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Fan air flow analysis and heat transfer enhancement of vehicle underhood cooling system – Towards a new control approach for fuel consumption reduction
Applied Energy 91 (2012) 439–450
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Fan air flow analysis and heat transfer enhancement of vehicle underhood cooling 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, France PSA Peugeot Citroën, Velizy A Center, 2 route de Gisy, 78 943 Vélizy Villacoublay, France c Fluid Mechanics, Heat and Thermodynamics Group, School of Engineering, Lebanese International University, Beirut, Lebanon b
a r t i c l e
i n f o
Article history: Received 12 August 2011 Received in revised form 9 October 2011 Accepted 12 October 2011 Available online 9 November 2011 Keywords: Vehicle underhood Car cooling module Heat exchangers Heat transfer enhancement Fan Car underhood aerothermal management
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 positioning according to the engine energy requirements. Measurements are carried out on a simplified vehicle body designed based on the real vehicle front block. We report here velocity and temperature measurements by Particle Image Velocimetry (PIV), by Laser Doppler Velocimetry (LDV) and by thermocouples. The underhood of the simplified body is instrumented by 59 surface and fluid thermocouples. Measurements are carried out for conditions simulating both the slowdown and the thermal soak phases with the fan in operation. Different fan rotational speeds, radiator water flow and underhood geometries have been experimented. The ultimate aim is to apply the new control approach to a real vehicle so as to reduce the energy delivered to the pump and compressor and therefore to reduce the vehicle fuel consumption. Ó 2011 Elsevier Ltd. All rights reserved.
1. Introduction The cooling module of vehicle constituted of an assembly of different heat exchangers and one or two fans have been pushed backward into the engine compartment, in order to evacuate heat in difficult operating conditions. This have caused the underhood aerothermal and aeraulic conditions to be largely conditioned by the different elements of the cooling module which at the same time interact with other components in the underhood. In heat exchangers’ applications which cover a large variety of technological domains [1–10] (such as nuclear reactors, chemical and biomedical processes, military and aerospace, and automotive), the trend is to reduce the heat exchangers’ weight and volume while keeping sufficient or even increasing thermal efficiencies. In the open literature, the large part of optimization studies is focused on the geometry and the design of the heat exchangers and also on the fluid flows structures. Several design and control approaches aiming to optimize the heat exchangers’ thermal performance are reported in [11–20], while few studies are focussed on the heat exchangers’ optimization taking into account their interaction with the geometry in which they are installed. Among the most commonly heat exchangers used are the tubes and fins heat exchangers which are composed of several tubes of elliptical cross section between which are positioned several ⇑ Corresponding author. Tel.: +33 2 40 68 31 24; fax: +33 2 40 68 31 41. E-mail address: [email protected] (H. Peerhossaini). 0306-2619/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2011.10.017
continuous and parallel fins. These types of heat exchangers are particularly adapted to most industrial applications, especially in automotive applications, and their fins can be simple, louvered, crimped, shutters, or with delta wings longitudinal vortex generators, etc. Several studies [21–26] on the louvered fin heat exchangers for underhood applications showed that the thermal performance of such exchangers are strongly affected by the geometrical parameters of the heat exchanger (e.g. distance between fins, the tubes height, and fins angle) and operating conditions of the two fluids, air and water (their mass flow rates and inlet temperatures). However, few studies have explicitly addressed the crucial question of how to control the different parameters for enhancing the exchanger’s heat evacuation capacity. Most of the underhood heat transfer analysis has been concentrated on the underhood thermal management by temperature and heat flux measurements [27–34]. In the present study we have experimentally investigated the heat exchanger optimization by acting on the geometry (underhood space here) in which they are integrated. The present analysis shows how to act on the different parameters which influence the heat exchanger performance in order to optimize its functioning. On the other hand, in critical phases of vehicle functioning such as cooling at the vehicle stopped situation, the stop and go driving mode, and the vehicle thermal soak, the performance of car heat exchanger is essentially influenced by the air flow generated by the fan. This air flow in its turn depends on the underhood architecture. Most experimental 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.
determination of the fan aerothermal characteristics (curves of pressure rise as a function of the air flow) or optimization studies (use of rectifiers in order to guide the airflow) where the fan alone is considered. On real cars, these studies are quite limited given the constraints of accessibility related to the presence of engine block downstream of the fan. For instance, it is very difficult to use optical measurement techniques such as PIV or LDV in the car underhood space. One solution is to work on underhood simplified models designed to allow the integration of optical measurement techniques. Compared to studies on isolated fan, these models present the advantage of reproducing the effects of the engine blocking imposed on the fan flow. This ‘‘blocking’’ due to the short distance between the fan and the engine block downstream, is a parameter that could affect the characteristics of the flow induced by the fan in the car underhood. Moreover, the space between the cooling module and the engine block varies from one vehicle type to another. For these
reasons, it was decided in the present study to work on a car underhood model containing an actual cooling module and simplified engine block both translatable in X and Y (respectively in the vehicle length and width) directions. This modularity allows covering a wide range of underhood architectures to explore different strategies adapted by the automobile manufacturers. Part of this paper will thus be focussed on the PIV and LDV measurements as well as thermal measurements performed on the model for different blockages downstream of the fan, that is to say, for different positions of the engine block with respect to the cooling system. The impact of these aerodynamic changes on the thermal performance of the heat exchangers and their surrounding area is also studied. We report here velocity and temperature measurements by Particle Image Velocimetry PIV, by Laser Doppler Velocimetry and by thermocouples. The underhood of the simplified car body is instrumented by 59 surface and fluid thermocouples. Measurements are
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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.
carried out for conditions simulating both the slowdown and the thermal soak phases with the fan in operation. Different fan rotational speeds, radiator water flow and underhood geometries have been tested. The final aim is to apply the new control approach developed here on a real vehicle so as to reduce the energy delivered to the pump and compressor functions and thus to reduce the vehicle fuel consumptions. The rest of this paper is organized as follows. Section 2 describes the underhood simplified body designed for the experimental investigations. In Section 3, measurements carried out on the simplified model by PIV, LDV and thermocouples are reported. Section 4 presents the experimental results on the heat exchangers thermal performance and the fan air flow topology as well as a new control approach permitting the adaptation of the cooling module of a given vehicle to its engine energy requirements.
geometries and sizes of air inlets used in vehicle series, by closing or opening a given and specific number of windows. The car simplified body is provided with several types of air outlets: a vertical exit, a tunnel passage (representing the outlet in the tunnel exhaust of a real vehicle), openings in the wheel arches and outlets on the underbody. In the experiments, not all the outlets are open simultaneously; combinations between different types of openings permit to simulate configurations of different car series: for example when the vertical outlet and outlets in the underbody are closed by leaving open outlets in the tunnel and the wheel arches, one can simulate the configurations of conventional vehicles series of PSA Peugeot Citroen. On the other hand, by keeping only the exit in the tunnel exhaust open, typical configurations of BMW vehicles are represented.
2. Car simplified body The car simplified body is described in [30], here we present some main features of it. The considered model is a simplified but representative form of the front end block of a Peugeot 207 passenger car (Fig. 1a). The dimensions of the front end projection are 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 at its 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 allows varying their positions in the X and Y directions (Fig. 1c). The initial configuration is that of a Peugeot 207 where the cooling module is offset from the center of the car to the right (as Y) and where the engine block is centered in the width (Y) of the car and is positioned at 6 cm from the cooling module. The engine block can then move from 6 cm to 6 cm in the Y direction in steps of 2 cm; and 6 cm to 20 cm from the cooling module in the X following steps of 2 cm. The air inlet of the model (Fig. 1d) is flat and flexible: with a fixed frame and several windows, it can represent different
Traps
Inlet Radiator
Outlet Flowmeter
Supply Main water supply
Drain Boiler with integrated pump
Fig. 4. Schematic diagram of the thermal loop used to measure the thermal power evacuation of radiator.
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Fig. 5. Grid of air thermocouples placed downstream of the cooling module.
3. Experimental setup
Exchangers
Aerodynamic and thermal measurements are carried out on the car simplified body. Aerodynamic measurements are carried out by Laser Doppler Velocimetry (LDV) (Section 3.1) and Particle Image Velocimetry (PIV) (Section 3.2) and temperatures are measured using thermocouples (Section 3.3).
