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SAE TECHNICALPAPER SERIES 2003-01-0613

CFRM Concept at Vehicle Idle Conditions

Zhigang Yang, Jeffrey Bozeman and Fred Z. ShenGeneral Motors Corporation

James A. AcreDelphi Corporation

Reprinted From: Thermal Management(SP-1751)

2003 SAE World CongressDetroit, Michigan

March 3-6, 2003

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ISSN 0148-7191Copyright © 2003 SAE International

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ABSTRACT

The concept of condenser, fan, and radiator power train cooling module (CFRM) was further evaluated via three-dimensional computational fluid dynamics (CFD) studies in the present paper for vehicle at idle conditions. The analysis shows that the CFRM configuration was more prone to the problem of front-end air re-circulation as compared with the conventional condenser, radiator, and fan power train cooling module (CRFM). The enhanced front-end air re-circulation leads to a higher air temperature passing through the condenser. The higher air temperature, left unimproved, could render the vehicle air conditioning (AC) unit ineffective. The analysis also shows that the front-end air re-circulation can be reduced with an added sealing between the CFRM package and the front of the vehicle, making the CFRM package acceptable at the vehicle idle conditions.

INTRODUCTION

Condenser, radiator, and fan are standard equipment of vehicle HVAC sub-system for vehicle underhood thermal management, power train cooling, and air conditioning. These components are typically packaged into a condenser, radiator, and fan module (CRFM) with the fan located at the rear of the package. Pusher fan configuration, in which the fan is placed at the front of the package, has also been in service and is mainly being used as auxiliary cooling fan to supplement the main fan located at the rear of the package. A recent entry for the configuration of packaging these components is the center-mounted-fan (CMF) configuration

proposed by Delphi Corporation [1]. In this packaging configuration, the fan is placed at the center of the package and the subsystem module is now a condenser, fan, and radiator module (CFRM).

In a previous work [2] reported in the SAE 2002 World Congress, three-dimensional numerical analysis was carried out with the aid of CFD tools to assess the performance of the CFRM concept in a vehicle level thermal system analysis. It was found that at the fan speed of 2500 rpm, the CFRM configuration delivers an air stream entering the vehicle underhood area with an average temperature that is about 10°C cooler than the air stream temperature delivered by its CRFM counterpart operating at the same fan speed. The difference in temperature was mainly due to the higher mass flow rate impelled by the fan in the CFRM configuration. The mass flow rate impelled by the CFRM fan can be reduced with reduced fan speed. It was found that at the fan speed of 2160 rpm, the mass flow rate for the CFRM package is the same as the mass flow rate of the CRFM package with the fan operating at the speed of 2500 rpm. The radiator exit temperatures were found to be similar for these two cases, and the vehicle underhood components experience similar thermal environment. However, the fan in the CFRM configuration was found to use 19% less power, due mainly to the reduction in the fan speed. Such a saving in fan power consumption suggests that the CFRM configuration could be a viable alternative to the CRFM configuration currently in use today.

The above performance comparison between the CFRM configuration and the CRFM

2003-01-0613

CFRM Concept at Vehicle Idle Conditions

Zhigang Yang, Jeffrey Bozeman and Fred Z. Shen General Motors Corporation

James A. Acre Delphi Corporation

Copyright © 2003 SAE International

configuration was for a passenger sedan driving at an extreme condition with the vehicle climbing a grade road while pulling the maximum rated trailer at the gross combined vehicle and trailer mass rating. [This driving condition is referred to as the trailer grade condition for the rest of the paper.] Automobile vehicles operate in a wide range of conditions, and it is thus necessary to assess the performance of the CFRM package at different operation conditions before the packaging concept can be installed in a real vehicle.

