international journal of heat and mass transfer · management and the proposed procedure delivered...

International Journal of Heat and Mass Transfer xxx (2012) xxx–xxx

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International Journal of Heat and Mass Transfer

journal homepage: www.elsevier .com/locate / i jhmt

A review on air flow and coolant flow circuit in vehicles’ cooling system

S.C. Pang, M.A. Kalam ⇑, H.H. Masjuki, M.A. HazratDepartment of Mechanical Engineering, University of Malaya, 50603 Kuala Lumpur, Malaysia

a r t i c l e i n f o

Article history:Received 13 June 2012Received in revised form 25 June 2012Accepted 2 July 2012Available online xxxx

Keywords:Cooling systemNumerical modelVehicles’ keying-offAir flowCoolant flowAfter-boiling phenomenon

0017-9310/$ - see front matter � 2012 Elsevier Ltd. Ahttp://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.07

⇑ Corresponding author. Tel./fax: +60 3 79674448.E-mail addresses: [email protected] (S.C. Pan

Kalam), [email protected] (H.H. Masjuki), alihazrat

Please cite this article in press as: S.C. Pang et al.http://dx.doi.org/10.1016/j.ijheatmasstransfer.2

a b s t r a c t

Engine cooling system plays an important role to maintain the operating temperature of engine. The cool-ant circuit initiates by picking up heat at water jackets. With the pressure gradient exists in coolant circuit,hot coolant flows out from engine to radiator or to bypass circuit (during cold start). The under hood airflow carries heat away at radiator after the air flows through numerous hood components. The coolant flowcircuit and air flow circuit meet each other and exchange heat at radiator. Extensive researches are carriedout to study vehicles’ cooling system extensively either numerically or experimentally. The research coversmany individual topics which include numerical modelling of engine cooling system, under hood air flow,heat transfer at water jacket, heat transfer at radiator and coolants’ after-boiling phenomenon.

� 2012 Elsevier Ltd. All rights reserved.

Contents

1. Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 002. Coupling of 3D CFD simulation and 1D thermo-fluid simulation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 003. Under hood cooling air . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 004. Heat transfer at water jackets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 005. Heat transfer at radiator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 00

5.1. Overall heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 005.2. Air side calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 00

6. Transient state of cooling system after engine rapid shutdown . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 007. Conclusions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 00

Acknowledgement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 00References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 00

1. Introduction

Engine cooling system is crucial to ensure engine could operateat its optimum temperature. However, it is always being neglectedor sub- prioritized in vehicles design. For instance, radiator air flowis compromised when lower hood design is desirable for vehicles’cosmetic. The coolant circuit initiates by picking up heat at waterjackets. With the pressure gradient exists in coolant circuit, hotcoolant flows out from engine to radiator or to bypass circuit (dur-ing cold start). The ratio of coolant flow to radiator and to bypasscircuit is actuated thermally by thermostat. The under hood air

ll rights reserved..002

g), [email protected] ([email protected] (M.A. Hazrat).

, A review on air flow and coola012.07.002

flow carries heat away at radiator after the air flows throughnumerous hood components. The coolant flow circuit and air flowcircuit meet each other and exchange heat at radiator.

In numerical modelling, the complexity of hoods’ geometry en-quires computational fluid dynamic (CFD) to capture the dynamicair flow to radiator. While one dimensional thermo-fluid simula-tion could model coolant circuit in components level and couldsimulate the transient coolant temperature. Secondly, the coolingair flow is affected by numerous components at under hood.Numerical and experimental works are conducted to study thecooling air flow. Thirdly, heat transfer at water jacket which cool-ant absorbs and carries away the combustion heat is anotherimportant branch of engine cooling research. Fourthly, heat trans-fer at radiator is another major research category. The effect of finpitch, number of tube rows and fin topography could affect the

nt flow circuit in vehicles’ cooling system, Int. J. Heat Mass Transfer (2012),

http://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.07.002

mailto:[email protected]





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Nomenclature

q density [kg/m3]ot difference in time [s]ox, oy, oz difference in space [m]ou difference in velocity [m/s]oP difference in pressure [Pa]A, B, C, D constantAw cylinder wall area [m2]Ao prime area [m2]At tube area [m2]Ar, Ai radiator face area and inlet area [m2]Cp specific heat of coolant and air [kJ/kg K]f friction factorFp fin pitch [m]h fluid convective coefficient [W/m2 K]ha air side convective coefficient [W/m2 K]hc coolant side convective coefficient [W/m2 K]hg gas side convective coefficient [W/m2 K]Dhgrille total head loss across the grille [kg/ s2 m]j Colburn factorKgrille Grille coefficientKloss Grille loss coefficient

