fan air flow analysis and heat transfer enhancement of vehicle underhood cooling system – towards...

Applied Energy 91 (2012) 439–450

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Applied Energy

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Fan air flow analysis and heat transfer enhancement of vehicle underhoodcooling system – Towards a new control approach for fuel consumption reduction

Mahmoud Khaled a,b,c, Fareed Mangi a, Hisham El Hage c, Fabien Harambat b, Hassan Peerhossaini a,⇑a Thermofluids, Complex Flows and Energy Research Group, Laboratoire de Thermocinétique, CNRS-UMR 6607, Nantes University, rue C.Pauc, BP 50609, 44 306 Nantes cedex 3, Franceb PSA Peugeot Citroën, Velizy A Center, 2 route de Gisy, 78 943 Vélizy Villacoublay, Francec Fluid Mechanics, Heat and Thermodynamics Group, School of Engineering, Lebanese International University, Beirut, Lebanon

a r t i c l e i n f o

Article history:Received 12 August 2011Received in revised form 9 October 2011Accepted 12 October 2011Available online 9 November 2011

Keywords:Vehicle underhoodCar cooling moduleHeat exchangersHeat transfer enhancementFanCar underhood aerothermal management

0306-2619/$ - see front matter � 2011 Elsevier Ltd. Adoi:10.1016/j.apenergy.2011.10.017

⇑ Corresponding author. Tel.: +33 2 40 68 31 24; faE-mail address: hassan.peerhossaini@univ-nantes.

a b s t r a c t

We report here experimental results focused on the optimization of vehicle underhood cooling module.These results constitute the basis for a new approach of controlling the cooling module positioningaccording to the engine energy requirements. Measurements are carried out on a simplified vehicle bodydesigned based on the real vehicle front block. We report here velocity and temperature measurementsby Particle Image Velocimetry (PIV), by Laser Doppler Velocimetry (LDV) and by thermocouples. The und-erhood of the simplified body is instrumented by 59 surface and fluid thermocouples. Measurements arecarried out for conditions simulating both the slowdown and the thermal soak phases with the fan inoperation. Different fan rotational speeds, radiator water flow and underhood geometries have beenexperimented. The ultimate aim is to apply the new control approach to a real vehicle so as to reducethe energy delivered to the pump and compressor and therefore to reduce the vehicle fuel consumption.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

The cooling module of vehicle constituted of an assembly of dif-ferent heat exchangers and one or two fans have been pushedbackward into the engine compartment, in order to evacuate heatin difficult operating conditions. This have caused the underhoodaerothermal and aeraulic conditions to be largely conditioned bythe different elements of the cooling module which at the sametime interact with other components in the underhood.

In heat exchangers’ applications which cover a large variety oftechnological domains [1–10] (such as nuclear reactors, chemicaland biomedical processes, military and aerospace, and automotive),the trend is to reduce the heat exchangers’ weight and volume whilekeeping sufficient or even increasing thermal efficiencies. In theopen literature, the large part of optimization studies is focused onthe geometry and the design of the heat exchangers and also onthe fluid flows structures. Several design and control approachesaiming to optimize the heat exchangers’ thermal performance arereported in [11–20], while few studies are focussed on the heatexchangers’ optimization taking into account their interaction withthe geometry in which they are installed.

Among the most commonly heat exchangers used are the tubesand fins heat exchangers which are composed of several tubes ofelliptical cross section between which are positioned several

ll rights reserved.

x: +33 2 40 68 31 41.fr (H. Peerhossaini).

continuous and parallel fins. These types of heat exchangers areparticularly adapted to most industrial applications, especially inautomotive applications, and their fins can be simple, louvered,crimped, shutters, or with delta wings longitudinal vortex genera-tors, etc.

Several studies [21–26] on the louvered fin heat exchangers forunderhood applications showed that the thermal performance ofsuch exchangers are strongly affected by the geometrical parame-ters of the heat exchanger (e.g. distance between fins, the tubesheight, and fins angle) and operating conditions of the two fluids,air and water (their mass flow rates and inlet temperatures). How-ever, few studies have explicitly addressed the crucial question ofhow to control the different parameters for enhancing the exchan-ger’s heat evacuation capacity. Most of the underhood heat transferanalysis has been concentrated on the underhood thermal manage-ment by temperature and heat flux measurements [27–34]. In thepresent study we have experimentally investigated the heat exchan-ger optimization by acting on the geometry (underhood space here)in which they are integrated. The present analysis shows how to acton the different parameters which influence the heat exchanger per-formance in order to optimize its functioning.

