heat transfer for the bare tube bank heat

exchanger has the form: 𝐻 =𝑄

𝑆𝛥𝑇

For the plain finned tube heat exchanger, it is calculated with:

𝐻 =𝑄

𝜂0𝑆𝛥𝑇

Where Q the heat flux exchanged, ΔT is the logarithmic mean

of the temperature difference: ΔT=(𝑇𝑖𝑛𝑡−𝑇𝑤)(𝑇𝑜𝑢𝑡−𝑇𝑤)

𝑙𝑛(𝑇𝑖𝑛𝑡−𝑇𝑤)

(𝑇𝑜𝑢𝑡−𝑇𝑤)

S is the external surface tube

𝜂0 is the surface efficiency calculated from :

𝜂0 = 1 − (1 −𝑇𝑓 − 𝑇𝑎𝑣

𝑇𝑤 − 𝑇𝑎𝑣

)𝑆𝑓

𝑆

Tav is the average temperature of air fluid of outlet and inlet

temperature, Tf is the mean temperature of fin, Sf is the fin

surface area.

Pressure drop:

The pressure drop is determined by ΔP= Pint - Pout

Colburn factor: j=ℎ

𝜌CpUmPr2/3 (7)

Friction factor: f=2𝛥𝑃

𝜌𝑈𝑚2

𝐴𝑎

𝐴 (8)

III. RESULTS AND DISCUSSIONS

A. verification of the model

The base of comparison used for the validation of the present

CFD simulation are the model of Min [4] and correlations of

Zhukaukas [6]. The diameter tubes of the used geometry in

simulations (bare tube bank heat exchanger) is upper to that

adapted in the geometry of both references models but with

the same dimensionless ratio of the diameter and pitch.

The adapted model of turbulence is k-e realizable, it is the

same model chosen by Jingchunb after comparing it with other

turbulence models including k-ε (standard, SST, RNG) and

asserted that it is the closest model in that of the Zhukauskas

correlation’s.

The validation was made on one hand by the comparison of

the Nusselt number according to the Reynolds number of the

present simulation and the model of Min and the Zhukauskas

correlation as shown in figure 7, on the other hand, figure 8

presents the comparison of coefficient of heat transfer and the

pressure drop according to the speed with the same models.

It can be seen that the Nusselt number difference between the

studied model and the reference models does not exceed

14.3% compared to the Zhukauskas model at point Re =

14697, and reaches its maximum value 10.9 % compared to

the model of Min at point Re = 7944, for figure 02 and 03 the

deviation of H and ΔP reaches its maximum (11.4% and 9.3%

respectively) at the high speed V = 20 m/ s (Fig 9).

https://en.wikipedia.org/wiki/Dimensionless_number

https://en.wikipedia.org/wiki/Convective

https://en.wikipedia.org/wiki/Heat_transfer_coefficient

https://en.wikipedia.org/wiki/Thermal_conductivity

https://en.wikipedia.org/wiki/Thermal_conductivity



Fig. 7. Comparison of Nusselt number generation by Zhuskauskas

correlation, Jingchun Min model and present CFD model

Fig. 7. Comparison of heat transfer coefficient generation by Zhuskauskas

correlation, Jingchun Min model and present CFD model

Fig. 8. Comparison of pressure drop generation by Zhuskauskas correlation, Jingchun Min model and present CFD model

B. Temperature

Figure 9 and 10 show the distribution of the local temperature by

comparing the figures relating to the two configurations. It appears

very clearly that the variation of the temperature is late for bare tube

bank configuration compared to plain finned tube configuration and

this delay increases with velocity. A solution to recover this delay is

to increase the number of rows of tubes, so the plain finned tube

configuration optimizes the number of rows to belay the same

temperature distribution in the air cooler.

Fig. 9. The distribution of the local temperature in bare tube bank

configuration for 5, 10, 15 and 20m/s.

Fig. 10. The distribution of the local temperature plain finned tube

configuration for 5, 10, 15 and 20m/s.

C. Presssure drop

Apart from the pressure drop due to the friction with

tubes, other losses by the dissipation of the kinetic energy of the air

particles by friction with the walls of the fins explain why the plain

finned tube configuration presents a more drop pressure than the bare

tube bank configuration. Figure 11 shows the comparison of the

pressure profile in the first row of the two configurations studied. We

note that the pressure drop is excessive in the plain finned tube

configuration.

5m/s

20m/s

15m/s

5m/s 10m/s



Fig. 11. The distribution of the local pressure in bare tube bank

configuration(a) and plain finned configuration(b)

D. Heat transfer performance

Logarithmic mean of the temperature difference (LMTD)

The LMTD calculation reflects the heat exchange between hot air and

cooling water (inversely proportional to the heat exchange

coefficient). Figure 12 shows that LMTD evolves with the flow

velocity and it is more important in plain finned tube configuration

with an average difference of 10.6%.

Table IV and V include the values of the air and water inlet

temperatures for the two configurations derived from the simulation

models.

TABLE IV

THE VALUES OF THE AIR AND WATER INLET TEMPERATURES

FOR BARE TUBE BANK CONFIGURATION

TABLE V

THE VALUES OF THE AIR AND WATER INLET TEMPERATURES FOR PLAIN FINNED TUBE CONFIGURATION

Fig. 12. Logarithmic mean of the temperature difference

Colburn factor j:

Colburn factor j given by equation (7), highlights the relation

between heat transfer and the kinetics of fluid in flow, and as

it was intended for the plain finned tube configuration presents

a factor j more important compared to bare tube bank

configuration due to the fins that constitute a flow obstacle.