Kram Ram pressure coefficientkair thermal conductivity of air [W/ m K]kwall thermal conductivity of wall [W/ m K]DPgrille, DPunderbody pressure change at grille and under body [Pa]Pr Prandtl numberQ heat [kW]Re Reynolds numberRth total thermal resistance [K/W]Tfluid fluid film temperature [K]Ta air temperature [K]Tc coolant temperature [K]Tg gas temperature [K]Tw,a air side wall temperature [K]Tw,c coolant side wall temperature [K]Tw,g gas side wall temperature [K]U overall heat transfer coefficient [W/ m2 K]V1 Free stream air velocity [m/s]Vr air velocity at radiator face [m/s]x thickness of wall [m]

2 S.C. Pang et al. / International Journal of Heat and Mass Transfer xxx (2012) xxx–xxx

thermal performance of radiator. Finally, this research would liketo emphasize also the after-boiling phenomenon of coolant dueto heat soak (after rapid engine shut down). The heat soak resultedtremendous increment in coolant temperature after vehicles’ key-ing-off. The after-boiling-phenomenon creates fatal damages tocomponents in coolant circuit.

2. Coupling of 3D CFD simulation and 1D thermo-fluidsimulation

In the evaluation of the performance of vehicle heat exchangers,two main portions of job which are the study of cooling air aerody-namic and heat exchanger system analysis are required. Manyresearchers completed either one portion of the job and someresearchers established a coupling of both portions. Fig. 1 dis-played the basic components in an engine cooling system.

Since the cooling air flow will pass through the front bumper,grille and other heat exchangers (i.e. condenser, charged air cool-er), the velocity distribution at the face of radiator is highly non-uniform, especially in low speed driving. Experiment investigationwas carried out to determine which shielding methods at front end

Fig. 1. Basic elements in e

Please cite this article in press as: S.C. Pang et al., A review on air flow and coolahttp://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.07.002

would provide more uniform velocity distribution hence highercooling performance [1]. Extensive experimental-based researchwas also carried out to control car under hood thermal conditions.An innovative and accurate method to measure radiative and con-vective heat flux at the under hood was proposed in the research.Furthermore, optimized thermal management was suggested byproper positioning of components (upstream & downstream) andplacing of deflector [2,3].

A lot of numerical models are developed to predict the coolingair flow, either in 2D [4] or 3D. Besides, there are few off-the-shelfcommercial CFD software like Fluent, Vectis, StarCD and StarCCM+[5]. The fundamental concept of computational fluid dynamics waswell described in many sources [6]. CFD is basically a Navier–Stokes solver which solves the equations of conservation of massand momentum. In CFD, the partial differential equation (PDE)form momentum equation is discretized into algebraic equations,by either finite volume method or finite difference method. If thetemperature data is crucial, it is necessary to include the numericalmodels with energy equation as well. Below is the momentumequation in x direction. It shows a strong relationship betweenpressure and velocity.

ngine cooling system.



S.C. Pang et al. / International Journal of Heat and Mass Transfer xxx (2012) xxx–xxx 3

q@u@tþ u

@u@xþ v

@u@yþw

@u@z

� �¼ � @P

@x

þ l @2u@x2 þ

@2u@y2 þ

@2u@z2

!fluid acceleration in x direction

¼ body forceðpressureÞ þ surface faceðshear stressÞ ð1Þ

In order to capture the geometry effect, dynamic air flow atvehicle under hood was always modelled using CFD. In the litera-ture, steady state cooling air flow path with simplified under hoodgeometry to study velocity profile and temperature distributionwas modelled in StarCD using SIMPISO solver. The boundary con-ditions for two scenarios moving and stationary were experimen-tally determined; the boundary conditions are the surfacetemperatures of heat sources (i.e. engine, reservoir, gearbox, ex-haust manifold and heat exchanger). The numerical models werevalidated with experiment at wind tunnel. The other componentsin the under hood were assumed to be adiabatic [7]. A CFD modelwas established using Fluent to minimize air recirculation insidethe radiator cover thus maximizes the flow through the radiatorcore and improves radiator efficiency. He used ANSA and TGRIDfor the surface and volume meshing, respectively. The radiatorwas modelled as a porous medium to account for the pressure dropinside the radiator [8]. The velocity over the radiator and air-to-boil temperature were calculated by CFD using PASSAGE [9].