On the other hand, in critical phases of vehicle functioning suchas cooling at the vehicle stopped situation, the stop and go drivingmode, and the vehicle thermal soak, the performance of car heat ex-changer is essentially influenced by the air flow generated by the fan.This air flow in its turn depends on the underhood architecture. Mostexperimental studies on fan airflow are carried out for

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Fig. 1. (a) Underhood simplified model (b) real cooling module and simplified engine block (c) X–Y displacement rail system (d) parametric air inlet.

Fig. 2. LDV measurements carried out on the zone between the fan and the engine block.

440 M. Khaled et al. / Applied Energy 91 (2012) 439–450

determination of the fan aerothermal characteristics (curves of pres-sure rise as a function of the air flow) or optimization studies (use ofrectifiers in order to guide the airflow) where the fan alone is consid-ered. On real cars, these studies are quite limited given the con-straints of accessibility related to the presence of engine blockdownstream of the fan. For instance, it is very difficult to use opticalmeasurement techniques such as PIV or LDV in the car underhoodspace. One solution is to work on underhood simplified models de-signed to allow the integration of optical measurement techniques.Compared to studies on isolated fan, these models present theadvantage of reproducing the effects of the engine blocking imposedon the fan flow. This ‘‘blocking’’ due to the short distance betweenthe fan and the engine block downstream, is a parameter that couldaffect the characteristics of the flow induced by the fan in the carunderhood. Moreover, the space between the cooling module andthe engine block varies from one vehicle type to another. For these

reasons, it was decided in the present study to work on a car under-hood model containing an actual cooling module and simplified en-gine block both translatable in X and Y (respectively in the vehiclelength and width) directions. This modularity allows covering awide range of underhood architectures to explore different strate-gies adapted by the automobile manufacturers. Part of this paperwill thus be focussed on the PIV and LDV measurements as well asthermal measurements performed on the model for different block-ages downstream of the fan, that is to say, for different positions ofthe engine block with respect to the cooling system. The impact ofthese aerodynamic changes on the thermal performance of the heatexchangers and their surrounding area is also studied.

We report here velocity and temperature measurements by Par-ticle Image Velocimetry PIV, by Laser Doppler Velocimetry and bythermocouples. The underhood of the simplified car body is instru-mented by 59 surface and fluid thermocouples. Measurements are

Fig. 3. (a) PIV windows for the main flow (b) PIV windows for secondary flow (c) laser source in the engine block (d) camera at the simplified model side.

M. Khaled et al. / Applied Energy 91 (2012) 439–450 441

carried out for conditions simulating both the slowdown and thethermal soak phases with the fan in operation. Different fan rota-tional speeds, radiator water flow and underhood geometries havebeen tested. The final aim is to apply the new control approachdeveloped here on a real vehicle so as to reduce the energy deliv-ered to the pump and compressor functions and thus to reduce thevehicle fuel consumptions.

The rest of this paper is organized as follows. Section 2 de-scribes the underhood simplified body designed for the experi-mental investigations. In Section 3, measurements carried out onthe simplified model by PIV, LDV and thermocouples are reported.Section 4 presents the experimental results on the heat exchangersthermal performance and the fan air flow topology as well as a newcontrol approach permitting the adaptation of the cooling moduleof a given vehicle to its engine energy requirements.

Flowmeter

Radiator

Traps

Outlet

Inlet

Boiler with integratedpump

DrainSupply

Main water supply

Flowmeter

Radiator

Traps

Outlet

Inlet

Boiler with integratedpump

DrainSupply

Main water supply

Fig. 4. Schematic diagram of the thermal loop used to measure the thermal powerevacuation of radiator.

2. Car simplified body

The car simplified body is described in [30], here we presentsome main features of it. The considered model is a simplifiedbut representative form of the front end block of a Peugeot 207passenger car (Fig. 1a). The dimensions of the front end projectionare the same as that of the real vehicle. The model is 1.7 m wide,1.3 m long and its height varies from 0.6 m at the front to 1 m atits rear. It includes in its body a real cooling module (a car radiator,a condenser, and a fan) and a simplified engine block (Fig. 1b).These two elements are placed on a traversing system that allowsvarying their positions in the X and Y directions (Fig. 1c). The initialconfiguration is that of a Peugeot 207 where the cooling module isoffset from the center of the car to the right (as Y) and where theengine block is centered in the width (Y) of the car and is posi-tioned at 6 cm from the cooling module. The engine block can thenmove from �6 cm to 6 cm in the Y direction in steps of 2 cm; and6 cm to 20 cm from the cooling module in the X following steps of2 cm. The air inlet of the model (Fig. 1d) is flat and flexible: with afixed frame and several windows, it can represent different

geometries and sizes of air inlets used in vehicle series, by closingor opening a given and specific number of windows.