Fan
Engine block
6 cm Air
3.1. LDV measurements For LDV measurements, the model is adjusted in a manner to reproduce the typical air inlets and outlets of a Peugeot 207 passenger car: the air inlets are of identical dimensions as that of the Peugeot 207, the air outlets are in the exhaust tunnel and in the wheel arches; the ratio between inlet and outlet sections is about 0.6. Paraffin oil (seeding) is added to air flow in front of the model (upstream of the fan) and the LDV probe is mounted on a three-dimensional traversing system and placed next to the model (Fig. 2). Two fan rotational speeds are studied: 1400 rpm and 2800 rpm. They are both the nominal operating speeds on actual Peugeot 207 vehicle. 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 spacing configurations are obtained by keeping the cooling module fixed and moving backward the engine block in increments of 2 cm in the X direction (car movement direction), the initial spacing being 6 cm. For each configuration (given fan rotation speed and radiator-engine block spacing), the horizontal velocity component U and vertical velocity component W (in the YZ plane) are measured at X = 1 cm downstream of the fan.
Z Y
Measurement window X Vortex
3.2. PIV measurements Contrary to a blower fan (located upstream of the heat exchangers), the air flow distribution induced by a suction fan (as is the case in the simplified body) in the underhood is no longer Table 1 Important parameters varied in the 48 different configurations examined in this work. Water flow rate (l/min)
Fan speed (rpm)
Cooling module/engine block spacing (cm)
8 and 10
1400, 2800 and 3300
6, 8, 10, 12, 14, 16, 18, 20
Impacting air jet Fig. 6. Averaged velocity field in a window of the lower part of the plane XZ Y = 0 of the main flow. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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conditioned by the cooling module. In this case, a swirling airflow can be generated and a secondary flow appears in addition to the main flow, giving rise to a three-dimensional structure to the resulting air flow. In order to obtain a three-dimensional picture of the resulting flows, we study first the topology of the primary and the secondary flows by PIV measurements. The main flow (XZ plane) is in a plane normal to the fan plane (Fig. 3a) and the secondary flow (YZ plane) is in a vertical plane parallel to the fan plane (Fig. 3b). By fixing the laser source inside the engine block (Fig. 3c), the PIV camera can be put on either the right or left side of the simplified body (Fig. 3d) to visualize the flow downstream of the fan. Measurements are made for different fan angular speeds (including those corresponding to functioning in a real vehicle) and different spaces between the cooling module and the engine block. Then, through several windows in the measurement plane and moving them in Y direction (i.e. on the width of the model), one can obtain an appropriate description of the main air flow between the fan and the engine block. For the topology of the secondary flow (detection of the swirling phenomenon), positions of the camera and the laser source are permuted. Positioned on the right side or left side of the model, the laser illuminates a plane parallel to the fan plane and the camera takes pictures of this plan within the engine block. For several positions of the camera with respect to the laser light plane, several windows of secondary flow were obtained. With these different windows, we covered an area large enough to probe the swirling flow. Note that other laser/ camera positions can be applied (e.g. the laser above the model and the camera below the model. . .), but we always chose the positions with less laser light attenuation and the sharpest images.
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3.3. Thermal measurements The purpose of such measures is to study the radiator performance for different fan speeds and relative engine block positions (including the one which corresponds to a real underhood of a Peugeot 207), that is the relation between the underhood aerodynamic and the heat exchangers and their environments. The thermal loop used for the measurement of radiator efficiency is shown in Fig. 4. The radiator is fed by a water circuit that controls the inlet temperature. The circuit essentially comprises a variable power boiler, the radiator of the cooling model (a real Peugeot 207 radiator), a main water supply, a flowmeter, valves and traps. The boiler has a three speeds (8, 10 and 12 l/min) built in pump and a temperature probe which allows to fix the hot temperature (the temperature at the inlet of the radiator) following instruction requested. Two 80 lm bead diameter thermocouples (type K) are positioned at the inlet and the outlet of the radiator. A grid of 30 thermocouples (80-micronsdiameters type K) is placed downstream of the cooling module (Fig. 5) which allows measuring the air temperature distribution on a plane downstream of the cooling module. The thermocouple grid was placed on a plane located 3 cm (in the X-direction) downstream of the cooling module and parallel to its surface. The thermocouples are distributed equally spaced in both X and Y directions. To ensure the reliability of the temperature measurement by the thermocouples grid with respect to vibration and deformation, we fixed the thermocouples on a grid made of very thin and rigid wires. The wire diameter was chosen in such a way that the Reynolds number for generation of Karman vortices, which are the cause of wire vibration and deformation, was under critical. Reproducibility tests were carried out and showed satisfactory results. The engine block surface opposite to the cooling module is instrumented with 27 type T thermocouples of 1 mm diameter
Fig. 7. Temperature increase (heat transfer enhancement) at the low part of the engine block in relation with the air jet flow apparition (plane XZ Y = 0). Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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beads, each in order to obtain the temperature distribution on the engine block surface. Three fan speeds are tested: the two nominal speeds of 1400 and 2800 rpm and a maximum speed of 3300 rpm. Two different water flow rates are imposed: 8 l/min and 10 l/min. Finally, eight configurations for positioning of the engine block in the X-direction with respect to the cooling module are tested: 6, 8, 10, 12, 14, 16, 18 and 20 cm. For each configuration, water flow, fan speed, position of the engine block, the radiator inlet temperature, and the temperatures at the different locations described above are recorded (59 locations). Table 1 summarizes the different parameters varied during the experiments (aerodynamic and thermal measurements), all corresponding to an inlet water temperature of 60 °C. A total of 2 3 8 = 48 configurations were examined. 4. Results and discussion Results presented here concern the heat transfer enhancement of the cooling module of a vehicle as well as the air flow induced by the fan and the aerothermal effects of the engine blockage on this air flow (Section 4.1). These results are then coupled with a scheme for active control of a vehicle cooling module (Section 4.2). 4.1. Fan airflow and engine blockage effects In experiments, ‘‘suction’’ type fan is placed behind the heat exchangers. Contrary to a ‘‘blowing’’ fan, which is located upstream of the heat exchangers, flow induced by a suction fan is likely to generate a three-dimensional flow with a large and strong swirl in the YZ plane. On the other hand, the presence of the engine block at only 6 cm downstream of the fan makes the topology of the air flow complex. Fig. 6 shows the velocity fields measured by PIV in the mid-height of radiator (at Y = 0) for a fan speed of 2800 rpm. A downward diagonal air flow is noticed in this measurement window. The diagonal direction of the flow is due to the difference in pressure between the region directly downstream of the fan and that of depression generated by the geometric restrictions below the engine block. On the other hand, a jet-type flow appears in the right part of the velocity field due to the air impact on the engine block. Reflected by the engine block, the jet flow is deflected in a direction diagonally opposite to the direction of the main flow with which it interacts (see the upper part of the air jet). The combination between the descending main flow and the reverse ascending jet generates a vortex midway from the engine block and the cooling module. The jet flow at the lower engine block region promotes the convective heat transfer in this part. Fig. 7 shows the temperature distribution in the jet area and the temperature profile over the total height of the engine block. Measurements are performed for the radiator fed with hot water at 60 °C. The air heated through the heat exchanger passage increases the temperature of the engine block (which in the simplified model is not a source of heat). A warmer area in the central part of the engine-block with a temperature peak located at the jet impact position is noticed. Going up in the Z direction (in the direction of the simplified model height), an upward vertical flow appears clearly in the middle and upper zones between the cooling module and the engine block (Fig. 8). This progressive emergence of a vertical velocity component is mainly due to the abrupt return flow that creates the engine-block near the fan and the depression prevalence generated at the engineblock and above. In Fig. 8, it is noticed that the air jet flow appeared in the lower part is not reproduced in the upper part. However, a vortex is detected at the lower part of the velocity field. This vortex results
Fig. 8. Averaged velocity fields in windows at the middle and high parts of the plane XZ Y = 0 of the main flow. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
from the interaction between the diagonally upward main flow and the flow near the engine block. Finally, the central portion (window 1 in Fig. 8) of the plane Y = 0 is always characterized by upward diagonalization of flow with the appearance of interaction between the diagonal flow due to the fan and the reflected flow from the engine block. One can 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 region and to the attracting flow at the upper ends of the engine block. To characterize the topology of the secondary flow, six windows marked from 1 to 6 of the YZ plane located in the mid-distance from the engine block and the cooling module are considered in Fig. 9; three distinct zones are shown:
