Requirement for more compact and more efficient fan systems to improve the automotive thermal management has urged Valeo to develop a complete ...
2003-01-0615
Detailed CFD Modeling of Engine Cooling Fan Systems Airflow M. Henner, A. Levasseur and S. Moreau Valeo Motors and Actuators Copyright © 2003 Society of Automotive Engineers, Inc.
ABSTRACT
and the secondary flow moving around the motor. Therefore to achieve the optimal cooling of a given electrical motor, fan systems designers nowadays need a better and a more detailed understanding of the flow phenomena under the fan hub and around the electrical motor. This will also improve the fan system efficiency.
Requirement for more compact and more efficient fan systems to improve the automotive thermal management has urged Valeo to develop a complete automatic CFD procedure to numerically simulate the fan system. The present study has focused on the description and the validation of the flexible grid template that is meant to simulate the internal flow field within the fan system. This template closely fits in the previously developed rotorstator parametric grid. The initial small rib reference design has been meshed and four different large rib cases with increasing geometrical difficulties have helped testing this new capability. The simulations have been performed on a medium size grid that has been shown to capture most of the important flow features. The ribs act as centrifugal pumps that drain airflow through the electrical motor. Larger ribs are shown to increase the flow rate through the motor by up to 25% with a marginal increase with swept ribs compared to radial ones.
Since 1996 [2-5], Valeo has developed a complete fan system design strategy using 3D Navier-Stokes computation fluid dynamics (CFD) intensively. It not only helped Valeo improving its products but also reducing the cost and the delay of the product development for a given client request. Today, CFD is used daily in our R&D group and has already lead to significant, constant and predictable improvement of the design of fan and stator blades. It has also led to optimized cooling passages within its state-of-the-art compact electrical motors. The overall CFD approach has already been presented in [3-5] and relies on accurate low diffusion simulations on multiblock structured grids with CFX-TASCflow™ from AEA Technology. As stressed in [5], the basic design level has evolved from a basic parametric 3D in-duct model where the fan blade is set in an ideal cylindrical duct to a complete model which can include the tip clearance, a recirculation area behind the hub and some stator vanes behind the fan blade. In the present work, this topology is further extended to include the detailed hub structure with its ribs, the exact outer boundary of the electrical motor and the inner ring of the support. Thus a new specific grid template has been developed to study the fan system internal airflow. This new template yields the various so-called “under-fan-hub”configurations. After going through the design background which involves both experimental and numerical investigations, the overall model and grid specifications of the under-fan-hub airflow are presented. The parametric features are particularly emphasized. The aim of this template is to quickly estimate the effect of various parameters on the airflow going through the motors and the impact on the fan system efficiency. The first results with various large rib designs are then achieved in the present investigation. They are compared to the reference calculation with small ribs and temperature measurements in climate cells
INTRODUCTION Automotive thermal management by the cooling system in tighter and hotter underhood environments requires more and more sophisticated fan systems, producing both a high flow rate and a high pressure rise in a minimum space. Higher pressure rise often means higher torque and higher electrical motor loading and internal heating. Moreover, better underhood airflow management suggests using puller fan systems, which create a more stringent environment for the electrical motor. Efficient cooling of the electrical motor then becomes a necessity. This is even more critical with Valeo latest compact motor technology, which brings the highest power density of electrically driven fan systems on the market. In [1], the detailed CFD methodology put together to improve the internal flow structure of the electrical motor was presented. This will improve the pressure-flow rate characteristics of a given motor. Yet the exact position of the actual operating conditions on this curve is a complex combination of the main flow going through the axial fan 1
DESIGN AND EXPERIMENTAL BACKGROUND
GLOBAL RIB DESIGN – The ribs within the fan hub act as two centrifugal pumps in series. As a preliminary guideline the quasi-1D approach developed in [6] was applied. The flow within the rib passage is assumed uniform. Loss functions are accounted for to yield the final flow characteristics of the centrifugal pump (figure 1). The operating point is found by the intersection with the internal flow characteristics of the electrical motor. The latter is obtained experimentally on the test rig shown in figure 2. From this initial design a range of inlet and outlet flow angles were determined. These angles then provided the necessary inputs for the CAD concept of the fan hub with CATIA (figure 3). At this stage only circular mean lines had been assumed. Four different prototypes were then built from this CAD study: one has radial ribs and is termed “radial”, three have swept ribs and are called 16090, 160-110 (shown in figure 4) and 160-130 respectively.