The car simplified body is provided with several types of air out-lets: a vertical exit, a tunnel passage (representing the outlet in thetunnel exhaust of a real vehicle), openings in the wheel arches andoutlets on the underbody. In the experiments, not all the outletsare open simultaneously; combinations between different types ofopenings permit to simulate configurations of different car series:for example when the vertical outlet and outlets in the underbodyare closed by leaving open outlets in the tunnel and the wheel ar-ches, one can simulate the configurations of conventional vehiclesseries of PSA Peugeot Citroen. On the other hand, by keeping onlythe exit in the tunnel exhaust open, typical configurations of BMWvehicles are represented.

Fig. 5. Grid of air thermocouples placed downstream of the cooling module.

ExchangersEngine blockFan


3. Experimental setup

Aerodynamic and thermal measurements are carried out on thecar simplified body. Aerodynamic measurements are carried out byLaser Doppler Velocimetry (LDV) (Section 3.1) and Particle ImageVelocimetry (PIV) (Section 3.2) and temperatures are measuredusing thermocouples (Section 3.3).

Measurement window

Air

X

YZ

6 cm

VortexX

YZ

6 cm
3.1. LDV measurements
For LDV measurements, the model is adjusted in a manner toreproduce the typical air inlets and outlets of a Peugeot 207 pas-senger car: the air inlets are of identical dimensions as that ofthe Peugeot 207, the air outlets are in the exhaust tunnel and inthe wheel arches; the ratio between inlet and outlet sections isabout 0.6. Paraffin oil (seeding) is added to air flow in front ofthe model (upstream of the fan) and the LDV probe is mountedon a three-dimensional traversing system and placed next to themodel (Fig. 2).

Two fan rotational speeds are studied: 1400 rpm and 2800 rpm.They are both the nominal operating speeds on actual Peugeot 207vehicle. For each speed, eight configurations of the engine blockage(eight spaces between the cooling module and the engine block)are studied: 6, 8, 10, 12, 14, 16, 18 and 20 cm. The different spacingconfigurations are obtained by keeping the cooling module fixedand moving backward the engine block in increments of 2 cm inthe X direction (car movement direction), the initial spacing being6 cm. For each configuration (given fan rotation speed and radia-tor-engine block spacing), the horizontal velocity component Uand vertical velocity component W (in the YZ plane) are measuredat X = 1 cm downstream of the fan.

3.2. PIV measurements

Contrary to a blower fan (located upstream of the heat exchang-ers), the air flow distribution induced by a suction fan (as is thecase in the simplified body) in the underhood is no longer

Table 1Important parameters varied in the 48 different configurations examined in thiswork.

Water flowrate (l/min)

Fan speed (rpm) Cooling module/engineblock spacing (cm)

8 and 10 1400, 2800 and 3300 6, 8, 10, 12, 14, 16, 18, 20

Impacting air jetFig. 6. Averaged velocity field in a window of the lower part of the plane XZ � Y = 0of the main flow. Velocity fields presented here are obtained by PIV. Fan speed is2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and60 �C respectively.


conditioned by the cooling module. In this case, a swirling airflowcan be generated and a secondary flow appears in addition to themain flow, giving rise to a three-dimensional structure to theresulting air flow. In order to obtain a three-dimensional pictureof the resulting flows, we study first the topology of the primaryand the secondary flows by PIV measurements. The main flow(XZ plane) is in a plane normal to the fan plane (Fig. 3a) and thesecondary flow (YZ plane) is in a vertical plane parallel to the fanplane (Fig. 3b).

By fixing the laser source inside the engine block (Fig. 3c), thePIV camera can be put on either the right or left side of the simpli-fied body (Fig. 3d) to visualize the flow downstream of the fan.Measurements are made for different fan angular speeds (includ-ing those corresponding to functioning in a real vehicle) and differ-ent spaces between the cooling module and the engine block. Then,through several windows in the measurement plane and movingthem in Y direction (i.e. on the width of the model), one can obtainan appropriate description of the main air flow between the fanand the engine block.

For the topology of the secondary flow (detection of theswirling phenomenon), positions of the camera and the lasersource are permuted. Positioned on the right side or left sideof the model, the laser illuminates a plane parallel to the fanplane and the camera takes pictures of this plan within the en-gine block. For several positions of the camera with respect tothe laser light plane, several windows of secondary flow wereobtained. With these different windows, we covered an arealarge enough to probe the swirling flow. Note that other laser/camera positions can be applied (e.g. the laser above the modeland the camera below the model. . .), but we always chose thepositions with less laser light attenuation and the sharpestimages.