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– the first at the center is characterized by a large number of vortices distributed randomly in the plane (window 3 of Fig. 9), – two circular shaped areas (zones) of revolution axis parallel to that of the fan where the flow follows the same direction as the rotating fan blades (the dashed and solid arc lines passing through windows 1, 2, 4, 5 and 6 of Fig. 9), – the vortex area at the center generated by the interaction between the dead zone caused by the presence of the fan hub and the flow into two swirl circles around it. In conclusion, the flow resulting from a suction type fan has a very strong threedimensional structure, characterized by the appearance of particular structures in the main flow (eddies, jet air flow. . .) and a swirling flow creating a dead zone vortex behind the fan center. In particular, the topology of the main flow is likely
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to be significantly dependent on the blockage induced by the engine block on the fan flow. Thus, it is crucial to investigate the effect of variation of the distance d between the cooling module and the engine block on the flow induced by the fan. To study the engine blockage effect, the distance d between the cooling module and the engine block is varied. For this, the position of the cooling module is fixed while the engine block is incrementally moved backward: d ranges from 6 to 20 cm with 2 cm increments. This study has explicitly shown that the particular structures of the main flow observed in the configuration d = 6 cm (see the above description) are largely due to the proximity of the engine block downstream of the fan. Indeed, the air jet flow present in this position disappears for the longer distances (above 8 cm). Similarly, the different vortices that appeared for d = 6 cm (following the interaction between the fan airflow and the flow) which ‘‘bounced’’ back
Fig. 9. Swirling phenomenon in secondary flow – Averaged velocity fields obtained by PIV measurements in the plane YZ X = 3 cm. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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are not present for the other distances. Another typical difference is that from the distance of 8 cm on, a reverse flow appears in the central area, where negative longitudinal velocities are observed. Fig. 10 shows the distribution of the flow velocity component U in the Y = 0 plane, downstream of the fan for d = 6 cm, 8 cm and 10 cm. These measurements are obtained by PIV and for the fan functioning at 2800 rpm. The blue color corresponds to negative values of U. Thus, it is seen that the reverse flow appears from the distance beyond 8 cm in the measurement area. Effects of the distance between the radiator and the engine block are not noticeable on the topology of the secondary flow. The swirling phenomenon shown for d = 6 cm persists for the other distances, although there were slight changes in velocity components V and W. Since it appears that it is essentially the main flow which is sensitive to the engine blockage effect, in the following the study will be focussed on the flow velocity component U, aiming to link the changes in the flow with the thermal modifications: thermal power evacuated by heat exchangers, air temperature downstream of the heat exchangers and temperature at the engine block. Fig. 11 shows the distribution of the axial velocity component U obtained by LDV in a YZ plane located at X = 2 cm from the cooling module. The measurement planes shown are of 38 cm width (Y) and 32 cm height (Z) with 340 LDV measurement points. It can be noted that the flow topology in a YZ plane directly downstream of the fan is slightly modified by moving the engine block backward: – for d = 6 cm, traces of the fan blades on right of the picture are not yet present and those of left are not yet complete, – for d = 8 cm, the traces at left are completed while those at right start to appear in the lower part of the picture,
– for d = 10 cm, the traces of the fan blades at the top right begin to form and become complete for d = 14 cm, the distance at which the topology of the flow velocity becomes almost invariant. Traces of fan blades correspond to the parts of the picture that are in red–orange. The same features are observed for the other fan rotation speeds. However, while the flow topology is slightly modified by increasing the distance d, the flow statistics vary significantly. Fig. 12 shows the changes with the distance d in the average speed and the standard deviation of the flow velocity distribution in the YZ plane for X = 2 cm. It is noticed that by increasing the distance d: – the mean flow velocity increases, i.e. the air mass flow rate through the cooling module increases: at 2800 rpm, the air flow 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 a distance of 6 cm blocks the flow but makes it more homogeneous. By pushing backward the engine block from the radiator, the flow induced by the fan meets less resistance and the flow velocity, but the non-homogeneity increases too. The increase in the non-homogeneity of the velocity distribution through a heat exchanger decreases the capacity of the heat exchanger to evacuate heat (contrary to the increase in the mean speed). Moreover, even if the decrease in the engine blockage
Fig. 10. Axial flow velocity (U) distributions in the plane Y = 0 for different distances between the fan and the engine block.
M. Khaled et al. / Applied Energy 91 (2012) 439–450
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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 increase also the air temperature downstream of the fan which may penalize the cooling of the underhood bodies exposed to the
airflow. To quantify this thermal behavior, the outlet temperature of the radiator cooling water as well as air temperature distribution on a plane downstream of the fan are measured for a given
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4,50
P radiator (kW)
Air velocity (m/s)
2,70 2,50 2,30 2,10 1,90
Mean air velocity Standard deviation air velocity
1,70
4,00 3,50
Fan low speed Fan high hig gh speed
3,00 2,50 2,00 1,50 1,00
1,50 6
8
10
12
14
16
6
18
8
Distance (cm) (a)
10
12
14
16
18
Distance (cm) (a) 88,0 87,0
Fan low speed Fan high speed
86,0
T air (°C)
Air velocity (m/s)
5,40
4,90
4,40
84,0 83,0 82,0 81,0
Mean air velocity Standard deviation air velocity
3,90
85,0
80,0 79,0
3,40
6
6
8
10
12
14
16
8
18
radiator water inlet temperature (Fig. 5). These experiments were performed for different positions of the engine block and different water flow rates. Fig. 13 shows the variation of the radiator evacuated thermal power and the averaged air temperature at the fan downstream, as a function of the spacing between the cooling module and the engine block, for fan nominal speeds of 1400 and 2800 rpm. For the two fan operating speeds, it was found that the thermal power evacuated by the radiator increases as the distance d increases. For example, for d ranging from 6 to 18 cm, the power evacuated 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 decreases when d increases from 6 to 14 cm (18 cm for the case at 2800 rpm). For example, at a fan speed of 1400 rpm, the air temperature decreases from 87 °C to 79.5 °C, a relative decrease of 9%. From 14 cm, the air temperature downstream of the fan begins to increase slightly to attain 80.67 °C when d = 18 cm. This behavior reflects the competition between two opposite effects: by increasing the distance d, the evacuated power by the radiator increases and thus the heat absorbed by the air passing through the heat exchanger also increases. At the same time, the airflow rate through the heat exchanger increases. Between d = 6 and 14 cm the airflow increase dominates the absorbed thermal power, resulting in a decrease in the air temperature downstream of the fan. The opposite situation occurs for d higher than 14 cm. Table 2 reports the influence (in percentage) of the distance d on the thermal power evacuated by the radiator and the air temperature downstream of the fan.
12
14
16
18
Distance (cm) (b)
Distance (cm) (b) Fig. 12. Variations of the mean and the standard deviation of axial flow velocity field 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 are tested: (a) 1400 rpm and (b) 2800 rpm.
10
Fig. 13. Variations of the (a) radiator thermal power and (b) air temperature downstream of the fan, as a functions of the distance d for two nominal fan speeds. Water flow rate is 8 l/min.
Table 2 Percentages of deviation from the reference configuration (d = 6 cm) on the heat exchanger thermal power evacuation and air temperature downstream of the fan. Radiator thermal power
Downstream air temperature
Distance, d (cm)
Fan low speed
Fan high speed
Fan low speed
Fan high speed
8 10 12 14 16 18 20
17.75 26.09 29.96 34.78 36.52 36.56 37.84
7.06 12.85 16.14 18.26 19.37 19.93 22.00
1.24 4.91 6.65 8.58 7.85 7.29 8.55
0.52 1.73 2.93 4.43 5.16 4.58 3.74
It is noted that the increase in d causes useful variations in the different parameters. In the particular case of a Peugeot 207 (d = 6 cm) for example, if the engine block is moved backward by a distance of 2 cm the following results will be observed: – an increase in the thermal power evacuated by the radiator of 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 radiator of about 5% at low speed and 2% at high speed. Applicability of the above results depends, of course, on the availability of space in the underhood. Generally, additional 10– 12 cm is available in the underhood for the displacement of the
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engine block with respect to the cooling module. However, this space varies from one vehicle to another: for example in a Peugeot 207, approximately 6 cm additional space is available, while for a Citroën C6 this space varies between 10 and 12 cm approximately. Therefore, geometrically there is no potential restrictions if one respects the other allowable distances for an existing car design (for example Peugeot 207, 307, Citroen C2, C6, etc.). However, for a new car design, a choice has to be made if there is a strong downsizing demand of the underhood.