Figure 13 compares the colburn factor of the two

configurations. The maximum difference is 31% at low

velocities and at least 14.2% at high velocities. This indicates

better heat transfer capacity for plain finned tube

configuration.

Fig. 13. Colburn factor as a function of Re

FRICTION FACTOR

The numerical results of friction factor with different

Reynolds number are shown in figure 14. The analyze of this

results may be presented as follows: first, the friction factor

decreases greatly with Reynold number for plain finned tube

configuration and is almost unchanged for bare tube bank

configuration. Secondly, plain finned configuration has a

higher coefficient of friction than the other configuration with

a maximum difference of 33%.

Fig. 14. Friction factor as a function of Re

Plain finned tube configuration

Velocity m/s

AIR WATER

Temperature inlet(k)

Temperature outlet(k)



5 490 359 292 341

10 490 347 292 349

15 490 332 292 352

20 490 304 292 357

Fin tube bank configuration

Velocity (m/s)

AIR WATER





5 490 372 292 337

10 490 363 292 339

15 490 346 292 344

20 490 321 292 352

15m/s

20m/s



Performance of the air cooler :

The plain finned tube configuration provides a

considerable improvement in the coefficient of heat

exchange, but with important constraint of pressure drop

compared to the bare tube bank configuration, Figure 15

gives the graph of h / dP which allows to compare the

performance of the two configurations, and present an

maximum advantage for the plain finned tube

configuration by 9,7% for 5m/s.

Fig. 15. Performance comparison of bare tube bank and plain finned tube

configuration

IV. CONCLUSION

The CFD simulation of a supercharged air cooler of a marine

engine with these real dimensions and with its current bare

tube bank configuration then with new plain finned tube

configuration has made it possible to focus on the effect of fins

on the performance of the exchanger and compare numerically

the airside heat transfer and pressure drop characteristics the

both configurations. The new configuration provides a higher

heat exchange coefficient provoke by the excellent

temperature distribution throughout the rows of tubes and also

additional pressure drop relative to the current configuration

as constraint, this drop is caused by the friction of the particles

of air with the walls of the fins, in conclusion the air cooler

with plain finned configuration more efficient than the one

with bare tube bank configuration.

NOMENCLATURE

Cp Specific heat Capacity (J kg- 1 K-1)

g Gravitational constant (m/s2)

P Pressure (Pa)

t Time (s)

DTLM Average temperature difference Logarithmic.

h Heat transfer coefficient [w/k]

m Mass flow rate [kg/s]

Pk Producing term of turbulent kinetic energy

k Turbulent kinetic energy;

ε Turbulent dissipation rate

σε , σk Prandtl numbers

d Tube diameter (mm)

L Length heat exchanger (mm)

λ Thermal conductivity (wm-1k-1)

ρ Density (kgm-3)

ϑ kinematic viscosity (m2s-1)

A Heat transfer area (m2)

Aa Total air-side surface area (m2)

U Velocity (m s−1)

Δp Pressure drop (Pa)

η Fin efficiency

ηo Overall fin efficiency

REFERENCES [1] Arafat A. Bhuiyan, A.K.M. Sadrul Islam, Thermal and hydraulic

performance of finned-tube heat exchangers under different flow ranges, International Journal of Heat and Mass Transfer 101 (2016) 38–59

[2] L.Gu, J. Min, X.Wu, L.Yang, Airside heat transfer and pressure loss characteristics of bare and finned tube heat exchangers used for aero engine cooling considering variable air properties, International Journal of Heat and Mass Transfer 108 (2017) 1839–1849

[3] P. Salmon, L. Könözsy, C. Temple, S. Grove, Numerical Investigation on Various Heat Exchanger Performances to Determine an Optimum Configuration for Charge Air Cooler, Oil and Water Radiators in F1 Sidepods, Applied Thermal Engineering (2017)

[4] J.C. Min, R.L. Webb, Numerical predictions of wavy fin coil performance, J. Enhanced Heat Transfer 8 (3) (2001) 159–174.

[5] J.Y. Jang, M.C. Wu, W.J. Chang, Numerical and experimental studies of three dimensional plate-fin and tube heat exchangers, Int. J. Heat Mass Transfer 39 (14) (1996) 3057–3066.

[6] A. Zukauskas, R. Ulinskas, Efficiency parameters for heat transfer in tube banks, Heat Transfer Eng. 6 (1) (1985) 19–25.

[7] W.M. Kays, A.L. London, Compact Heat Exchangers, McGraw-Hill, New York, 1984.

[8] H.C. Kang, J.H. Eoh, J.E. Cha, et al., Numerical study on pressure drop and heat transfer for designing sodium-to-air heat exchanger tube banks on advanced sodium- cooled fast reactor, Nucl. Eng. Des. 254 (2013) 5–15.

[9] R.L. Puterbaugh, J. Brown, R. Battelle, Impact of heat exchanger location on engine performance, SAE Technical Paper, 2012.

[10] Z. Liu, H. Li, L. Shi, Y. Zhang, Numerical study of the air inlet angle influence on the air–side performance of plate-fin heat exchangers, Applied Thermal Engineering (2015),

[11] Claude Jean, B. Beaulieu, S. Bédard, M. Blaquière, L. Breton, B. Leclerc, Caractéristiques fonctionnelles des moteurs diesels marins, Institut Maritimr du Québéc .

[12] F. Inc., FLUENT User’s Guide,



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Cfd Simulations of Thermal Performance of Bare and Finned ...ijens.org/Vol_18_I_01/182201-7373-IJMME-IJENS.pdf · compounding and the complex supercharging system. Often, turbocharging