Fig. 7. Temperature increase (heat transfer enhancement) at the low part of the enginepresented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flo

3.3. Thermal measurements

The purpose of such measures is to study the radiator perfor-mance for different fan speeds and relative engine block positions(including the one which corresponds to a real underhood of a Peu-geot 207), that is the relation between the underhood aerodynamicand the heat exchangers and their environments. The thermal loopused for the measurement of radiator efficiency is shown in Fig. 4.The radiator is fed by a water circuit that controls the inlet temper-ature. The circuit essentially comprises a variable power boiler, theradiator of the cooling model (a real Peugeot 207 radiator), a mainwater supply, a flowmeter, valves and traps. The boiler has a threespeeds (8, 10 and 12 l/min) built in pump and a temperature probewhich allows to fix the hot temperature (the temperature at the inletof the radiator) following instruction requested. Two 80 lm beaddiameter thermocouples (type K) are positioned at the inlet andthe outlet of the radiator. A grid of 30 thermocouples (80-microns-diameters type K) is placed downstream of the cooling module(Fig. 5) which allows measuring the air temperature distributionon a plane downstream of the cooling module. The thermocouplegrid was placed on a plane located 3 cm (in the X-direction) down-stream of the cooling module and parallel to its surface. The thermo-couples are distributed equally spaced in both X and Y directions. Toensure the reliability of the temperature measurement by the ther-mocouples grid with respect to vibration and deformation, we fixedthe thermocouples on a grid made of very thin and rigid wires. Thewire diameter was chosen in such a way that the Reynolds numberfor generation of Karman vortices, which are the cause of wire vibra-tion and deformation, was under critical. Reproducibility tests werecarried out and showed satisfactory results.

The engine block surface opposite to the cooling module isinstrumented with 27 type T thermocouples of 1 mm diameter

block in relation with the air jet flow apparition (plane XZ � Y = 0). Velocity fieldsw rate and inlet temperature are 8 l/min and 60 �C respectively.


beads, each in order to obtain the temperature distribution on theengine block surface.

Three fan speeds are tested: the two nominal speeds of 1400and 2800 rpm and a maximum speed of 3300 rpm. Two differentwater flow rates are imposed: 8 l/min and 10 l/min. Finally, eightconfigurations for positioning of the engine block in the X-directionwith respect to the cooling module are tested: 6, 8, 10, 12, 14, 16,18 and 20 cm. For each configuration, water flow, fan speed, posi-tion of the engine block, the radiator inlet temperature, and thetemperatures at the different locations described above are re-corded (59 locations).

Table 1 summarizes the different parameters varied during theexperiments (aerodynamic and thermal measurements), all corre-sponding to an inlet water temperature of 60 �C. A total of2 � 3 � 8 = 48 configurations were examined.

4. Results and discussion

Results presented here concern the heat transfer enhancementof the cooling module of a vehicle as well as the air flow inducedby the fan and the aerothermal effects of the engine blockage onthis air flow (Section 4.1). These results are then coupled with ascheme for active control of a vehicle cooling module (Section 4.2).

Fig. 8. Averaged velocity fields in windows at the middle and high parts of theplane XZ � Y = 0 of the main flow. Velocity fields presented here are obtained byPIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlettemperature are 8 l/min and 60 �C respectively.

4.1. Fan airflow and engine blockage effects

In experiments, ‘‘suction’’ type fan is placed behind the heatexchangers. Contrary to a ‘‘blowing’’ fan, which is located upstreamof the heat exchangers, flow induced by a suction fan is likely togenerate a three-dimensional flow with a large and strong swirlin the YZ plane. On the other hand, the presence of the engine blockat only 6 cm downstream of the fan makes the topology of the airflow complex. Fig. 6 shows the velocity fields measured by PIV inthe mid-height of radiator (at Y = 0) for a fan speed of 2800 rpm.A downward diagonal air flow is noticed in this measurement win-dow. The diagonal direction of the flow is due to the difference inpressure between the region directly downstream of the fan andthat of depression generated by the geometric restrictions belowthe engine block. On the other hand, a jet-type flow appears inthe right part of the velocity field due to the air impact on the en-gine block. Reflected by the engine block, the jet flow is deflected ina direction diagonally opposite to the direction of the main flowwith which it interacts (see the upper part of the air jet). The com-bination between the descending main flow and the reverseascending jet generates a vortex midway from the engine blockand the cooling module. The jet flow at the lower engine block re-gion promotes the convective heat transfer in this part.