For other applications, cooling fan and heat shield were designedand distance between fan and heat exchanger was optimized for aheavy duty truck using CFD [10]. Several geometry modificationswere performed in StarCD CFD model to improve air flow approach-ing intercooler [11]. On the other hand, a combination of CFD andFlow Network Modeling instead of single CFD model was being uti-lized to model under hood air flow path [12]. Meaningful investiga-tion on a computational procedure was carried out and wasimplemented in Vectis. The research studied under hood thermalmanagement and the proposed procedure delivered accurate tran-sient predictions with substantially reduced computational time[13,14]. Some researchers used StarCCM+ as CFD platform to simu-late ram air pressure, grille, heat exchangers, cooling fan, shroud,engine bay restriction and positioning of outlet. They emphasizedon three main aspects which are porous media modelling of heatexchanger, fan modelling and system modelling [15].

In order to save experiment time and prototype cost, CFD is al-ways the primary choice to test any creative and innovative auto-motive front-end designs. In another research, a total of 22 fasciadesigns were simulated in the CFD software (Fluent). The CFD pre-dictions were compared with experimental measurements of radi-ator specific dissipation [16]. Some researchers tested CFD

Fig. 2. Coupling of 3D CFD model and 1D thermo- fluid model.


applications for automotive front-end design. They simulated oneCFD model to reach three ultimate goals which were power traincooling, vehicles’ aerodynamics and climate control [17].

In several literatures, one dimensional thermo-fluid simulationis a supreme option to reflect the system effect of coolant path.Some researchers modelled both air flow and coolant flow basedon one dimensional thermo-fluid simulation, with the softwarenamed KULI. The flow models are based on network flow theory,which permits the construction of complex cooling circuits [18].Some research utilized AMESim to perform system analysis for en-gine cooling system, mainly coolant flow path [19].

The methodology to integrate the 3D CFD model (i.e. Fluent,StarCD, Vectis, StarCCM+, PowerFLOW and Cool3D) with 1D ther-mo-fluid system model (i.e. KULI, Flowmaster, Dymola, GT-cool) isfairly interesting. 3D CFD model could reflect the non-uniformvelocity distribution of cooling air at the frontal surface of heatexchangers (i.e. condenser or radiator). The 1D thermo-fluid systemmodel is utilized as a system performance computation for enginecooling system or vehicle HVAC system. Basically, it involves a pro-cess which the models exchange boundary conditions until theyconverge with each other. 3D CFD model will feed the 1D thermo-fluid model with boundary conditions, such as convection coeffi-cient, velocity and fluid temperature. On the other hand, 1D ther-mo-fluid model will feed CFD model with boundary conditions,such as heat rejection to air. Fig. 2 elaborated how the integrationof two numerical models is and how the whole model convergedby exchanging boundary conditions. A procedure was discussedregarding integration of a CFD model using Vectis (of air flow) anda 1D thermal-fluid model (of coolant flow and oil flow) for vehiclethermal management study [20]. The effect of under hood recircu-lation flows on AC system performance was evaluated by couplingPowerFLOW and Dymola [21]. On the other hand, CFD model wasbuilt using Cool3D and thermo-fluid model was built using GT-cool.Then the integrated models were exploited to study engine coolingsystem [22]. For cooling air flow at condenser, industry based ex-perts integrated the model of StarCD and Flowmaster [23] for pas-senger cabin cooling analysis [24]. Last but not the least,integrated model of CFD model using Fluent (under hood air flow)and thermo-fluid model using Flowmaster (lubrication and coolantcircuit) was established to study under hood thermal analysis [25].Researchers also integrated model of Fluent, StarCD and KULI forautomotive air conditioning analysis including air flow approachingcondenser [26]. A procedure was also described to integrate CFDmodel with KULI. The researcher emphasized that 1D thermo-hydraulic model gives the possibility to analyze split flow patternsor non-uniform air velocity distribution. In the 1D model (KULI),the heat exchanger will be modelled in such a way that it is parti-tioned into a number of rectangular segments. In each segment, afictive flow resistance will be determined according to CFD distribu-tion [27]. All the formers illustrated above are 3D CFD models whilethe latters are 1D thermo-fluid system models. After the coupling ofmodels, a fruitful and complete cooling/ air-conditioning systemanalysis could be performed with different scenarios. It is worthmentioning that it is not necessary to run 3D simulation for thewhole system. By coupling the 1D and 3D models, it is the bestmethod to yield the advantages from both 1D and 3D modeling.