2
Cooling module
Automatic Command system
6 cm
1 Radiator inlet duct
2
Engine block
Water pump
X 4.2. Active control of vehicle cooling modules
(a) 2
Cooling module
Automatic Command system
10 cm
1 Radiator inlet duct
Engine block
X
2
Water pump
Control thermocouple
(b) 2
Cooling module 6 cm
In this section we describe a new monitoring tool [31] that can be used to optimize the underhood aerothermal management. It permits to adapt the heat exchangers’ performance to the engine energy requirements and to increase the heat exchangers’ heat evacuation power in critical situations such as the slowdown phase or the thermal soak phase of a vehicle (vehicle stop after a significant heating load). Motivated by the results of Section 4.1 on the aerothermal effects of the distance d between the cooling module and the engine block, the control scheme presented here is based on moving away (increase d) in the X-direction (direction of the vehicle length) the cooling module from the engine block in order to increase the heat exchangers evacuation power in critical situations where the engine overheats under conditions of vehicle slowdown or vehicle thermal soak. This control is managed by a thermocouple placed in the cooling water upstream of the radiator inlet (the engine outlet). Depending on the radiator inlet temperature in a given thermal situation during vehicle slowdown or thermal soak with the fan functioning phases, the heat exchangers energy requirements and the need to increase in heat evacuation power compared to the normal situation are determined and the cooling module is then moved to a distance that corresponds to the desired increase in power. To implement the control approach we apply it to the case of a simplified vehicle cooling module consisting only of a condenser and a radiator (as is the case in a Peugeot 207 without turbo-compressor for example). On other vehicle types, the same principles are applicable to other heat exchangers (such as the charged air cooler or engine oil cooler). The basic principle of the control scenario is described in Fig. 14a. In this Figure and for the sake of simplification, the control principle is presented only for the radiator; the same principle is applicable for the other heat exchangers (and is shown also for the condenser in Fig. 13c). The thermocouple placed at the radiator inlet returns the instantaneous temperature value to the automatic control system (step 1 marked 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 correspondence law between the critical temperatures Ti and the distances di between the cooling module and the engine block. Whenever one of the temperatures Ti is reached, the control system automatically commands the cooling module to move for a distance di from the engine block which corresponds to values recorded at the temperature Ti reached. It also controls the elongation of the connection pipes to the same distance di (both commands described correspond to the number 2 in black square of the technical description). Fig. 14b shows the system of Fig. 14a (reference configuration) after the application of the control. The active control consists therefore of the simultaneous application of the various commands described above (in Fig. 14a and c). In the case of several heat heat exchangers, other than the two main (radiator and condenser), the control is applied to the
Control thermocouple
Automatic Command system
1 Condenser inlet duct
Engine block X
2
Climate compressor
Control thermocouple
(c) Fig. 14. (a) Technical description of the first control principle for the radiator duct displacement (b) systems after application of the first control approach (c) technical description of the first control principle for the condenser duct displacement.
different heat exchanger pipes at the same time as the displacement in the X-direction of the cooling module. The economic interest of the proposed control approach can be seen in the application of its design to normal and non critical engine operation. If the control is used in normal vehicle operations, it can then increase the thermal power of the cooling module for the same work delivered for the water pump and the air conditioning system compressor. Thus, one can have the same evacuated thermal powers by the heat exchangers for smaller pumping and compressor works, which reduces the vehicle fuel consumption (since the pump and compressor works correspond to losses of the engine power). 5. Conclusions To allow the cooling module (several heat exchangers and one or two fans) of a vehicle to evacuate heat in difficult operating
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conditions, it has been pushed backward into the engine compartment. Therefore, the aeraulic and aerothermal situations in the underhood are largely conditioned by the cooling module which performance continues to present demanding aerothermal management challenges in their design as well as their interaction with other components in the underhood. In this work, experimental aerodynamic and thermal investigations were performed on a vehicle simplified body. These investigations have enabled us to study the flow induced by the fan and the blockage effect imposed by the engine to this flow, in conjunction with the thermal performance of real cooling module integrated in the model. In particular, we discuss, by the PIV and LDV measurements, the three-dimensional character of the fan flow. A first component of the threedimensional flow was the swirl phenomenon in planes parallel to the fan plane. A second corresponded to the main flow in the direction of the fan axis: properties of the main flow proved to be strongly related to the distance d between the cooling module and the engine block. For example, increasing the distance d, the air jet which appears for d = 6 cm does no longer exist. It is also found that by increasing the 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 monitoring tool that can be used for optimizing the underhood aerothermal management was proposed. It permits to adapt the heat exchangers’ performances to the engine energy requirements and to increase the heat exchangers’ evacuation power in critical situations such as the slowdown phase or the thermal soak phase of a vehicle (vehicle stops after a significant heating load). The aim here is to apply this new control approach to a real vehicle so as to reduce the energy delivered to the pump and compressor functions, and therefore to reduce the vehicle fuel consumption. References [1] Xie G, Sunden B, Wang Q, Tang L. Performance predictions of laminar and turbulent heat transfer and fluid flow of heat exchangers having large tubediameter and large tube-row by artificial neural networks. Int J Heat Mass Transfer 2009;52:2484–97. [2] Hasan MI, Rageb AA, Yaghoubi M, Homayoni H. Influence of channel geometry on the performance of a counter flow microchannel heat exchanger. Int J Thermal Sci 2009. [3] Salimpour MR. Heat transfer coefficients of shell and coiled tube heat exchangers. Exp Therm Fluid Sci 2009;33:203–7. [4] Ajakh A, Kestoras MD, Toe R, Peerhossaini H. Influence of forced perturbations in the stagnation region on Görtler instability. AIAA J 1999;37:1572–7. [5] Zhang P, Hrnjak PS. Air-side performance evaluation of three types of heat exchangers in dry, wet and periodic frosting conditions. Int J Refrig 2009. [6] Lasbet Y, Auvity B, Castelain C, Peerhossaini H. A chaotic heat-exchanger for PEMFC cooling applications. J Power Sources 2006;156:114–8. [7] Peerhossaini H, Castelain C, Le Guer Y. Heat exchanger design based on chaotic advection. Exp Therm Fluid Sci 1993;7:333–44. [8] Habchi C, Lemenand T, Della Valle D, Peerhossaini H. Liquid–liquid dispersion in a chaotic advection flow. Int J Multiphase Flow 2009;35:485–97. [9] Habchi C, Ouarets S, Lemenand T, Della Valle D, Bellettre J, Peerhossaini H. Influence of viscosity ratio on droplets formation in a chaotic advection flow. Int J Chem Reactor Eng 2009;7:A50.
[10] Lai Y, Cordero N, Barthel F, Tebbe F, Kuhn J, Apfelbeck R, et al. Liquid cooling of bright LEDs for automotive applications. Appl Therm Eng 2009;29:1239–44. [11] Lawson MJ, Thole KA. Heat transfer augmentation along the tube wall of a louvered fin heat exchanger using practical delta winglets. Int J Heat Mass Transfer 2008;51:2346–60. [12] Ferouillat S, Tochon P, Garnier C, Peerhossaini H. Intensification of heattransfer and mixing in multifunctional heat exchangers by artificially generated streamwise vorticity. Appl Therm Eng 2006;26:1820–9. [13] Ferouillat S, Tochon P, Peerhossaini H. Open-loop thermal control of exothermal chemical reactions in multifunctional heat exchangers. Int J Heat Mass Transfer 2006;26:2479–90. [14] Doodman AR, Fesanghray M, Hosseini R. A robust stochastic approach for design optimization of air cooled heat exchangers. Appl Energy 2009;86:1240–5. [15] Guo J, Xu M, Cheng L. The application of field synergy number in shell-andtube heat exchanger optimization design. Appl Energy 2009;86:2079–87. [16] Ferrouillat S, Tochon P, Peerhossaini H. Micromixing enhancement by turbulence. application to multifunctional heat exchangers. Chem Eng Process 2006;45:633–40. [17] Gholap AK, Khan JA. Design and multi-objective optimisation of heat exchangers for refrigerators. Appl Energy 2007;84:1226–39. [18] Hilbert R, Janiga G, Baron R, Thévenin D. Multi-objective shape optimization of a heat exchanger using parallel genetic algorithms. Int J Heat Mass Transfer 2006;49:2567–77. [19] Foli K, Okabe T, Olhofer M, Jin Y, Sendhoff B. Optimization of micro heat exchanger: CFD, analytical approach and multi-objective evolutionary algorithms. Int J Heat Mass Transfer 2006;49:1090–9. [20] Habchi C, Lemenand T, Della Valle D, Peerhossaini H. Alternating mixing tabs in multifunctional heat exchanger-reactor. Chem Eng Process 2010;49:653–61. [21] Oliet C, Oliva A, Castro J, Pérez-Segarra CD. Parametric studies on automotive radiators. Appl Therm Eng 2007;27:2033–43. [22] Qi ZG, Chen JP, Chen ZJ. Parametric study on the performance of a heat exchanger with corrugated louvered fins. Appl Therm Eng 2007;27:539–44. [23] Witry A, Al-Hajeri MH, Bondok AA. Thermal performance of automotive aluminium plate radiator. Appl Therm Eng 2005;25:1207–18. [24] Carluccio E, Starace G, Ficarella A, Laforgia D. Numerical analysis of a crossflow compact heat exchanger for vehicle applications. Appl Therm Eng 2005;25:1995–2013. [25] Khaled M, Harambat F, Peerhossaini H. Analytical and empirical determination of thermal performance of louvered heat exchanger-Effects of air flow statistics. Int J Heat Mass Transfer 2011;54:356–65. [26] Khaled M, Harambat F, Peerhossaini H. A quantitative method for assessment of car inclination effects on thermal management of the underhood compartment, Journal of Thermal Science and Engineering Applications – Transaction of ASME, March 2009, vol. 1/014501-5. [27] Khaled M, Harambat F, Peerhossaini H. Effects of car inclination on air flow and aerothermal behaviour in the underhood compartment. In: Proceedings of the ASME Fluids Engineering Division Summer Conference, vol. 1, PT A and B; 2008. p. 517–22. [28] Khaled M, Harambat F, Peerhossaini H. Temperature and heat flux behavior of complex flows in car underhood compartment. Heat Transfer Eng 2010;31(13):1–11. [29] Khaled M, Harambat F, Peerhossaini H. Underhood thermal management: temperature and flux measurements and physical analysis. Appl Therm Eng 2010;30:590–8. [30] Khaled M, Garnier B, Harambat F, Peerhossaini H. A new method for simultaneous measurement of convective and radiative heat flux in car underhood applications. Measur Sci Technol J 2010;21:025903. 10pp. [31] Khaled M, Yammine A, Harambat F, Peerhossaini H. Aerodynamic forces on a simplified car body – towards innovative designs for car drag reduction. In: Proceedings of the ASME fluids engineering division summer conference, August 1–5, Montréal, Canada; 2010. FEDSM-ICNMM2010-30321. [32] Khaled M, Harambat F, Peerhossaini H. Véhicule Automobile Présentant un Refroidissement optimisé, French patent application filed on January 22 2009, Patent no. 1050417. [33] Khaled M, Garnier B, Harambat F, Peerhossaini H. Towards the control of car underhood thermal conditions. Appl Therm Eng 2011;31:902–10. [34] Khaled M, Harambat F, El Hage H, Peerhossaini H. Spatial optimization of an underhood cooling module: towards an innovative control approach. Appl Energy 2011;88:3841–9.