Fig. 7 shows the temperature distribution in the jet area and thetemperature profile over the total height of the engine block. Mea-surements are performed for the radiator fed with hot water at60 �C. The air heated through the heat exchanger passage increasesthe temperature of the engine block (which in the simplified modelis not a source of heat). A warmer area in the central part of the en-gine-block with a temperature peak located at the jet impact posi-tion is noticed.

Going up in the Z direction (in the direction of the simplified mod-el height), an upward vertical flow appears clearly in the middle andupper zones between the cooling module and the engine block(Fig. 8). This progressive emergence of a vertical velocity componentis mainly due to the abrupt return flow that creates the engine-blocknear the fan and the depression prevalence generated at the engine-block and above.

In Fig. 8, it is noticed that the air jet flow appeared in the lowerpart is not reproduced in the upper part. However, a vortex is de-tected at the lower part of the velocity field. This vortex results

from the interaction between the diagonally upward main flowand the flow near the engine block.

Finally, the central portion (window 1 in Fig. 8) of the planeY = 0 is always characterized by upward diagonalization of flowwith the appearance of interaction between the diagonal flowdue to the fan and the reflected flow from the engine block. Onecan distinguish clearly that the air tends to rise rather than fall.This is due to the blockage of the flow at the engine block lower re-gion and to the attracting flow at the upper ends of the engineblock.

To characterize the topology of the secondary flow, six windowsmarked from 1 to 6 of the YZ plane located in the mid-distancefrom the engine block and the cooling module are considered inFig. 9; three distinct zones are shown:


– the first at the center is characterized by a large number ofvortices distributed randomly in the plane (window 3 ofFig. 9),

– two circular shaped areas (zones) of revolution axis parallelto that of the fan where the flow follows the same directionas the rotating fan blades (the dashed and solid arc linespassing through windows 1, 2, 4, 5 and 6 of Fig. 9),

– the vortex area at the center generated by the interactionbetween the dead zone caused by the presence of the fanhub and the flow into two swirl circles around it.

In conclusion, the flow resulting from a suction type fan has avery strong threedimensional structure, characterized by theappearance of particular structures in the main flow (eddies, jetair flow. . .) and a swirling flow creating a dead zone vortex behindthe fan center. In particular, the topology of the main flow is likely

Fig. 9. Swirling phenomenon in secondary flow – Averaged velocity fields obtained byobtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet te

to be significantly dependent on the blockage induced by the en-gine block on the fan flow. Thus, it is crucial to investigate the ef-fect of variation of the distance d between the cooling module andthe engine block on the flow induced by the fan.

To study the engine blockage effect, the distance d between thecooling module and the engine block is varied. For this, the positionof the cooling module is fixed while the engine block is incremen-tally moved backward: d ranges from 6 to 20 cm with 2 cmincrements.

This study has explicitly shown that the particular structures ofthe main flow observed in the configuration d = 6 cm (see the abovedescription) are largely due to the proximity of the engine blockdownstream of the fan. Indeed, the air jet flow present in this posi-tion disappears for the longer distances (above 8 cm). Similarly,the different vortices that appeared for d = 6 cm (following the inter-action between the fan airflow and the flow) which ‘‘bounced’’ back

PIV measurements in the plane YZ � X = 3 cm. Velocity fields presented here aremperature are 8 l/min and 60 �C respectively.


are not present for the other distances. Another typical difference isthat from the distance of 8 cm on, a reverse flow appears in the cen-tral area, where negative longitudinal velocities are observed. Fig. 10shows the distribution of the flow velocity component U in the Y = 0plane, downstream of the fan for d = 6 cm, 8 cm and 10 cm. Thesemeasurements are obtained by PIV and for the fan functioning at2800 rpm. The blue color corresponds to negative values of U. Thus,it is seen that the reverse flow appears from the distance beyond8 cm in the measurement area.

Effects of the distance between the radiator and the engineblock are not noticeable on the topology of the secondary flow.The swirling phenomenon shown for d = 6 cm persists for the otherdistances, although there were slight changes in velocity compo-nents V and W.