In some cases, researchers also included a finite-element-anal-ysis thermal model/temperature model (besides CFD and ther-mo-fluid models) to consider all the three modes of heat transfer,which are conduction, radiation and convection. This occurs whenthe impacts of conduction and radiation are significant towards thewhole system. CFD only accounts for convective heat transfer influids. Thermal model will provide wall temperature of compo-nents as boundary conditions for the respective cooling air CFDmodel. In return, CFD model would give convective coefficient(h) and fluid film temperature (Tfluid) as boundary conditions for



Fig. 3. Schematic configuration of cooling module and air flow pattern in vehicleengine room [26].

Fig. 4. The pressure jump and pressure drop of cooling air in engine compartment[26].


the thermal model since convective heat (h) is the common outputfrom both models. The value difference of convective heat fromCFD model and thermal model will decide the convergence of thecoupled models. By exchanging boundary conditions betweenmodels, it could ensure higher accuracy solution from both models.In under hood thermal analysis, the coupling of thermal modelusing PERMAS and POSRAD with the CFD model using StarCDwas illustrated. In the models, six units of heat exchangers werebeing considered [28].

Besides coupling cooling air CFD model with 1D system model,researchers developed a second CFD model for air flow inside thepassenger cabin. In order to include the effect of radiation and con-duction, the second CFD model was coupled with a thermal modelusing THESEUS. Initial boundary condition for the thermal modelwas obtained from experiment, which is a wind tunnel test. In short,two CFD models, one thermal model and one thermo-fluid systemmodels were developed to solve a single problem [24]. In evaluationof AC performance of combat vehicle, researchers also coupled boththe volume based CFD model (under hood cooling air and passengercabin) with a shell based thermal model using MuSES. A detailedcoupling procedure was elaborated by exchanging boundary condi-tions [29]. For analysis of air flow in a building, researchers illus-trated the coupling of energy system model and CFD model [30].

3. Under hood cooling air

There are two main sources of energy which contribute to thecooling air flow through under hood, one is the ram air and anotheris radiator fan. For vehicle at high speeds, the main driving force forthe cooling air flow is ram air. For vehicle at low speeds, the maindriving force for the cooling air is radiator fan. Ram air is a flowdriving force resulted from favorable static pressure gradient be-tween the vehicle frontal opening inlets and underbody. As the freestream air approaching the frontal opening, the air velocity or dy-namic pressure is reduced while static pressure increases, in orderto maintain same total pressure (Bernoulli Law). The static pres-sure is highest at a stagnant point before frontal opening (fluidvelocity is zero). On the other hand, the air acceleration as a resultof Venturi Effect (lower cross sectional area, higher velocity) cre-ates a low static pressure area at the underbody. Meanwhile, elec-trical fan, acting as a momentum source, incurs a static pressurejump and total pressure jump for the air which flows throughthe fan. Fans do work on the cooling air by giving the air both staticand dynamic energy. Roseberry [31] emphasized the importance ofinteraction between external air flow and internal air flow. Hecommented that two interfaces of this interaction are the frontalopening and undersize of engine bay. He defined a grille coefficient(Kgrille) as the fraction of the free stream total pressure which isdelivered through the grille, Equation Set (2). Grille loss coefficient(Kloss) is obtained by normalizing total pressure difference acrossthe grille with free stream dynamic pressure. Experimental meth-od to study the relationship between total pressure loss acrossgrille versus air flow rate, vehicle speed and layout of frontal open-ings was used. Alternatively, the grille coefficient may be ex-pressed as a function of the ratio of an average velocity at theface of the radiator to the free stream velocity and the ratio ofthe radiator face area to the inlet area [31].

Dhgrille ¼ p1 þ12qV2

1

� �� pr þ

12qV2

r

� �

K loss ¼Dhgrille

1=2qV1Kgrille ¼ 1� K loss

Kgrille ¼ Kg0 þ Kg1Ar

Ai

� �Vr

V1

� �ð2Þ


Another researcher defined a ram pressure coefficient (Kram) asram pressure normalized by the average dynamic pressure at radi-ator, Equation Set (3). It is important to note that static pressuredifference was used instead of total pressure difference in his for-mulation. This is due to the assumption that average air velocityfor internal flow is similar, thus total pressure difference is equiv-alent to static pressure difference [32].

DPgrilleþDPcondenserþDPradiatorþDPfanþDPengine bayþDPunderbody¼0DPram¼�ðDPgrilleþDPunderbodyþDPenginebayÞ¼DPcondenserþDPradiatorþDPfan

Kram¼DPram

1=2qV2r

Kram¼V1Vr

� �2

K0þðAr

AiÞKA�Kbay

ð3Þ

Besides, several researches provided significant insights andfundamental exposure regarding vehicle cooling air flow [33–35].Finally, Fig. 3 exhibited the main components at vehicles’ hoodwhich could affect air flow pattern. Fig. 4 displayed the pressuredevelopment as air flows through the hood [26]. Fig. 5 showedthe pressure definition at each hood components which ease theequation derivation later [33].