The objective of the present paper is to assess the performance of the CFRM configuration at vehicle idle conditions. Because of the lack of the swirl velocity and thus the lower momentum of the air stream exiting the CFRM package, its performance at vehicle idle conditions was suspected to be of concern, as discussed in [2]. The assessment was carried out using three-dimensional CFD analysis in the present study. The structure of the present paper is as follows: First, framework and methodology of the present assessment analysis are described. Next, detailed numerical results of the CFRM configuration and the CRFM configuration at vehicle idle conditions are shown. Conclusions and discussions are given in the final section of the paper.

METHOD OF ANALYSIS

Our analysis in the present study was a vehicle level thermal system analysis, in which the thermal fields due to each of the vehicle segments (front end, underhood, and underbody) are calculated and analyzed together as one thermal system. This is in contrast with the traditional approach in which thermal analysis for automobile is carried out by focusing on a segment of the vehicle at a time, for example, underbody thermal analysis, underhood thermal analysis, front end flow analysis. The vehicle thermal system approach is made feasible due to the development in computational resources in terms of both computer hardware and computational technologies. The vehicle thermal system approach has the advantage over the segment-based approach in that the difficulties and ambiguities of setting up the conditions at the inter-segment locations are removed.

In the vehicle thermal system analysis, the heat exchangers (condenser and radiator) are still

modeled rather than being solved directly, due to the difficulties in resolving the small scale geometries related to the heat exchanger fins and the airflow passing through them. In this modeling approach, rather than simulating the heat exchangers explicitly, only the effects of the heat exchangers on the airflow field are considered and modeled.

The heat exchangers have an effect on both the velocity field and the thermal field of the airflow passing through the heat exchangers. For the velocity field of the airflow, the heat exchangers are represented as porous media, which yields a drag force to the airflow passing through them. The effect of this drag force is that the airflow loses pressure when passing through the heat exchangers. For a given heat exchanger, the magnitude of the pressure loss is a function of the velocity of the airflow going through it. For the thermal field of the airflow, the heat exchangers are represented by volume heat sources injecting heat to the incoming airflow passing through the heat exchangers. The amount of heat injected into the airflow, for a given construction of the heat exchanger, is related to the mass flow rate of the airflow, the mass flow rate of the coolant passing the heat exchanger, and the temperatures of the airflow and the coolant. Both the pressure loss and the thermal performance characteristics of a given heat exchanger can be determined in a component level test, or are provided by the heat exchanger manufacturers.

Heat exchangers are modeled in this way in both the previous study [2] and the present study. The commercial CFD software Star-CD [3] was used in the previous study. Details of modeling the heat exchangers and implementing the model in Star-CD are given in references [4-5]. In the present study, the commercial CFD software Fluent [6] was used to study the performance of the CFRM concept at vehicle idle conditions. In the Fluent code, the implementation of the heat exchanger model is achieved via a user-defined function (UDF), which was provided to us by Fluent Inc. [7]. It should be pointed out that the heat exchanger models implemented in the Star-CD code and the Fluent code have the same level of sophistication, and should not produce any noticeable differences for the same physical problem.

While the heat exchangers were modeled in a similar manner in the previous study and in the

present analysis, fan treatment was completely different. In the previous study, the fan was modeled using the momentum source method. In this modeling approach, similar to the modeling of heat exchangers, only the effect of the fan on the airflow is considered and modeled. The modeling is achieved by adding source terms in the momentum equations such that the solution of the governing equations will reproduce the fan’s pressure rise vs. flow rate relationship, which is normally called fan performance curve. The fan performance curve can be obtained from the fan performance test or is made available by the fan manufacturers. Details of modeling the fan using the momentum source method can be found in reference [8], for example. In the present analysis, however, the flow field due to the fan blade rotation was simulated directly using the multiple reference frame (MRF) methodology. In the MRF approach, the fluid domain is divided into a few sub-domains, with the fluid enclosing the fan solved in a reference frame that is rotating with the fan and the bulk of fluid solved in the laboratory reference frame. Numerical solutions for these sub-domains are matched at the interface of the sub-domains during the solution process. The MRF approach for fan simulation is available in both the Star-CD code and the Fluent code.