Since it appears that it is essentially the main flow which is sen-sitive to the engine blockage effect, in the following the study will befocussed on the flow velocity component U, aiming to link thechanges in the flow with the thermal modifications: thermal powerevacuated by heat exchangers, air temperature downstream of theheat exchangers and temperature at the engine block. Fig. 11 showsthe distribution of the axial velocity component U obtained by LDVin a YZ plane located at X = 2 cm from the cooling module. The mea-surement planes shown are of 38 cm width (Y) and 32 cm height (Z)with 340 LDV measurement points. It can be noted that the flowtopology in a YZ plane directly downstream of the fan is slightlymodified by moving the engine block backward:

– for d = 6 cm, traces of the fan blades on right of the pictureare not yet present and those of left are not yet complete,

– for d = 8 cm, the traces at left are completed while those atright start to appear in the lower part of the picture,

Fig. 10. Axial flow velocity (U) distributions in the plane Y = 0 fo

– for d = 10 cm, the traces of the fan blades at the top rightbegin to form and become complete for d = 14 cm, the dis-tance at which the topology of the flow velocity becomesalmost invariant. Traces of fan blades correspond to theparts of the picture that are in red–orange. The same fea-tures are observed for the other fan rotation speeds.

However, while the flow topology is slightly modified byincreasing the distance d, the flow statistics vary significantly.Fig. 12 shows the changes with the distance d in the average speedand the standard deviation of the flow velocity distribution in theYZ plane for X = 2 cm. It is noticed that by increasing the distanced:

– the mean flow velocity increases, i.e. the air mass flow ratethrough the cooling module increases: at 2800 rpm, the airflow rate increases by 40% between d = 6 cm and d = 20 cm,

– the standard deviation of the flow velocity increases too,i.e. the flow becomes increasingly non-homogeneous.Again, at 2800 rpm, when d increases from 6 to 20 cm,the standard deviation increases by more than 40%.

Thus, approaching the engine block to the cooling module at adistance of 6 cm blocks the flow but makes it more homogeneous.By pushing backward the engine block from the radiator, the flowinduced by the fan meets less resistance and the flow velocity, butthe non-homogeneity increases too.

The increase in the non-homogeneity of the velocity distribu-tion through a heat exchanger decreases the capacity of the heatexchanger to evacuate heat (contrary to the increase in the meanspeed). Moreover, even if the decrease in the engine blockage

r different distances between the fan and the engine block.

Fig. 11. Flow velocity fields in the YZ plane and X = 2 cm for different distances d at a fan rotational speed of 2800 rpm – LDV measurements.


increases the evacuation power of the heat exchanger, it will in-crease also the air temperature downstream of the fan whichmay penalize the cooling of the underhood bodies exposed to the

airflow. To quantify this thermal behavior, the outlet temperatureof the radiator cooling water as well as air temperature distribu-tion on a plane downstream of the fan are measured for a given

1,50

1,70

1,90

2,10

2,30

2,50

2,70

6 8 10 12 14 16 18

Distance (cm)

Air

velo

city

(m/s

)

Mean air velocityStandard deviation air velocity

3,40

3,90

4,40

4,90

5,40

6 8 10 12 14 16 18Distance (cm)

Air

velo

city

(m/s

)


(a)

(b)

1,50

1,70

1,90

2,10

2,30

2,50

2,70

6 8 10 12 14 16 18


3,40

3,90

4,40

4,90

5,40

6 8 10 12 14 16 18


Fig. 12. Variations of the mean and the standard deviation of axial flow velocityfield U, as a function of the distance d between the fan and the cooling module.Values are integrated through the plane X = 2 cm. Two fan rotational speeds aretested: (a) 1400 rpm and (b) 2800 rpm.

1,00

1,50

2,00

2,50

3,00

3,50

4,00

4,50

6 8 10 12 14 16 18

Distance (cm)

P ra

diat

or (k

W)

Fan low speedFan high speed

79,0

80,0

81,0

82,0

83,0

84,0

85,0

86,0

87,0

88,0

6 8 10 12 14 16 18

Distance (cm)

T ai

r (°

C) Fan low speed

Fan high speed

(a)

(b)

1,00

1,50

2,00

2,50

3,00

3,50

4,00

4,50

6 8 10 12 14 16 18

Fan low speedFan higgh speed

79,0

80,0

81,0

82,0

83,0

84,0

85,0

86,0

87,0

88,0

6 8 10 12 14 16 18

Fan low speedFan high speed

Fig. 13. Variations of the (a) radiator thermal power and (b) air temperaturedownstream of the fan, as a functions of the distance d for two nominal fan speeds.Water flow rate is 8 l/min.

Table 2Percentages of deviation from the reference configuration (d = 6 cm) on the heatexchanger thermal power evacuation and air temperature downstream of the fan.