4. Heat transfer at water jackets

In 2006, a research on parametric study for heat transfer atwater jackets was completed [36]. The gas side convective heattransfer coefficient (hg) was calculated using Annand and Woschini



Fig. 5. Full scale vehicle cooling air flow stream tube [33].

Fig. 6. Schematic of thermal resistance concept at water jackets.

Table 1Trend of gas pressure, gas temperature, gas side convective coefficient and coolantside convective coefficient in response to engine speed, compression ratio and etc.

Pg Tg hg hc

Engine speed " " (slight) " " "Compression ratio " " ; " "Excess air ratio " ; ; ; ;Inlet pressure, Pi " " " (slight) " "Inlet temperature, Ti " ; " (slight) ; ;

Fig. 7. Schematic of overall heat transfer coefficient and thermal resistance conceptat radiator.

Fig. 8. Sketch illustrating the variation in heat transfer coefficient along the length


equation, Eq. (4). The coolant side convective coefficient (hc) wascalculated with the measurement of gas temperature (Tg), gas sidewall temperature (Tw,g), coolant side wall temperature (Tw,c) andcoolant temperature (Tc), Eq. (5). The gas side coefficient increaseswith the increase of in-cylinder pressure, increase of gas velocityand decrease of in-cylinder temperature. Despite the increase ofgas side temperature (adverse effect toward coefficient), increaseof gas side pressure poses dominant effect towards gas side heattransfer coefficient. It is shown that the engine speed and compres-sion ratio impact more significantly than other engines parame-ters. Meanwhile, the effect of engine speed and excess airtowards coolant side convective coefficient are more severe. Itwas also concluded that very high coolant side coefficient is evi-dence of nucleate boiling [36]. At 1000 rpm, gas side convectioncoefficient is 1097 W/m2 K while coolant side convection coeffi-cient is 4263 W/m2 K. As thermal resistance is an inverse of con-vection coefficient, it could be observed that gas side thermalresistance conquer significant percentage of the total thermalresistance, Fig. 6. Fig 6 illustrated the heat flow path across enginewall and its respective thermal resistance. For same heat flow,higher thermal resistance will incur higher temperature gradient.Table 1 displayed the trend of gas pressure, gas temperature, gas


side convective coefficient and coolant side convective coefficientin response to engine speed, compression ratio and etc.

hgðhÞ ¼ 3:26X10�3D�0:2pðhÞ0:8wðhÞ0:8TðhÞ�0:55 ð4Þ

�hc ¼�hgð�Tg � �Tw;gÞð�Tw;c � �TcÞ

ð5Þ

5. Heat transfer at radiator

5.1. Overall heat transfer coefficient

In computation of radiator overall heat transfer coefficient (U), anumerical study was developed for various design parameters ofheat exchanger. Some researchers utilized thermal resistance con-cept for the heat transfer at radiator in their modelling (Fig. 7) [37].

of a continuous fin.




Besides, some researchers mentioned that in the applications ofthe fin-and-tube heat exchangers, the air side thermal resistancegenerally comprises over 90% of the total thermal resistance [38].In this connection, significant efforts were made to improve theairside performance. In order to reduce the airside thermal resis-tance, the airside convection heat transfer coefficient, overall area(Ao = prime surface + finned surface) and fin efficiency could be in-creased. Prime surface is the basic surface which allows heat flowsfrom heat source to heat sink. Finned surface is an extended sur-face of prime surface to enhance heat transfer. Fin efficiency is de-fined as a ratio of actual heat dissipation of a fin to its idealdissipation if the entire fin is of the same temperature as its base.Eq. (6) indicated the total thermal resistance for the heat flow pathacross radiator tube and fin. Eq. (7) computed the quantitative heatflow across radiator tube and fin.