Of the two approaches, the momentum source approach has the advantage of being numerically more stable and faster in analysis time since the fan blade geometry does not need to be resolved. The MRF approach has the advantage of being able to generate more information, including the flow field over the fan blades, the fan performance characteristics, fan efficiency, and load distribution over the fan blade. In our previous study, one uncertainty associated with the use of the momentum source approach was whether it was valid to use the same set of fan performance curves in our assessment of the performance comparisons between the CRFM configuration and the CFRM configuration. One of the motivations of adopting the MRF approach for fan in our present analysis was to remove this uncertainty.

The models for the CFRM configuration and the CRFM configuration were then formed by putting together the models for each component, i.e., condenser, radiator, and fan with fan shroud, in different orders, as shown in figures 1 and 2, respectively. Identical components were used in

these two configurations and the performance characteristics for condenser and radiator were assumed the same for both the CFRM and the CRFM configurations. The condenser location remains the same for both packaging configurations. Also fixed is the total size of the package, measured in the direction from the front to the rear of the vehicle. In addition, the CFRM package and the CRFM package are perfectly sealed on the side of their packages.

The resulting model for the CFRM package or the CRFM package was then mounted to a vehicle model representing a typical family sedan. In our vehicle level thermal system analysis, other major underhood, underbody, and exterior components of the vehicle were included in the vehicle model, and the flow and thermal fields about these components were directly calculated. An effort was made to include all the components that have a dimension of 75mm or larger. Altogether, more than 100 components were included in the final vehicle model. The model structure is quite flexible in that any other component can be easily added, if it is deemed thermally important. The resulting vehicle model was then placed in a computational domain consisting of a box of 23m in length, 10m in width, and 5.5m in height, the same size as the test section of the General Motors Aerodynamics Laboratory [9]. This is the

Figure 1: CRFM configuration

Figure 2: CFRM configuration

typical computational domain for our vehicle thermal system analysis. For the present study on the thermal performance at vehicle idle conditions, a much smaller computational domain could have been used.

In the present numerical analysis, the vehicle component surfaces were discretized with triangle meshes. The size of the triangle meshes varies with the component in question. A typical mesh on component of interest has a dimension of 10-15mm. Altogether, more than 200,000 surface elements were used. Tetra mesh was used to discretize the fluid in the computational domain. Such a triangle/tetra mesh structure for surface/volume discretization allows fast and easy representation of the complex geometry often found in an automobile. The number of the volume meshes is strongly dependent on the size of the surface elements and the size of the computational domain. In the present study, more than 2 million tetra cells were used for the fluid volume. These tetra cells were generated using TGrid [10].

The spatial location of the computational domain and the vehicle are characterized by their coordinates in a coordinate system with x in the direction from the front to the rear of the vehicle, z in the direction normal to the ground, and y in the direction that makes the coordinate system right-handed. The origin of the coordinate system is locates such that the front wheels have the x-coordinate of zero. This is also the location of the balance at GMAL.

Due to the heat rejection from the condenser and the radiator, the temperature of the airflow increases, and the density of the airflow decreases. The density variation leads to a coupling between the velocity field and the thermal field. In the present study, the coupling is facilitated by the use of the following equation of state,

ρ = pa/RT, (1)

where pa is the ambient atmospheric pressure. Such an equation of state is necessary in order to describe the fundamental physics causing the performance differences between the CFRM configuration and the CRFM configuration in a vehicle level thermal analysis. A detailed analysis of and a description for such a necessity is given in Yang et al. [11].