Radiator thermal power Downstream airtemperature

Distance, d(cm)

Fan lowspeed

Fan highspeed

Fan lowspeed

Fan highspeed

8 17.75 7.06 �1.24 �0.5210 26.09 12.85 �4.91 �1.7312 29.96 16.14 �6.65 �2.9314 34.78 18.26 �8.58 �4.4316 36.52 19.37 �7.85 �5.1618 36.56 19.93 �7.29 �4.5820 37.84 22.00 �8.55 �3.74


radiator water inlet temperature (Fig. 5). These experiments wereperformed for different positions of the engine block and differentwater flow rates. Fig. 13 shows the variation of the radiator evacu-ated thermal power and the averaged air temperature at the fandownstream, as a function of the spacing between the coolingmodule and the engine block, for fan nominal speeds of 1400and 2800 rpm.

For the two fan operating speeds, it was found that the thermalpower evacuated by the radiator increases as the distance d in-creases. For example, for d ranging from 6 to 18 cm, the powerevacuated by the radiator is increased by 37% at the fan low speed.At high speed, this power increases by about 20%. Paradoxically,the air mean temperature downstream of the fan strongly de-creases when d increases from 6 to 14 cm (18 cm for the case at2800 rpm). For example, at a fan speed of 1400 rpm, the air tem-perature decreases from 87 �C to 79.5 �C, a relative decrease of9%. From 14 cm, the air temperature downstream of the fan beginsto increase slightly to attain 80.67 �C when d = 18 cm. This behav-ior reflects the competition between two opposite effects: byincreasing the distance d, the evacuated power by the radiator in-creases and thus the heat absorbed by the air passing through theheat exchanger also increases. At the same time, the airflow ratethrough the heat exchanger increases. Between d = 6 and 14 cmthe airflow increase dominates the absorbed thermal power,resulting in a decrease in the air temperature downstream of thefan. The opposite situation occurs for d higher than 14 cm.

Table 2 reports the influence (in percentage) of the distance don the thermal power evacuated by the radiator and the air tem-perature downstream of the fan.

It is noted that the increase in d causes useful variations in thedifferent parameters. In the particular case of a Peugeot 207(d = 6 cm) for example, if the engine block is moved backward bya distance of 2 cm the following results will be observed:

– an increase in the thermal power evacuated by the radiatorof about 18% when the fan operates at low speed and 7%when it operates at high speed

– a decrease in the air temperature downstream of the radi-ator of about 5% at low speed and 2% at high speed.

Applicability of the above results depends, of course, on theavailability of space in the underhood. Generally, additional 10–12 cm is available in the underhood for the displacement of the

Cooling module

Water pumpEngine block

Radiator inlet duct

6 cm

AutomaticCommand

system1

2

2

Cooling module


Radiator inlet duct

6 cm

AutomaticCommand

system1

2

2

Cooling module


Radiator inlet duct

6 cm

AutomaticCommand

system1

2

2

Cooling module


Radiator inlet duct

6 cm

AutomaticCommand

system1

2

2


engine block with respect to the cooling module. However, thisspace varies from one vehicle to another: for example in a Peugeot207, approximately 6 cm additional space is available, while for aCitroën C6 this space varies between 10 and 12 cm approximately.Therefore, geometrically there is no potential restrictions if one re-spects the other allowable distances for an existing car design (forexample Peugeot 207, 307, Citroen C2, C6, etc.). However, for a newcar design, a choice has to be made if there is a strong downsizingdemand of the underhood.

Control thermocouple

X


X

Cooling module


Radiator inlet duct

AutomaticCommand

system


1

2

2

10 c

m

X

Cooling module


Radiator inlet duct

AutomaticCommand

system


1

2

2

10 c

m

X

Cooling module

Climatecompressor

Engine block

Condenser inlet duct

6 cm

Automatic Command

system


1

2

2

X

Cooling module

Climatecompressor

Engine block


6 cm

Automatic Command

system


1

2

2

X

(a)

(b)

(c)


X


X

Cooling module


Radiator inlet duct

AutomaticCommand

system


1

2

2

10 c

m

X

Cooling module

Engine block

Radiator inlet duct

AutomaticCommand

system


1

2

2

10 c

m

X

Cooling module

Climatecompressor

Engine block


6 cm

Automatic Command

system


1

2

2

X

Cooling module

Climatecompressor

Engine block


6 cm

Automatic Command

system


1

2

2

X

(a)

(b)

(c)Fig. 14. (a) Technical description of the first control principle for the radiator ductdisplacement (b) systems after application of the first control approach (c) technicaldescription of the first control principle for the condenser duct displacement.

4.2. Active control of vehicle cooling modules

In this section we describe a new monitoring tool [31] that canbe used to optimize the underhood aerothermal management. Itpermits to adapt the heat exchangers’ performance to the engineenergy requirements and to increase the heat exchangers’ heatevacuation power in critical situations such as the slowdown phaseor the thermal soak phase of a vehicle (vehicle stop after a signif-icant heating load).