INITIAL CFD STUDIES – A preliminary CFD investigation was first undertaken to understand the general features of the internal flow of engine cooling fan systems. The simplified model did not account for the fan blades and replaced the external flow through the blade by the theoretical exit speed triangles. Similarly the airflow through the electrical motor was simplified to the main internal airflow passage. The geometrical parameters were the hub size, the external motor ring size and the rib height. The rotational velocity was also varied. It was found that the ring and the hub radii should be similar and that a larger hub diameter and a larger rotational velocity increased the airflow through the motor. The presence of ribs was also found to be crucial for motor cooling. This led us to a first overall topology to focus on for the next detailed study. Having investigated the overall configuration it was time to consider the influence of the rib geometry and the effect of the coupling between the external flow through the blades and the internal flow under the fan hub. 11TA45 Restriction Curve
0.14 y = 0.0048x2 + 0.002x + 0.0015 y = 0.0038x2 + 0.0031x + 0.0017
Pressure rise dp in inH2O
0.12
0.1 11TA45Mot#1 11TA45Mot#2 Polynomial (11TA45Mot#1)
0.08
Polynomial (11TA45Mot#2) 0.06
0.04
0.02
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
flow cfm
Flow rate
Figure 3. CATIA design of fan hub with swept ribs
Figure 1. Typical rib flow characteristics
Figure 4. Fan prototype 160-110 with swept ribs Figure 2. Test rig for electrical motors flow characteristics 2
EXPERIMENTAL INVESTIGATIONS – The previously built fan prototypes were then assembled with the same support and several instrumented compact electrical motors (shown in figure 5). Each resulting fan system was then put on the same set of heat exchangers and tested in a climate cell at an ambient temperature of 100°C. Temperatures were measured with thermocouples at several key locations within the electrical motor: the brushes, the bushing on the back plate, the bearing on the front plate and the diode on the plastic brush holder. Strong temperature differences were found between the reference small ribs fan and the above four prototypes: the large ribs decreased the local temperatures by 5 to 10 degrees. Yet little temperature difference was found between the radial and the curved ribs. The latter result was a further incentive for a more detailed CFD analysis of the internal flow under the fan hub.
Figure 6. Overall model topology FAN SYSTEM MODEL– The fan is assumed symmetric and the support is assumed to have a number of stator vanes that are multiple of the rotor blades. The fan system model then involves only one rotor blade section and some multiple (here two) stator vane passages. The grid template for the external airflow is the stage configuration shown in figure 7 presented and validated in details in [5]. It comes from the extensive experience acquired on several blade families that has guided the technical choices for the grid generation, such as the overall topology, the node distribution and density. The end result is a high quality grid well suited to our very compact turbomachine stage. In the present study, the stator blades have not been considered to limit the final grid size.
Figure 5. Instrumented electrical motor
NEW UNDER-FAN-HUB DEVELOPMENTS OVERALL MODEL – The overall model consists of four main components: the fan model which provides the detailed external airflow; the electrical motor model which retains only the necessary aerodynamic characteristics, the under-fan-hub model which exhibits the detailed internal airflow and the support model which accounts for the structural blockage and possible flow leakage. The overall topology is outlined in figure 6. To get high quality and consistent results in the numerical simulations, parametric grid templates have been emphasized for all detailed models to ensure repeatability of grid generation and consequently comparable quality level of simulations on similar cases. These templates are written in the ACL command language of TASCgrid the integrated mesh generator of CFX-TASCflow from AEA Technology. The grid of each part is built once for a generic case and can be duplicated in a few minutes for all variations on the same topology. The four parts are then assembled to yield the final grid for the flow simulations.