In the present analysis, the turbulent flow field was modeled using the realizable k-ε model of Shih et al. [12]. The convective heat transfer due to turbulent fluid motion was modeled with the Reynolds Analogue and the assumption of a constant turbulent Prandtl number. At the solid wall surfaces, the wall-functions approach of Launder and Spalding [13] was used to represent the turbulent wall boundary conditions. In a typical vehicle level thermal system analysis such as the one in the previous study [2], all three mechanisms of heat transfer, conduction, convection, and radiation, are included. In addition, conjugate heat transfer is assumed to take place so that the temperature distribution on a component of interest can be accurately found. In the present study, a simplified formulation for the heat transfer and the thermal field was adopted in which the radiation heat transfer was neglected and the thermal boundary condition at any solid wall surface was assumed adiabatic. Such a simplified approach was possible because in the present study we were only interested in the air temperature, particularly the air temperature at the condenser inlet, rather than the temperature on the surface of any solid component. Since the radiation heat transfer takes place between solid surfaces, our neglect of this heat transfer mechanism and the assumption of the adiabatic wall conditions would only lead to inaccurate prediction of the temperature distribution on the solid surfaces of the components. Because the turbulent Prandtl number is close to one (0.9 for air), the thermal boundary layer and the momentum boundary layer have comparable thickness (the ratio between the two boundary layers is √Pr, where Pr is the turbulent Prandtl number), and both become vanishingly small as the Reynolds number becomes large. For the flow filed at the vehicle idle conditions, the Reynolds number is large mainly due to the small viscosity of air. Thus, any inaccuracy of the temperature on component solid surfaces will affect only the fluid next to the surface, and have a negligible effect on the temperature of the bulk of fluid due to the low thermal conductivity of air. Thus, our simplifications for the heat transfer and the thermal boundary conditions at solid wall surfaces are justified.

The commercial CFD code Fluent was used for the current analysis, as mentioned earlier. The code is a finite volume based code using unstructured grid system. It supports the triangle and tetra grid structure for the surface mesh and

the volume mesh as used in our discretization. Computations using the Fluent code were carried on GM’s HPC (High Performance Computing) platform. For the present framework of analysis and the number of cells mentioned above, the numerical computation took about nine (9) hours to finish using eight (8) processors, making it possible to generate useful results overnight and able to impact design consideration the following day.

RESULTS OF THE ANALYSIS

Boundary conditions need to be specified at the boundary of the computational domain to carry out the numerical analysis. Theoretically, the boundary conditions should be set such that the pressure at the boundary of the computational domain will be in balance with the ambient pressure. However, such a setting of the ambient pressure boundary condition is very difficult to achieve in the numerical study, and the resulting numerical analysis is less stable numerically. Instead, in the present study, a small wind is imposed on the ambient environment, and the resulting setup for the boundary conditions are then the inflow condition, the outflow condition, and the inviscid wall condition at the boundaries of the computational domain, just like the setup for the case of moving vehicle. Such a setup facilitates the numerical convergence of our study. It is observed that for a small ambient wind speed the airflow rate going through the heat exchangers and the flow field in the underhood region are determined solely by the fan operation. In addition, it can be argued that the setup with a small ambient wind gives a more representation of a vehicle at idle conditions, since the actual vehicle operates in an environment that always has some (small) ambient wind.

In the numerical solution process, solution for the flow only (cold flow at iso-thermal condition) was solved first. After reaching numerical convergence, the thermal field was then turned on and both the equations for the flow field and the energy equation were solved as a coupled set of equations. First order upwind scheme was used throughout our numerical study for its numerical stability properties. Since we were interested in the airflow temperature at the condenser inlet and the temperature in the bulk of the fluid domain in the underhood area rather than the temperatures at and/or near the

component surfaces, the first order scheme was deemed sufficient. The conclusion was later confirmed by calculations using the second order scheme.

For all the cases to be presented later, the analysis was carried out by assuming that the ambient air had a temperature of 40°C, the heat rejection rate for the condenser and for the radiator was set to 10KW each, and each of the two parallel fans operated at the fan speed of 2500 rpm. In each case, the objective was to compare the performance of the CRFM configuration and the performance of the CFRM configuration, especially in terms of the overall underhood thermal environment and the problem of front-end air re-circulation.