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Fan air flow analysis and heat transfer enhancement of vehicle underhood cooling 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, France PSA Peugeot Citroën, Velizy A Center, 2 route de Gisy, 78 943 Vélizy Villacoublay, France c Fluid Mechanics, Heat and Thermodynamics Group, School of Engineering, Lebanese International University, Beirut, Lebanon b
a r t i c l e
i n f o
Article history: Received 12 August 2011 Received in revised form 9 October 2011 Accepted 12 October 2011 Available online 9 November 2011 Keywords: Vehicle underhood Car cooling module Heat exchangers Heat transfer enhancement Fan Car underhood aerothermal management
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 positioning according to the engine energy requirements. Measurements are carried out on a simplified vehicle body designed based on the real vehicle front block. We report here velocity and temperature measurements by Particle Image Velocimetry (PIV), by Laser Doppler Velocimetry (LDV) and by thermocouples. The underhood of the simplified body is instrumented by 59 surface and fluid thermocouples. Measurements are carried out for conditions simulating both the slowdown and the thermal soak phases with the fan in operation. Different fan rotational speeds, radiator water flow and underhood geometries have been experimented. The ultimate aim is to apply the new control approach to a real vehicle so as to reduce the energy delivered to the pump and compressor and therefore to reduce the vehicle fuel consumption. Ó 2011 Elsevier Ltd. All rights reserved.
1. Introduction The cooling module of vehicle constituted of an assembly of different heat exchangers and one or two fans have been pushed backward into the engine compartment, in order to evacuate heat in difficult operating conditions. This have caused the underhood aerothermal and aeraulic conditions to be largely conditioned by the different elements of the cooling module which at the same time interact with other components in the underhood. In heat exchangers’ applications which cover a large variety of technological domains [1–10] (such as nuclear reactors, chemical and biomedical processes, military and aerospace, and automotive), the trend is to reduce the heat exchangers’ weight and volume while keeping sufficient or even increasing thermal efficiencies. In the open literature, the large part of optimization studies is focused on the geometry and the design of the heat exchangers and also on the fluid flows structures. Several design and control approaches aiming to optimize the heat exchangers’ thermal performance are reported in [11–20], while few studies are focussed on the heat exchangers’ optimization taking into account their interaction with the geometry in which they are installed. Among the most commonly heat exchangers used are the tubes and fins heat exchangers which are composed of several tubes of elliptical cross section between which are positioned several ⇑ Corresponding author. Tel.: +33 2 40 68 31 24; fax: +33 2 40 68 31 41. E-mail address: [email protected] (H. Peerhossaini). 0306-2619/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.apenergy.2011.10.017
continuous and parallel fins. These types of heat exchangers are particularly adapted to most industrial applications, especially in automotive applications, and their fins can be simple, louvered, crimped, shutters, or with delta wings longitudinal vortex generators, etc. Several studies [21–26] on the louvered fin heat exchangers for underhood applications showed that the thermal performance of such exchangers are strongly affected by the geometrical parameters of the heat exchanger (e.g. distance between fins, the tubes height, and fins angle) and operating conditions of the two fluids, air and water (their mass flow rates and inlet temperatures). However, few studies have explicitly addressed the crucial question of how to control the different parameters for enhancing the exchanger’s heat evacuation capacity. Most of the underhood heat transfer analysis has been concentrated on the underhood thermal management by temperature and heat flux measurements [27–34]. In the present study we have experimentally investigated the heat exchanger optimization by acting on the geometry (underhood space here) in which they are integrated. The present analysis shows how to act on the different parameters which influence the heat exchanger performance in order to optimize its functioning. On the other hand, in critical phases of vehicle functioning such as cooling at the vehicle stopped situation, the stop and go driving mode, and the vehicle thermal soak, the performance of car heat exchanger is essentially influenced by the air flow generated by the fan. This air flow in its turn depends on the underhood architecture. Most experimental 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.
determination of the fan aerothermal characteristics (curves of pressure rise as a function of the air flow) or optimization studies (use of rectifiers in order to guide the airflow) where the fan alone is considered. On real cars, these studies are quite limited given the constraints of accessibility related to the presence of engine block downstream of the fan. For instance, it is very difficult to use optical measurement techniques such as PIV or LDV in the car underhood space. One solution is to work on underhood simplified models designed to allow the integration of optical measurement techniques. Compared to studies on isolated fan, these models present the advantage of reproducing the effects of the engine blocking imposed on the fan flow. This ‘‘blocking’’ due to the short distance between the fan and the engine block downstream, is a parameter that could affect the characteristics of the flow induced by the fan in the car underhood. Moreover, the space between the cooling module and the engine block varies from one vehicle type to another. For these
reasons, it was decided in the present study to work on a car underhood model containing an actual cooling module and simplified engine block both translatable in X and Y (respectively in the vehicle length and width) directions. This modularity allows covering a wide range of underhood architectures to explore different strategies adapted by the automobile manufacturers. Part of this paper will thus be focussed on the PIV and LDV measurements as well as thermal measurements performed on the model for different blockages downstream of the fan, that is to say, for different positions of the engine block with respect to the cooling system. The impact of these aerodynamic changes on the thermal performance of the heat exchangers and their surrounding area is also studied. We report here velocity and temperature measurements by Particle Image Velocimetry PIV, by Laser Doppler Velocimetry and by thermocouples. The underhood of the simplified car body is instrumented by 59 surface and fluid thermocouples. Measurements are
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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.
carried out for conditions simulating both the slowdown and the thermal soak phases with the fan in operation. Different fan rotational speeds, radiator water flow and underhood geometries have been tested. The final aim is to apply the new control approach developed here on a real vehicle so as to reduce the energy delivered to the pump and compressor functions and thus to reduce the vehicle fuel consumptions. The rest of this paper is organized as follows. Section 2 describes the underhood simplified body designed for the experimental investigations. In Section 3, measurements carried out on the simplified model by PIV, LDV and thermocouples are reported. Section 4 presents the experimental results on the heat exchangers thermal performance and the fan air flow topology as well as a new control approach permitting the adaptation of the cooling module of a given vehicle to its engine energy requirements.
geometries and sizes of air inlets used in vehicle series, by closing or opening a given and specific number of windows. The car simplified body is provided with several types of air outlets: a vertical exit, a tunnel passage (representing the outlet in the tunnel exhaust of a real vehicle), openings in the wheel arches and outlets on the underbody. In the experiments, not all the outlets are open simultaneously; combinations between different types of openings permit to simulate configurations of different car series: for example when the vertical outlet and outlets in the underbody are closed by leaving open outlets in the tunnel and the wheel arches, one can simulate the configurations of conventional vehicles series of PSA Peugeot Citroen. On the other hand, by keeping only the exit in the tunnel exhaust open, typical configurations of BMW vehicles are represented.