Rth ¼1

UAt¼ 1

hcAt|ffl{zffl}coolant side

þ xkAt|{z}tube

þ 1hagAo|fflffl{zfflffl}air side

ð6Þ

Q ¼ hcAtDT1 ¼kAt

xDT2 ¼ hagAoDT3 ð7Þ

Fig. 10. Effect of number of tube rows on heat transfer and friction characteristics(Fp=1.23 mm) [39].
5.2. Air side calculations
In a research, the history developments of fin patterns for fin-and-tube heat exchanger were summarized [38]. First generationpattern was plain and wavy fin geometry. Second generation hadinterrupted surfaces features which enhanced heat transfer mech-anism like boundary restarting, wake management and flow desta-bilization. In Fig. 8, it shows that heat transfer coefficient is alwayshighest at the leading edge of a plate when the flow is laminar andwith thinner boundary layer. Sometimes, this phenomenon iscalled as entrance effect. This explains why boundary restartingcould improve heat transfer coefficient. Though, interrupted sur-face can significantly improve the heat transfer performance, theassociated penalty of the pressure drop is tremendous. A properre-design of heat exchanger should maximize the heat transferperformance with minimum penalty of pressure drop. The thirdgeneration with the enhanced surfaces employ longitudinal vortex

Fig. 9. Effect of fin pitch (Fp) on heat transfer and friction characteristics havingtube rows N=1 and N=2 [39].

Fig. 11. Schematics illustration of 3D computational results for the effect of finpitch [40].


generation that provided swirling motion to the flow field. Usuallythe swirling motions may be classified as transverse and longitudi-nal vortices.

The basic surface characteristics of heat exchangers are gener-ally presented in dimensionless form as friction factor f and theColburn j factor. The characteristic length for the Reynolds numberis fin collar outside diameters. Listed below are the characteriza-tion equations for the air side heat transfer and friction.

j ¼ ha

qVmaxCp;aPr2=3 ð8Þ



Fig. 12. Flow pattern at different Reynolds number [43].


j ¼ A

ReBDc

ð9Þ

f ¼ C

ReDDc

ð10Þ

Researcher also examined the effects of fin pitch, number of tuberows and tube diameters towards the thermal hydraulic character-istics for a plain fin-and-tube heat exchanger as shown in Fig. 9 and10 [39]. The researcher concluded that for ReDc < 5000 and low tuberows (N = 1, 2), heat transfer performance increases with thedecreasing of fin pitch. The researcher explained this phenomenonwith the numerical results of the effect of fin pitch carried out[40]. The investigation, displayed in Fig. 11(a) and (b) below, shownthat the vortex forms behind the tube can be suppressed and theentire flow region can be kept steady and laminar when the finpitch is small enough. For numerical results of two rows configura-tion, the first row cylinder is stabilized due to the existence of thesecond tube row, Fig. 11(c) [40]. The effect of fin pitch is negligibleat high Reynolds number, as the effect of vortex shedding gain inimportance.

Researchers displayed a result that for small fin pitch and lowReDc, the heat transfer performance decrease with the increase of

Fig. 13. Nu over the fin surface for S = Fp/Dt


number of tube rows, as shown in Fig. 10 [39] [41] [42]. The resultscan be interpreted from the observations by Mochizuki [43] whoreported steady laminar flow patterns prevailed throughout theradiator core at low Reynolds number. As the Reynolds number in-creased, the vortex shedding site would migrate upstream, asshown in Fig. 12. However, the transition from one flow regimeto the succeeding one was so gradual that a distinct transition Rey-nolds number could be determined. At very large Reynolds num-ber, the flow was turbulent for the whole radiator core. In short,the effect of tube rows was more pronounced in laminar flows(where boundary development is important) and diminished inturbulent flow.

However, some researchers studies the effect of fin spacing onconvection in a plate fin-and-tube heat exchanger by normalizedfin pitch Fp with tube diameter Dt [44]. In this study, the convectiveheat transfer coefficient increases with fin pitch (if tube diameter isconstant). This seemed to contradict (in fact not) with Wang [39],because this experiment was carried out by varying the distance oftwo plates. Higher fin pitch means higher mass flow thus betterheat transfer coefficient as shown in Fig. 13.

A researcher completed an experiment using an infrared ther-mo-vision to monitor temperature distribution over a plate fin sur-face inside the plate finned-tube heat exchangers [45]. The localconvective heat transfer coefficient is derivation of the differentia-tion of the temperature functions. In-line arrangement and stag-gered arrangement are compared in term of temperaturedistribution and local convective coefficient. Staggered arrange-ment shows a lower temperature distribution thus higher convec-tive heat transfer coefficient (Figs. 14 and 15). The highestconvective heat transfer coefficient was indicated at the leadingedge of the plate fin because the velocity boundary layer is initiallydeveloped in the z-direction. A most pronounced is that the tem-perature gradient on the fin surface is sharper near the front andsides of first two row tubes due to repeated growth and destruc-tion of the boundary layer by tubes. At the rear of the tube, thetemperature gradient is gentler because the airflow is swept down-stream into the wake. An exit effect is observed at the third tuberows. In Fig. 16, it shown that convective heat transfer coefficientis higher with increase fluid velocity (mass flow). Increase convec-tive heat transfer coefficient show the entrance effect at the firstrow and exit effect at third tube rows.