Motivated by the results of Section 4.1 on the aerothermal ef-fects of the distance d between the cooling module and the engineblock, the control scheme presented here is based on moving away(increase d) in the X-direction (direction of the vehicle length) thecooling module from the engine block in order to increase the heatexchangers evacuation power in critical situations where the en-gine overheats under conditions of vehicle slowdown or vehiclethermal soak. This control is managed by a thermocouple placedin the cooling water upstream of the radiator inlet (the engine out-let). Depending on the radiator inlet temperature in a given ther-mal situation during vehicle slowdown or thermal soak with thefan functioning phases, the heat exchangers energy requirementsand the need to increase in heat evacuation power compared tothe normal situation are determined and the cooling module isthen moved to a distance that corresponds to the desired increasein power.

To implement the control approach we apply it to the case of asimplified vehicle cooling module consisting only of a condenserand a radiator (as is the case in a Peugeot 207 without turbo-com-pressor for example). On other vehicle types, the same principlesare applicable to other heat exchangers (such as the charged aircooler or engine oil cooler).

The basic principle of the control scenario is described inFig. 14a. In this Figure and for the sake of simplification, the controlprinciple is presented only for the radiator; the same principle isapplicable for the other heat exchangers (and is shown also forthe condenser in Fig. 13c).

The thermocouple placed at the radiator inlet returns the instan-taneous temperature value to the automatic control system (step 1marked by black square in the technical description). This, in turn,verifies in real time if the temperature reaches critical values Ti.On the other hand, in the control system is registered a correspon-dence law between the critical temperatures Ti and the distancesdi between the cooling module and the engine block. Wheneverone of the temperatures Ti is reached, the control system automati-cally commands the cooling module to move for a distance di fromthe engine block which corresponds to values recorded at the tem-perature Ti reached. It also controls the elongation of the connectionpipes to the same distance di (both commands described correspondto the number 2 in black square of the technical description). Fig. 14bshows the system of Fig. 14a (reference configuration) after theapplication of the control.

The active control consists therefore of the simultaneous appli-cation of the various commands described above (in Fig. 14a and c).In the case of several heat heat exchangers, other than the twomain (radiator and condenser), the control is applied to the

different heat exchanger pipes at the same time as the displace-ment in the X-direction of the cooling module.

The economic interest of the proposed control approach can beseen in the application of its design to normal and non critical engineoperation. If the control is used in normal vehicle operations, it canthen increase the thermal power of the cooling module for the samework delivered for the water pump and the air conditioning systemcompressor. Thus, one can have the same evacuated thermal powersby the heat exchangers for smaller pumping and compressor works,which reduces the vehicle fuel consumption (since the pump andcompressor works correspond to losses of the engine power).

5. Conclusions

To allow the cooling module (several heat exchangers and oneor two fans) of a vehicle to evacuate heat in difficult operating


conditions, it has been pushed backward into the engine compart-ment. Therefore, the aeraulic and aerothermal situations in theunderhood are largely conditioned by the cooling module whichperformance continues to present demanding aerothermal man-agement challenges in their design as well as their interaction withother components in the underhood.

In this work, experimental aerodynamic and thermal investiga-tions were performed on a vehicle simplified body. These investi-gations have enabled us to study the flow induced by the fan andthe blockage effect imposed by the engine to this flow, in conjunc-tion with the thermal performance of real cooling module inte-grated in the model. In particular, we discuss, by the PIV andLDV measurements, the three-dimensional character of the fanflow. A first component of the threedimensional flow was the swirlphenomenon in planes parallel to the fan plane. A second corre-sponded to the main flow in the direction of the fan axis: proper-ties of the main flow proved to be strongly related to thedistance d between the cooling module and the engine block. Forexample, increasing the distance d, the air jet which appears ford = 6 cm does no longer exist. It is also found that by increasingthe distance d:

– the air flow rate through the heat exchanger increases,– the thermal power evacuated by the radiator increases,– the air temperature downstream of the cooling module

decreases.

Finally, based on the experimental results, a new monitoringtool that can be used for optimizing the underhood aerothermalmanagement was proposed. It permits to adapt the heat exchang-ers’ performances to the engine energy requirements and toincrease the heat exchangers’ evacuation power in critical situa-tions such as the slowdown phase or the thermal soak phase of avehicle (vehicle stops after a significant heating load). The aim hereis to apply this new control approach to a real vehicle so as to re-duce the energy delivered to the pump and compressor functions,and therefore to reduce the vehicle fuel consumption.

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