2. Car simplified body The car simplified body is described in [30], here we present some main features of it. The considered model is a simplified but representative form of the front end block of a Peugeot 207 passenger car (Fig. 1a). The dimensions of the front end projection are 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 at its 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 allows varying their positions in the X and Y directions (Fig. 1c). The initial configuration is that of a Peugeot 207 where the cooling module is offset from the center of the car to the right (as Y) and where the engine block is centered in the width (Y) of the car and is positioned at 6 cm from the cooling module. The engine block can then move from 6 cm to 6 cm in the Y direction in steps of 2 cm; and 6 cm to 20 cm from the cooling module in the X following steps of 2 cm. The air inlet of the model (Fig. 1d) is flat and flexible: with a fixed frame and several windows, it can represent different
Traps
Inlet Radiator
Outlet Flowmeter
Supply Main water supply
Drain Boiler with integrated pump
Fig. 4. Schematic diagram of the thermal loop used to measure the thermal power evacuation of radiator.
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Fig. 5. Grid of air thermocouples placed downstream of the cooling module.
3. Experimental setup
Exchangers
Aerodynamic and thermal measurements are carried out on the car simplified body. Aerodynamic measurements are carried out by Laser Doppler Velocimetry (LDV) (Section 3.1) and Particle Image Velocimetry (PIV) (Section 3.2) and temperatures are measured using thermocouples (Section 3.3).
Fan
Engine block
6 cm Air
3.1. LDV measurements For LDV measurements, the model is adjusted in a manner to reproduce the typical air inlets and outlets of a Peugeot 207 passenger car: the air inlets are of identical dimensions as that of the Peugeot 207, the air outlets are in the exhaust tunnel and in the wheel arches; the ratio between inlet and outlet sections is about 0.6. Paraffin oil (seeding) is added to air flow in front of the model (upstream of the fan) and the LDV probe is mounted on a three-dimensional traversing system and placed next to the model (Fig. 2). Two fan rotational speeds are studied: 1400 rpm and 2800 rpm. They are both the nominal operating speeds on actual Peugeot 207 vehicle. 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 spacing configurations are obtained by keeping the cooling module fixed and moving backward the engine block in increments of 2 cm in the X direction (car movement direction), the initial spacing being 6 cm. For each configuration (given fan rotation speed and radiator-engine block spacing), the horizontal velocity component U and vertical velocity component W (in the YZ plane) are measured at X = 1 cm downstream of the fan.
Z Y
Measurement window X Vortex
3.2. PIV measurements Contrary to a blower fan (located upstream of the heat exchangers), the air flow distribution induced by a suction fan (as is the case in the simplified body) in the underhood is no longer Table 1 Important parameters varied in the 48 different configurations examined in this work. Water flow rate (l/min)
Fan speed (rpm)
Cooling module/engine block spacing (cm)
8 and 10
1400, 2800 and 3300
6, 8, 10, 12, 14, 16, 18, 20
Impacting air jet Fig. 6. Averaged velocity field in a window of the lower part of the plane XZ Y = 0 of the main flow. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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conditioned by the cooling module. In this case, a swirling airflow can be generated and a secondary flow appears in addition to the main flow, giving rise to a three-dimensional structure to the resulting air flow. In order to obtain a three-dimensional picture of the resulting flows, we study first the topology of the primary and the secondary flows by PIV measurements. The main flow (XZ plane) is in a plane normal to the fan plane (Fig. 3a) and the secondary flow (YZ plane) is in a vertical plane parallel to the fan plane (Fig. 3b). By fixing the laser source inside the engine block (Fig. 3c), the PIV camera can be put on either the right or left side of the simplified body (Fig. 3d) to visualize the flow downstream of the fan. Measurements are made for different fan angular speeds (including those corresponding to functioning in a real vehicle) and different spaces between the cooling module and the engine block. Then, through several windows in the measurement plane and moving them in Y direction (i.e. on the width of the model), one can obtain an appropriate description of the main air flow between the fan and the engine block. For the topology of the secondary flow (detection of the swirling phenomenon), positions of the camera and the laser source are permuted. Positioned on the right side or left side of the model, the laser illuminates a plane parallel to the fan plane and the camera takes pictures of this plan within the engine block. For several positions of the camera with respect to the laser light plane, several windows of secondary flow were obtained. With these different windows, we covered an area large enough to probe the swirling flow. Note that other laser/ camera positions can be applied (e.g. the laser above the model and the camera below the model. . .), but we always chose the positions with less laser light attenuation and the sharpest images.
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3.3. Thermal measurements The purpose of such measures is to study the radiator performance for different fan speeds and relative engine block positions (including the one which corresponds to a real underhood of a Peugeot 207), that is the relation between the underhood aerodynamic and the heat exchangers and their environments. The thermal loop used for the measurement of radiator efficiency is shown in Fig. 4. The radiator is fed by a water circuit that controls the inlet temperature. The circuit essentially comprises a variable power boiler, the radiator of the cooling model (a real Peugeot 207 radiator), a main water supply, a flowmeter, valves and traps. The boiler has a three speeds (8, 10 and 12 l/min) built in pump and a temperature probe which allows to fix the hot temperature (the temperature at the inlet of the radiator) following instruction requested. Two 80 lm bead diameter thermocouples (type K) are positioned at the inlet and the outlet of the radiator. A grid of 30 thermocouples (80-micronsdiameters type K) is placed downstream of the cooling module (Fig. 5) which allows measuring the air temperature distribution on a plane downstream of the cooling module. The thermocouple grid was placed on a plane located 3 cm (in the X-direction) downstream of the cooling module and parallel to its surface. The thermocouples are distributed equally spaced in both X and Y directions. To ensure the reliability of the temperature measurement by the thermocouples grid with respect to vibration and deformation, we fixed the thermocouples on a grid made of very thin and rigid wires. The wire diameter was chosen in such a way that the Reynolds number for generation of Karman vortices, which are the cause of wire vibration and deformation, was under critical. Reproducibility tests were carried out and showed satisfactory results. The engine block surface opposite to the cooling module is instrumented with 27 type T thermocouples of 1 mm diameter
Fig. 7. Temperature increase (heat transfer enhancement) at the low part of the engine block in relation with the air jet flow apparition (plane XZ Y = 0). Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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beads, each in order to obtain the temperature distribution on the engine block surface. Three fan speeds are tested: the two nominal speeds of 1400 and 2800 rpm and a maximum speed of 3300 rpm. Two different water flow rates are imposed: 8 l/min and 10 l/min. Finally, eight configurations for positioning of the engine block in the X-direction with respect to the cooling module are tested: 6, 8, 10, 12, 14, 16, 18 and 20 cm. For each configuration, water flow, fan speed, position of the engine block, the radiator inlet temperature, and the temperatures at the different locations described above are recorded (59 locations). Table 1 summarizes the different parameters varied during the experiments (aerodynamic and thermal measurements), all corresponding to an inlet water temperature of 60 °C. A total of 2 3 8 = 48 configurations were examined. 4. Results and discussion Results presented here concern the heat transfer enhancement of the cooling module of a vehicle as well as the air flow induced by the fan and the aerothermal effects of the engine blockage on this air flow (Section 4.1). These results are then coupled with a scheme for active control of a vehicle cooling module (Section 4.2). 4.1. Fan airflow and engine blockage effects In experiments, ‘‘suction’’ type fan is placed behind the heat exchangers. Contrary to a ‘‘blowing’’ fan, which is located upstream of the heat exchangers, flow induced by a suction fan is likely to generate a three-dimensional flow with a large and strong swirl in the YZ plane. On the other hand, the presence of the engine block at only 6 cm downstream of the fan makes the topology of the air flow complex. Fig. 6 shows the velocity fields measured by PIV in the mid-height of radiator (at Y = 0) for a fan speed of 2800 rpm. A downward diagonal air flow is noticed in this measurement window. The diagonal direction of the flow is due to the difference in pressure between the region directly downstream of the fan and that of depression generated by the geometric restrictions below the engine block. On the other hand, a jet-type flow appears in the right part of the velocity field due to the air impact on the engine block. Reflected by the engine block, the jet flow is deflected in a direction diagonally opposite to the direction of the main flow with which it interacts (see the upper part of the air jet). The combination between the descending main flow and the reverse ascending jet generates a vortex midway from the engine block and the cooling module. The jet flow at the lower engine block region promotes the convective heat transfer in this part. Fig. 7 shows the temperature distribution in the jet area and the temperature profile over the total height of the engine block. Measurements are performed for the radiator fed with hot water at 60 °C. The air heated through the heat exchanger passage increases the temperature of the engine block (which in the simplified model is not a source of heat). A warmer area in the central part of the engine-block with a temperature peak located at the jet impact position is noticed. Going up in the Z direction (in the direction of the simplified model height), an upward vertical flow appears clearly in the middle and upper zones between the cooling module and the engine block (Fig. 8). This progressive emergence of a vertical velocity component is mainly due to the abrupt return flow that creates the engine-block near the fan and the depression prevalence generated at the engineblock and above. In Fig. 8, it is noticed that the air jet flow appeared in the lower part is not reproduced in the upper part. However, a vortex is detected at the lower part of the velocity field. This vortex results