6. Transient state of cooling system after engine rapidshutdown

Several researchers established an experiment to study theafter-boiling phenomenon in small internal combustion engine[46–48]. In the experiments, the small engine was firstly being

= 0.116 and S = 0.365 at Re = 630 [44].



Fig. 14. Temperature map on the plate fin surface for both in- line and staggered arrangement, Fp = 10 mm, U = 1m/s [45].

Fig. 15. Convection heat coefficient on the plate fin surface for both in-line and staggered arrangement, Fp = 10 mm, U = 1 m/s [45].


operated at full load, and then it was kept idle for a period beforekeying-off. The result analysis of a research is summarized in Table2, Figs. 17 and 18 [48]. It is elaborated that during idle time whenthe pump is working at slower speed, the longer the coolant resi-dence time in radiator and coolant inlet temperature (into engine)decreases. On the other hand, the longer the coolant residence timein engine, the coolant outlet temperature (from engine) increases.The researchers divided the after-boiling phenomenon into threephases, which are air compression (b to d1), air leakage (d1 to d2)and coolant leakage (f1 to f2). In the air compression phase, thermalexpansion and vapor generation increases the system pressure and


compresses the air and air pockets inside the system. The boilinginside the coolant circuit is suppressed with the presence of pres-sure cap and expansion tank. This is attributed to saturation tem-perature increases with the increment of system pressure. Thesystem pressure increases until the pressure limit of the pressurecap, and the air and coolant start to leak out (during air and coolantleakage phases). The leaked volume of coolant, thermal contractionof coolant and condensation of vapor at radiator slowly bring thesystem pressure below the pressure limit of pressure cap. Then,the pressure cap closes and coolant cools as stagnant liquid. It isthought-provoking, when the mass diffusion theory was utilized



Fig. 16. The variation of the span- averaged local convective coefficients with Fp = 20 mm for the in-line array [45].

Table 2Summary of the after- boiling phenomenon [48].

Observation Trends Remarks

1. Idle (a) Coolantpressure

Decreased – Due to lower pump speed

(b) Coolanttemperature

Increased – Longer residence time with cylinder head due to small mass flow through engine

(c) Cylinder headtemperature

Decreasedby �60 �C

– Lower loading, heat which generated by engine combustion decreases

– Temperature gradient exists between cylinder head and coolant, heat transfer happens in themode of convection– The cylinder wall temperature (130 �C) is greater than coolant saturation temperature at thesystem pressure. Thus, nucleate boiling at cylinder head further boosts the heat transfer fromcylinder

2. Phase A (a) Coolantpressure

Increased – Due to fraction of coolant vaporizes at engine outlet

Air compression (immediatelyafter keying off)


Increased – Pressure increases at cylinder head, resulted a pressure gradient between cylinder head andradiator, coolant flows out from cylinder to radiator



– Due to evaporation of coolant (latent heat is used for phase change) and convection (sensibleheat is used for increment of coolant temperature) transfers the heat away from cylinder

(d) Others – Air in expansion tank and air pockets in cooling circuits are compressed(e) Cylinderbodytemperature

Increased – Conductive heat transfer from other warmer parts of cylinder body

3. Phase B (a) Coolantpressure

Increased – Due to fraction of coolant vaporizes at engine outlet

Air leakage (after opening ofpressure cap atpressure = 2 bar)


Increased – Accumulated heat is transferred away from cylinder head to coolant. Some energy is used forcoolant evaporation, some energy are stored as internal energy for coolant

– Coolant flows out from cylinder head, as pressure gradient exists between cylinder head andradiator



– Heat is transferred away by sub-cooled boiling and convection

(d) Others – At the end of this phase, expansion tank is full with liquid

4. Phase C (a) Coolantpressure

Increased(f1 –max)

– Vapor production is still the dominant factor

Coolant leakage Decreased(max – fv)

– Volume reduction due to coolants’ leakage predominant over the effect of vapor production

Decreased(fv – f2)

– Condensation of vaporized mass occurs at radiator. Finally, pressure cap closes and leakagestops


Remainedthe same

– Diminished temperature gradient between cylinder head and coolant, thus no existence ofheat transfer

c) Cylinder HeadTemperature

Remainedthe same

– Diminished temperature gradient between cylinder head and coolant, thus no existence ofheat transfer

5. After Phase C (a) Coolanttemperature

Decreased – Cools as stagnant vapor and stagnant liquid

Coolant temperature is measured at engine outlet.