Figure 7. External fan system topology Mesh Topology – The resulting grid topology has four blocks wrapped around the blade, one small inlet block 3
and one large outlet block. The central blocks require the ring, the hub and the motor profiles, and the blade airfoil sections at different radii. All grids used for fan simulation involve a hybrid H-O-H or H-C-H topology [8]. The O- or Cgrids are wrapped around the blade to ensure a proper resolution of the boundary layers around the profile. Four other H-blocks are included in the domain limited by each periodic face on the pressure and suction side of the blades. Various additional blocks model the tip clearance and the gap between the hub limit and the motor back plate. The resulting grid on the “solid” surfaces is shown in figure 8.
the symmetry and the 40° periodicity, the actual motor cooling passages are modeled by an equivalent annular passage at the same mean radius on the motor case with the same area as sketched in figure 10. The porous medium is then applied to this annulus. The actual characteristics of this porous medium are set by an external subroutine in the solver with a law that best fits the experimental measurements: ∆p = K ρ (v2/2) ∆L where K represents the internal resistance of the motor, v the mean speed going through the motor and ∆L the length of the motor case. This model was first validated on a simple cube. 11TA45 Restriction Curve
0.14 y = 0.0048x2
y = 0.0038x2 + 0.0031x + 0.00
Pressure drop dp in inH2O
0.12
0.1 11TA45Mot#1 11TA45Mot#2
Polynomial (11TA
0.08
Polynomial (11TA 0.06
0.04
0.02
0 0
0.5
1
1.5
2
2.5
3
3.5
4
4.5
5
flow cfm
Flow rate
Figure 9. Compact electrical motor flow characteristics
Figure 8. Surface grid of the blade section Boundary Conditions – The mass flow rate is set at the inlet and a constant pressure (atmospheric conditions) is imposed at the outlet. No slip condition is imposed in the proper reference frame for all solid walls. At the periodic boundaries, a periodic general grid interface (GGI) is imposed to account for the high fan blade stagger angles and to maintain the grid orthogonality at a reasonable level. Other frozen-rotor GGI are set at the interfaces between the stationary and rotary parts.
Actual geometry
Modeled geometry
Figure 10. Model of electrical motor flow passages
MOTOR MODEL– The internal structure of the electrical motor is not fully accounted for to keep the grid and model size reasonable. A typical grid size within the motor would add about a million nodes and would require simulating the full fan systems with its nine blades [1]. Yet a porous media models the effect of the internal structure with the experimental internal flow characteristics as inputs. Figure 9 is the latter for the present compact electrical motors. This pressure drop is measured between the front and back plates of the motor for a given flow rate. The two experimental curves in figure 9 give an idea of the dispersion due to measurements and process. To keep
HUB AND RIB MODEL– The parametric hub model is built in the same way as the fan system model. The number of ribs is assumed to be a multiple of the rotor blades. Presently only two rib passages have been considered following the results of a previous design of experiments that had been achieved on this parameter to decrease the motor temperatures with the same experimental set-up as above. This hub topology is then meant to fit exactly in the previous fan system topology: under the red part (fan blade root) of figure 7.
4
Mesh Topology – The grid topology requires the rib, the hub and the motor profiles. Similarly to the above fan blade section, the hub parametric grids involve a simple Hgrid topology. The node distribution in the H-grids wrapped around the ribs are fine enough to ensure a proper resolution of the boundary layers around their profile. The node distribution is also set to insure a smooth transition with the fan system grid (figure 8). The resulting grid is shown in figure 11. Again it should be stressed that a new grid can be quickly created for any change of sweep, stagger and profile of the ribs.
mesh sizes in the order of 600,000 nodes. Similar grid studies as in [5] have been done. A fine grid would require at least a million nodes. Accounting for the internal structure of the electrical motor would then require at least 10 millions nodes without the stator vanes!
Boundary Conditions –No slip condition is imposed in the proper reference frame for all solid walls. A one-to-one periodic interface is imposed at the periodic boundaries. A stationary frozen-rotor interface is set between the electrical motor cooling passage and the hub grid. Finally a GGI is set at the interface between the outlet of the hub grid and the fan system grid.
Figure 12. Complete under-fan-hub topology PHYSICAL MODEL – To account for the usual incompressible, highly rotational and three-dimensional environment of fan systems, the flow field has been assumed to be fully turbulent and has been modeled by the 3D RANS Navier-Stokes equations with a twoequations k -ε turbulence model. The resulting set of conservative equations has been solved with CFXTASCflow from AEA Technology on multi-block structured grids using its Multiple-Frame-of-Reference capability. Details of the numerical second-order schemes used for this type of computation can be found in [3]. Figure 11. Internal hub topology and mesh
At the solid boundaries wall laws have been applied to reduce the number of nodes close to the walls. All computations have been performed at a rotational speed of 2500 rpm and a volumetric flow rate of 2500 m3/h.