CASE I: U = 1M/S

First, a small velocity of 1m/s was imposed at the inlet of the computational domain. Both the CRFM configuration and the CFRM configuration were calculated. To illustrate the overall vehicle underhood thermal environment, temperature distributions on two perpendicular planes, the first being the vehicle symmetry plane with y=0 and the second being parallel to the ground with z=0.4m, were plotted. These temperature distributions are shown in figures 3-4 for the CRFM configuration and figures 5-6 for the CFRM configuration, respectively. By comparing these temperature distributions, it is seen that the peak underhood air temperature is slightly higher for the CRFM package. However, the temperature for the bulk of the air near the vehicle front dash area is higher for the CFRM package. This is partly because the airflow spreads more widely for the CRFM package due to the swirl velocity of the fan.

Overall, the temperature distributions for the CRFM package and the CFRM package are in the similar temperature range. In addition, it is noted that the peak air temperatures for both the CRFM package and the CFRM package are a lot lower than the corresponding peak air temperatures at the trailer grade condition, as reported in the previous study [2]. This is because the fan draws more airflow per unit heat rejection rate from the heat exchangers at the vehicle idle conditions than at the trailer grade conditions. Thus, from the point of view of the underhood thermal environment, idle conditions do not pose any challenge, at least for vehicles with electrical fan(s).

At vehicle idle conditions, the critical issue from the point of view of the HVAC performance is the airflow temperature at the condenser inlet location. This airflow temperature can be a lot higher than the ambient air temperature due to the phenomenon of front-end air re-circulation. When front-end air re-circulation takes place, rather than taking in fresh air from the ambient, the fan draws the hot air exiting from the radiator back into the heat exchangers again. As a result, the air temperature at the condenser inlet can become substantially higher than the ambient air temperature. The airflow temperature at the condenser inlet has a direct impact on the performance of the vehicle AC system at vehicle idle conditions. When the air temperature at the condenser inlet is raised, the AC compressor must operate at a higher pressure in order to reject the required amount of heat. The higher pressure needs more compressor work, and thus less cooling capacity is available for the vehicle cooling comfort. When the temperature is too high, the required pressure can be too high for the vehicle AC compressor, which can lead to shutting down of the vehicle AC system.

The air temperature at the condenser inlet is shown in figure 7 and figure 8 for the CRFM package and the CFRM package, respectively. There are some hot spots, suggesting the existence of the front-end air re-circulation. The average air temperature at the condenser inlet for the CRFM package is 49.5°C, and the average temperature for the CFRM package is 49.9°C. In both cases, the temperature distribution is rather uniform and the average condenser inlet temperature is less than 10°C higher than the ambient temperature. Overall, the issue of front-end air re-circulation is not very serious for this case.

Figure 3: Underhood temperature (°C) distribution for CRFM (on y=0 plane)

Figure 4: Underhood temperature (°C) distribution for CRFM (on z=0.4m plane)

Figure 5: Underhood temperature (°C) distribution for CFRM (on y=0 plane)

Figure 6: Underhood temperature (°C) distribution for CFRM (on z=0.4m plane) Figure 7: Temperature (°C) distribution at

condenser inlet (CRFM configuration)

CASE II: U = -1M/S

For the second case, the small ambient wind was assumed to be blowing from the rear of the vehicle. In actuality, a wind from the rear of the vehicle can happen as much as the ambient wind from the front of the vehicle. The key question for us is if such a change in the direction of the (small) ambient wind causes much change in the thermal field of our interest.