Please cite this article in press as: S.C. Pang et al., A review on air flow and coolant flow circuit in vehicles’ cooling system, Int. J. Heat Mass Transfer (2012),http://dx.doi.org/10.1016/j.ijheatmasstransfer.2012.07.002


Fig. 17. Coolant temperature after engine rapid shutdown: (a) temperature at engine outlet and pressure of coolant; (b) temperature of cylinder head and cylinder body [48].

Fig. 18. Schematic explanation of mass diffusion theory [48].

Fig. 19. Coolant temperature at engine inlet and engine outlet [46].


to explain the migration of vapor from engine to radiator [48]. Highconcentration of vaporized mass at cylinder head is the drivingforce for vaporized mass to flow from cylinder head to radiator.

Researcher explained that the temperature increases at engineoutlet, as coolant flows out due to existence of pressure gradient


between cylinder head and radiator. Next, the researcher describedthat the temperature increases at engine inlet two minutes afterengines’ keying-off, as there is a reverse flow from cylinder bodyto radiator due to existence of pressure gradient between them(Fig. 19). Yet, most of the time, the coolant at engine inlets coolsas stagnant liquid and the coolant at engine outlet continues itsflow [46].

Furthermore, a researcher investigated the effects of idle time,initial load and piping length towards after-boiling phenomenon[47]. The investigation results showed that as the engines’ shutdown and coolant flow stops, the cylinder-head metal is hot en-ough to vaporize a fraction of coolant. As a result, the system pres-sure increases, while the air in expansion tank and air pocketsinside coolants are being compressed until a limit where the pres-sure cap is opened, a significant volume of air and coolant leak out.As the idle time is prolonged, cylinder head is given longer time tocool down to lower temperature. For the case of idle time = 80 s,mild boiling and convection heat transfer occurs after the vehicles’shut-down, and increases the coolant temperature and pressure.With the presence of pressure cap, the boiling is suppressed aspressure increases, coolant saturation temperature also increases.




The cooling system solely undergoes air compression stage. Noleakage is observed as limited temperature gradient (thus limitedheat transfer) exists between cylinder head and coolant. Withoutheat transfer, coolant temperature could no longer be increasedto saturation temperature corresponding to the pressure limit ofpressure cap. In short, for prolonged idle time, the volume ofleaked coolant decreases and the starting of leakage is delayed.On the other hand, in order to observe the effect of engine loadingtowards behavior of after-boiling phenomenon, the researcher var-ied the brake mean effective pressure (bmep) in the range of 495–819 kPa. For the case of 495 kPa, only phase A (air compression)and phase B (air leakage) are observed. In conclusion, with lowerbmep, the volume of leaked coolant is reduced and the startingof leakage is delayed. Finally, the researcher repeated three scenar-ios for pipe length of 40 cm, 85 cm and 190 cm. Significant impactsare observed especially in phase C (coolant leakage), volume ofleaked coolant and leakage duration are reduced with shorter pipe.Longer pipe will extend the vapor arrival time at radiator (fv), thusdelays vapor condensation time, pressure reduction and closure ofpressure cap (f2). For pipe 190 cm, the duration of phase A and B islonger as more volume for the thermal expansion and air compres-sion. The vaporized mass of coolant does not reach the radiator, butcools as stagnant vapor. Finally, pressure cap is closed due to thepressure reduction attributed to coolants’ leakage, but it is notattributed to vapor condensation.

7. Conclusions

In numerical modelling, integration of several simulations is re-quired to represent the whole vehicles’ cooling system. CFD simu-lation is always a better option to model under hood air flow.Thermo-fluid simulation is always chosen by researchers to modelthe coolant circuit. The existence of pressure gradient provokes theair flow at under hood. Researchers derived grille coefficient bynormalization of total pressure at hood. Research regarding heattransfer at water jacket and radiator could provide some insightabout how to improve the whole cooling system. Researcherscould find out which is the components with higher thermal resis-tance and could eliminate the obstacles for heat transfer. The studyin after-boiling-phenomenon helps to highlight among the crucialproblem for engine cooling system.

Acknowledgement

We would like to express our gratitude to the Ministry of Sci-ence, Technology & Innovation (MOSTI- TF001/2010) of Malaysiaand University Malaya Institute of Research Management & Con-sultancy (IPPP- PV007-2011A) for providing us with the researchfunding.

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http://dx.doi.org/10.1115/1.3126262


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