Fig. 8. Averaged velocity fields in windows at the middle and high parts of the plane XZ Y = 0 of the main flow. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
from the interaction between the diagonally upward main flow and the flow near the engine block. Finally, the central portion (window 1 in Fig. 8) of the plane Y = 0 is always characterized by upward diagonalization of flow with the appearance of interaction between the diagonal flow due to the fan and the reflected flow from the engine block. One can 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 region and to the attracting flow at the upper ends of the engine block. To characterize the topology of the secondary flow, six windows marked from 1 to 6 of the YZ plane located in the mid-distance from the engine block and the cooling module are considered in Fig. 9; three distinct zones are shown:
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– the first at the center is characterized by a large number of vortices distributed randomly in the plane (window 3 of Fig. 9), – two circular shaped areas (zones) of revolution axis parallel to that of the fan where the flow follows the same direction as the rotating fan blades (the dashed and solid arc lines passing through windows 1, 2, 4, 5 and 6 of Fig. 9), – the vortex area at the center generated by the interaction between the dead zone caused by the presence of the fan hub and the flow into two swirl circles around it. In conclusion, the flow resulting from a suction type fan has a very strong threedimensional structure, characterized by the appearance of particular structures in the main flow (eddies, jet air flow. . .) and a swirling flow creating a dead zone vortex behind the fan center. In particular, the topology of the main flow is likely
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to be significantly dependent on the blockage induced by the engine block on the fan flow. Thus, it is crucial to investigate the effect of variation of the distance d between the cooling module and the engine block on the flow induced by the fan. To study the engine blockage effect, the distance d between the cooling module and the engine block is varied. For this, the position of the cooling module is fixed while the engine block is incrementally moved backward: d ranges from 6 to 20 cm with 2 cm increments. This study has explicitly shown that the particular structures of the main flow observed in the configuration d = 6 cm (see the above description) are largely due to the proximity of the engine block downstream of the fan. Indeed, the air jet flow present in this position disappears for the longer distances (above 8 cm). Similarly, the different vortices that appeared for d = 6 cm (following the interaction between the fan airflow and the flow) which ‘‘bounced’’ back
Fig. 9. Swirling phenomenon in secondary flow – Averaged velocity fields obtained by PIV measurements in the plane YZ X = 3 cm. Velocity fields presented here are obtained by PIV. Fan speed is 2800 rpm, distance d is 6 cm, water flow rate and inlet temperature are 8 l/min and 60 °C respectively.
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are not present for the other distances. Another typical difference is that from the distance of 8 cm on, a reverse flow appears in the central area, where negative longitudinal velocities are observed. Fig. 10 shows the distribution of the flow velocity component U in the Y = 0 plane, downstream of the fan for d = 6 cm, 8 cm and 10 cm. These measurements are obtained by PIV and for the fan functioning at 2800 rpm. The blue color corresponds to negative values of U. Thus, it is seen that the reverse flow appears from the distance beyond 8 cm in the measurement area. Effects of the distance between the radiator and the engine block are not noticeable on the topology of the secondary flow. The swirling phenomenon shown for d = 6 cm persists for the other distances, although there were slight changes in velocity components V and W. Since it appears that it is essentially the main flow which is sensitive to the engine blockage effect, in the following the study will be focussed on the flow velocity component U, aiming to link the changes in the flow with the thermal modifications: thermal power evacuated by heat exchangers, air temperature downstream of the heat exchangers and temperature at the engine block. Fig. 11 shows the distribution of the axial velocity component U obtained by LDV in a YZ plane located at X = 2 cm from the cooling module. The measurement planes shown are of 38 cm width (Y) and 32 cm height (Z) with 340 LDV measurement points. It can be noted that the flow topology in a YZ plane directly downstream of the fan is slightly modified by moving the engine block backward: – for d = 6 cm, traces of the fan blades on right of the picture are not yet present and those of left are not yet complete, – for d = 8 cm, the traces at left are completed while those at right start to appear in the lower part of the picture,
– for d = 10 cm, the traces of the fan blades at the top right begin to form and become complete for d = 14 cm, the distance at which the topology of the flow velocity becomes almost invariant. Traces of fan blades correspond to the parts of the picture that are in red–orange. The same features are observed for the other fan rotation speeds. However, while the flow topology is slightly modified by increasing the distance d, the flow statistics vary significantly. Fig. 12 shows the changes with the distance d in the average speed and the standard deviation of the flow velocity distribution in the YZ plane for X = 2 cm. It is noticed that by increasing the distance d: – the mean flow velocity increases, i.e. the air mass flow rate through the cooling module increases: at 2800 rpm, the air flow 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 a distance of 6 cm blocks the flow but makes it more homogeneous. By pushing backward the engine block from the radiator, the flow induced by the fan meets less resistance and the flow velocity, but the non-homogeneity increases too. The increase in the non-homogeneity of the velocity distribution through a heat exchanger decreases the capacity of the heat exchanger to evacuate heat (contrary to the increase in the mean speed). Moreover, even if the decrease in the engine blockage
Fig. 10. Axial flow velocity (U) distributions in the plane Y = 0 for different distances between the fan and the engine block.
M. Khaled et al. / Applied Energy 91 (2012) 439–450
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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 increase also the air temperature downstream of the fan which may penalize the cooling of the underhood bodies exposed to the
airflow. To quantify this thermal behavior, the outlet temperature of the radiator cooling water as well as air temperature distribution on a plane downstream of the fan are measured for a given
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4,50
P radiator (kW)
Air velocity (m/s)
2,70 2,50 2,30 2,10 1,90
Mean air velocity Standard deviation air velocity
1,70
4,00 3,50
Fan low speed Fan high hig gh speed
3,00 2,50 2,00 1,50 1,00
1,50 6
8
10
12
14
16
6
18
8
Distance (cm) (a)
10
12
14
16
18
Distance (cm) (a) 88,0 87,0
Fan low speed Fan high speed
86,0
T air (°C)
Air velocity (m/s)
5,40
4,90
4,40
84,0 83,0 82,0 81,0
Mean air velocity Standard deviation air velocity
3,90
85,0
80,0 79,0
3,40
6
6
8
10
12
14
16
8
18
radiator water inlet temperature (Fig. 5). These experiments were performed for different positions of the engine block and different water flow rates. Fig. 13 shows the variation of the radiator evacuated thermal power and the averaged air temperature at the fan downstream, as a function of the spacing between the cooling module and the engine block, for fan nominal speeds of 1400 and 2800 rpm. For the two fan operating speeds, it was found that the thermal power evacuated by the radiator increases as the distance d increases. For example, for d ranging from 6 to 18 cm, the power evacuated 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 decreases when d increases from 6 to 14 cm (18 cm for the case at 2800 rpm). For example, at a fan speed of 1400 rpm, the air temperature decreases from 87 °C to 79.5 °C, a relative decrease of 9%. From 14 cm, the air temperature downstream of the fan begins to increase slightly to attain 80.67 °C when d = 18 cm. This behavior reflects the competition between two opposite effects: by increasing the distance d, the evacuated power by the radiator increases and thus the heat absorbed by the air passing through the heat exchanger also increases. At the same time, the airflow rate through the heat exchanger increases. Between d = 6 and 14 cm the airflow increase dominates the absorbed thermal power, resulting in a decrease in the air temperature downstream of the fan. The opposite situation occurs for d higher than 14 cm. Table 2 reports the influence (in percentage) of the distance d on the thermal power evacuated by the radiator and the air temperature downstream of the fan.