The temperature distributions for the overall underhood area are shown in figures 9 and 10 for the CRFM configuration and the CFRM configuration, respectively. Compared with the temperatures for the case of u=1m/s shown in figure 3 and figure 5, the temperatures become higher for both configurations with a more pronounced temperature increase for the CFRM package. It should be pointed out that the iso-thermal calculation would have yielded the same cooling flow rate passing through the heat exchangers as for the case of u=1m/s, suggesting that the iso-thermal cooling flow rate is NOT the only parameter determining the vehicle underhood thermal environment. On the other hand, it is still not as high as what one finds in the trailer grade conditions.

The air temperature at the condenser inlet has seen an even more dramatic change. These temperatures are shown in figures 11 and 12 for the CRFM configuration and the CFRM configuration, respectively. Compared with the case of wind blowing from the vehicle front given in figures 7 and 8, the average air temperature at the condenser inlet now goes up to 56.5°C for the CRFM package, and up to 58.7°C for the CFRM package. Such a high average air temperature at the condenser inlet reduces the capacity of the vehicle AC system due to the higher compressor pressure that is required.

Figure 8: Temperature (°C) distribution at condenser inlet (CFRM configuration)

Figure 9: Temperature (°C) distribution for the CRFM configuration for u=-1m/s (y=0 plane)

Figure 10: Temperature (°C) distribution for the CFRM configuration for u=-1m/s (y=0 plane)

Figure 11: Temperature distribution at condenser inlet for u=-1m/s (CRFM

configuration)

Figure 12: Temperature distribution at condenser inlet for u=-1m/s (CFRM

configuration)

Since the ambient air has a temperature of 40°C, the hot air spots at the condenser inlet plane must be due to the front-end air re-circulation. To see clearly, the velocity vector at the y=0 plane and the z=0.4m plane are shown in figure 13 and figure 14. For better visual clarity, velocity vectors are projected in the plane of choice and all velocity vectors are drawn with the same length. The color on the velocity vector indicates the temperature at the point where velocity is drawn.

It is seen clearly that the hot air exiting from the radiator is being drawn back and fed into the heat exchanger again. As a result, the temperature at the condenser inlet becomes higher than the ambient air temperature. Such an air re-circulation exists for both the CRFM package and the CFRM package. Due to the small swirl velocity for the CFRM package, the impact is more severe and the resulting average air temperature at the condenser inlet is higher. The heat rejection from the heat exchangers heats up the incoming airflow. For a fixed heat rejection rate, the air temperature at the radiator exit will be higher if the air temperature at the condenser inlet is higher. This explains why the underhood air temperatures shown in figures 9

and 10 are higher. Figures 13 and 14 also show that the air re-circulation has the main path of going through the top and the sides of the CRFM and CFRM packages. Such information can help to make design modifications to reduce the air re-circulation. This is the topic of the next sub-section.

CASE III: U = -1M/S WITH SEALING

One remedy of reducing the front-end air re-circulation is to block the air re-circulation path shown in figures 13 and 14 by sealing the gaps between the CRFM (and the CFRM) package and the vehicle front. A seal was added, and the computations were carried out again. The resulting air temperatures at the condenser inlet plane are shown in figure 15 for the CFRM package.

It is seen from figure 15 that the air temperature at the condenser inlet plane is much cooler now as compared with the case without the seal, which was shown in figure 12. The average air temperature at the condenser inlet plane is reduced from 58.7°C to 48.5°C. In addition, the temperature distribution is more uniform. To confirm that the reduction in average temperature was achieved by the reduction in vehicle front-end air re-circulation, velocity vector at the y=0 plane is shown in figure 16, and the color again indicates the temperature at the location of the velocity vector. It is seen that the air re-circulation from the top of the CFRM package is blocked and only fresh air from the ambient is drawn into the heat exchangers. As a result of reduction in front-end air re-circulation, the vehicle underhood thermal environment is improved, as shown in figure 17 (and in comparison with figure 10). It is noted that the seal that reduces the front-end air re-circulation

Figure 13: Air re-circulation at the y=0 plane

Figure 14: Air re-circulation at the z=0.4m plane

Figure 15: Temperature distribution at condenser inlet for u=-1m/s (CFRM

configuration with seal)

is preferred for other reasons as well. For example, the seal would help reducing the vehicle aerodynamic drag and improve the cooling flow rate at the vehicle trailer grade conditions.

CONCLUSIONS

The performance of the CFRM packaging concept at the vehicle idle conditions are analyzed and assessed in the present study. For vehicle with electrical fans, vehicle idle conditions do not pose much challenge for power train cooling, since the vehicle must withstand more severe power train cooling requirement at trailer grade conditions, as the results of the previous study [2] had shown. On the other hand, a unique thermal issue facing the vehicle at idle conditions is the air temperature at the condenser inlet. Because air has little momentum due to the vehicle motion, hot air exiting the radiator can be drawn back into the heat exchanger, thus much increasing the temperature at the condenser inlet.

The present study assesses the performances of the CRFM package and the CFRM package in terms of the problem of front-end air re-circulation. Due to the low swirl velocity and thus the lower momentum of the hot air leaving the radiator, the CFRM package generates a higher degree of air re-circulation when compared with its CRFM counterpart. On the other hand, sealing of the CFRM package with the vehicle front can reduce the degree of air re-circulation, and thus reduce the average air temperature at the condenser inlet substantially. With an added seal, the air temperature at condenser inlet was brought down to acceptable level. The average air temperatures at the condenser inlet plane for the few cases calculated in the present study are summarized in the following table.

Table I: Average condenser inlet air temperature

U=1m/s U=-1m/s U=-1m/s (with seal)

CRFM 49.5°C 56.5°C 48.0°C

CFRM 49.9°C 58.7°C 48.5°C

Test was also carried out at the vehicle idle conditions. The following three types of systems were tested: 1) Production CRFM package on a vehicle; 2) Prototype CFRM on the same vehicle; and 3) Prototype CFRM with an air discharge deflector on the same vehicle. The average air temperatures at the condenser inlet plane for these three cases are given in table 2.

Table 2: Average condenser inlet air temperature

Production CRFM

Prototype CFRM

Prototype CFRM

with deflector

Temp. 57.2°C 60.6°C 56.1°C

Because of the variations in vehicle geometry and vehicle operating conditions such as the heat loads and the fan speed, direct comparisons between the CFD simulation and

Figure 16: Air re-circulation at the y=0 plane for CFRM package with seal

Figure 17: Temperature (°C) distribution for the CFRM configuration with seal for u=-1m/s

the test can not be made. On the other hand, it is worth noting that both the simulation and the test show the same general trend: the CFRM package will provide a lower cooling performance than the CRFM package with the addition of seal (in the analysis) or the air deflector (in the test). With the addition of the seal (in analysis) or the air deflector (in test), the performance of the CFRM package is improved and is acceptable for vehicle at idle conditions.

The present study also demonstrates the importance of carrying out calculations for vehicle at idle conditions with the ambient velocity of both u=1ms and u=-1m/s. In the case shown, u=-1m/s is much more critical in terms of the vehicle front-end air re-circulation, and thus the vehicle AC performance. It can be argued that for most vehicles, the case for u=-1m/s will yield results that are more critical than those obtained with u=1m/s if the air re-circulation is of concern.

ACKNOWLEDGMENTS

The surface mesh of the vehicle was provided by K. Elankumaran. Z. Yang would like to thank Mark Franchett, Che-Hsi Yu, and Vijay Damodaran for interesting discussions. He would also like to thank Greg Fadler, his manager, for generous support during the writing of this paper. The authors would like to thank the SAE reviewers for their comments on this paper.

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CONTACT

For information regarding this paper, please contact: Zhigang Yang General Motors Corporation Phone: 586-578-3688 E-mail: Zhigang.Yang@gm.com For information regarding center-mounted fan (CMF), please contact: James A. Acre Delphi Corporation Phone: 716-439-2932 E-mail: james.a.acre@delphiauto.com