12
14
16
18
Distance (cm) (b)
Distance (cm) (b) Fig. 12. Variations of the mean and the standard deviation of axial flow velocity field 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 are tested: (a) 1400 rpm and (b) 2800 rpm.
10
Fig. 13. Variations of the (a) radiator thermal power and (b) air temperature downstream of the fan, as a functions of the distance d for two nominal fan speeds. Water flow rate is 8 l/min.
Table 2 Percentages of deviation from the reference configuration (d = 6 cm) on the heat exchanger thermal power evacuation and air temperature downstream of the fan. Radiator thermal power
Downstream air temperature
Distance, d (cm)
Fan low speed
Fan high speed
Fan low speed
Fan high speed
8 10 12 14 16 18 20
17.75 26.09 29.96 34.78 36.52 36.56 37.84
7.06 12.85 16.14 18.26 19.37 19.93 22.00
1.24 4.91 6.65 8.58 7.85 7.29 8.55
0.52 1.73 2.93 4.43 5.16 4.58 3.74
It is noted that the increase in d causes useful variations in the different parameters. In the particular case of a Peugeot 207 (d = 6 cm) for example, if the engine block is moved backward by a distance of 2 cm the following results will be observed: – an increase in the thermal power evacuated by the radiator of 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 radiator of about 5% at low speed and 2% at high speed. Applicability of the above results depends, of course, on the availability of space in the underhood. Generally, additional 10– 12 cm is available in the underhood for the displacement of the
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engine block with respect to the cooling module. However, this space varies from one vehicle to another: for example in a Peugeot 207, approximately 6 cm additional space is available, while for a Citroën C6 this space varies between 10 and 12 cm approximately. Therefore, geometrically there is no potential restrictions if one respects the other allowable distances for an existing car design (for example Peugeot 207, 307, Citroen C2, C6, etc.). However, for a new car design, a choice has to be made if there is a strong downsizing demand of the underhood.
2
Cooling module
Automatic Command system
6 cm
1 Radiator inlet duct
2
Engine block
Water pump
X 4.2. Active control of vehicle cooling modules
(a) 2
Cooling module
Automatic Command system
10 cm
1 Radiator inlet duct
Engine block
X
2
Water pump
Control thermocouple
(b) 2
Cooling module 6 cm
In this section we describe a new monitoring tool [31] that can be used to optimize the underhood aerothermal management. It permits to adapt the heat exchangers’ performance to the engine energy requirements and to increase the heat exchangers’ heat evacuation power in critical situations such as the slowdown phase or the thermal soak phase of a vehicle (vehicle stop after a significant heating load). Motivated by the results of Section 4.1 on the aerothermal effects of the distance d between the cooling module and the engine block, the control scheme presented here is based on moving away (increase d) in the X-direction (direction of the vehicle length) the cooling module from the engine block in order to increase the heat exchangers evacuation power in critical situations where the engine overheats under conditions of vehicle slowdown or vehicle thermal soak. This control is managed by a thermocouple placed in the cooling water upstream of the radiator inlet (the engine outlet). Depending on the radiator inlet temperature in a given thermal situation during vehicle slowdown or thermal soak with the fan functioning phases, the heat exchangers energy requirements and the need to increase in heat evacuation power compared to the normal situation are determined and the cooling module is then moved to a distance that corresponds to the desired increase in power. To implement the control approach we apply it to the case of a simplified vehicle cooling module consisting only of a condenser and a radiator (as is the case in a Peugeot 207 without turbo-compressor for example). On other vehicle types, the same principles are applicable to other heat exchangers (such as the charged air cooler or engine oil cooler). The basic principle of the control scenario is described in Fig. 14a. In this Figure and for the sake of simplification, the control principle is presented only for the radiator; the same principle is applicable for the other heat exchangers (and is shown also for the condenser in Fig. 13c). The thermocouple placed at the radiator inlet returns the instantaneous temperature value to the automatic control system (step 1 marked 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 correspondence law between the critical temperatures Ti and the distances di between the cooling module and the engine block. Whenever one of the temperatures Ti is reached, the control system automatically commands the cooling module to move for a distance di from the engine block which corresponds to values recorded at the temperature Ti reached. It also controls the elongation of the connection pipes to the same distance di (both commands described correspond to the number 2 in black square of the technical description). Fig. 14b shows the system of Fig. 14a (reference configuration) after the application of the control. The active control consists therefore of the simultaneous application of the various commands described above (in Fig. 14a and c). In the case of several heat heat exchangers, other than the two main (radiator and condenser), the control is applied to the
Control thermocouple
Automatic Command system
1 Condenser inlet duct
Engine block X
2
Climate compressor
Control thermocouple
(c) Fig. 14. (a) Technical description of the first control principle for the radiator duct displacement (b) systems after application of the first control approach (c) technical description of the first control principle for the condenser duct displacement.
different heat exchanger pipes at the same time as the displacement in the X-direction of the cooling module. The economic interest of the proposed control approach can be seen in the application of its design to normal and non critical engine operation. If the control is used in normal vehicle operations, it can then increase the thermal power of the cooling module for the same work delivered for the water pump and the air conditioning system compressor. Thus, one can have the same evacuated thermal powers by the heat exchangers for smaller pumping and compressor works, which reduces the vehicle fuel consumption (since the pump and compressor works correspond to losses of the engine power). 5. Conclusions To allow the cooling module (several heat exchangers and one or two fans) of a vehicle to evacuate heat in difficult operating
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conditions, it has been pushed backward into the engine compartment. Therefore, the aeraulic and aerothermal situations in the underhood are largely conditioned by the cooling module which performance continues to present demanding aerothermal management challenges in their design as well as their interaction with other components in the underhood. In this work, experimental aerodynamic and thermal investigations were performed on a vehicle simplified body. These investigations have enabled us to study the flow induced by the fan and the blockage effect imposed by the engine to this flow, in conjunction with the thermal performance of real cooling module integrated in the model. In particular, we discuss, by the PIV and LDV measurements, the three-dimensional character of the fan flow. A first component of the threedimensional flow was the swirl phenomenon in planes parallel to the fan plane. A second corresponded to the main flow in the direction of the fan axis: properties of the main flow proved to be strongly related to the distance d between the cooling module and the engine block. For example, increasing the distance d, the air jet which appears for d = 6 cm does no longer exist. It is also found that by increasing the 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 monitoring tool that can be used for optimizing the underhood aerothermal management was proposed. It permits to adapt the heat exchangers’ performances to the engine energy requirements and to increase the heat exchangers’ evacuation power in critical situations such as the slowdown phase or the thermal soak phase of a vehicle (vehicle stops after a significant heating load). The aim here is to apply this new control approach to a real vehicle so as to reduce the energy delivered to the pump and compressor functions, and therefore to reduce the vehicle fuel consumption. References [1] Xie G, Sunden B, Wang Q, Tang L. Performance predictions of laminar and turbulent heat transfer and fluid flow of heat exchangers having large tubediameter and large tube-row by artificial neural networks. Int J Heat Mass Transfer 2009;52:2484–97. [2] Hasan MI, Rageb AA, Yaghoubi M, Homayoni H. Influence of channel geometry on the performance of a counter flow microchannel heat exchanger. Int J Thermal Sci 2009. [3] Salimpour MR. Heat transfer coefficients of shell and coiled tube heat exchangers. Exp Therm Fluid Sci 2009;33:203–7. [4] Ajakh A, Kestoras MD, Toe R, Peerhossaini H. Influence of forced perturbations in the stagnation region on Görtler instability. AIAA J 1999;37:1572–7. [5] Zhang P, Hrnjak PS. Air-side performance evaluation of three types of heat exchangers in dry, wet and periodic frosting conditions. Int J Refrig 2009. [6] Lasbet Y, Auvity B, Castelain C, Peerhossaini H. A chaotic heat-exchanger for PEMFC cooling applications. J Power Sources 2006;156:114–8. [7] Peerhossaini H, Castelain C, Le Guer Y. Heat exchanger design based on chaotic advection. Exp Therm Fluid Sci 1993;7:333–44. [8] Habchi C, Lemenand T, Della Valle D, Peerhossaini H. Liquid–liquid dispersion in a chaotic advection flow. Int J Multiphase Flow 2009;35:485–97. [9] Habchi C, Ouarets S, Lemenand T, Della Valle D, Bellettre J, Peerhossaini H. Influence of viscosity ratio on droplets formation in a chaotic advection flow. Int J Chem Reactor Eng 2009;7:A50.
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