SUPPORT MODEL AND ASSEMBLY – The final part of the under-fan-hub model is the support. It involves the central ring of the support with the actual clearance with the motor case. It fits exactly in the previous fan system topology: under the blue part (stator feet) of figure 7. The grid is a simple H-grid topology whose node density is meant to match the node density of the fan system and hub templates as closely as possible. No-slip condition is imposed on the ring. One-to-one periodic boundary conditions are imposed on the periodic faces. GGI interfaces are set at the fluid boundaries with the other templates.
RESULTS Figures 13 and 14 show the induced flow topology in the hub and around the motor case for the large and straight rib case. As already seen in the initial simplified CFD model, the ribs act as centrifugal pumps that drain air from the back of the motor into the hub. Figure 13 shows the streamlines colored by speed within the hub with the motor and the support removed from the plot for sake of clarity. It emphasizes that the large ribs induce a significant flow field (large speeds) and therefore a significant mass flow rate into the electrical motor. Moreover significant momentum fluxes are injected in the hub boundary layer at the blade root as conjectured in [8] and measured by LDV. A comparison with the initial small ribs show similar patterns but with a much smaller flow rate. Figure 14 is a velocity vector plot in a meridional cut
For the final assembly, all the above parts are located in the same reference frame with their proper relative position to yield the topology shown in figure 12. The corresponding partial grids are attached to yield the final simulation grid. The same grid quality has been targeted for the different cases according to previous test rig simulations with medium resolution [4-5]. This has led to 5
through the electrical motor case, the fan hub and the support ring between two ribs. It first points out that the inlet flow rate splits between a main flow rate that enters the back of the motor and a secondary leakage flow between the support ring and the motor case. The latter goes back all the way to the front plate of the motor. The outlet flow is all concentrated between the ring and hub end. Figure 14 also shows that several secondary flow patterns happen in the hub: a large flow recirculation in the hub corner and another one in front of the support ring.
Rib Qmotor
Radial 12.35
160-90 12.93
160-110 12.7
160-130 12.21
(m3/h)
Gain % base 4.70 2.83 -1.13 Table 1: Volumetric flow rates in the electrical motor When comparing the mean flow velocity through the motor, all four large rib designs lie around 5 m/s whereas the initial small rib design lie 25% lower at about 4 m/s. The corresponding aerodynamic operating points of the electrical motor are shown in figure 15. These results closely correlate with the temperature measurements in climate cells described above. The flow field in the portion of the rib lying along the hub edge can explain the marginal increase in flow rate observed with the swept ribs. This second centrifugal pump currently operates offdesign in a quasi-stalled mode. On the contrary, the inlet angle of the first centrifugal pump is improved over the radial design yielding the marginal efficiency improvement. Dp dans le moteur 120.00
100.00
Optimized ribs (+25%)
Dp (Pa)
80.00
60.00
Initial small ribs 40.00 Essais Curve
Figure 13. Streamlines colored by speed with the support removed
fit
valeurs essais Test data
20.00
Obtenu calcu K98l Simulations 0.00 0
1
2
3
4
5
6
7
8
V (m/s)
Hub
Figure 15. Electrical motor operating conditions The final investigation has been the influence of the rib design and particularly the large rib case on the axial fan performances. Table 2 summarizes the results on the fan pressure rise and torque. No significant difference is found.
Q Motor
Q exit
Support
Q support
Rib ∆Pfan (Pa) Gain % Torque
Motor
Radial 216 base 1.3385
160-90 215.7 -0.14 1.3379
160-110 215.8 -0.09 1.3352
160-130 215 -0.46 1.3348
(Nm)
Gain %
Figure 14. Velocity vectors in a cut between the ribs The next step has been to simulate and compare the different rib designs. Similar flow patterns as figure 14 are obtained in all cases. Table 1 shows that quantitatively the flow rates achieved in the motor with the four large designs are very similar with a marginal increase of 5% for the 160-90 configuration.
base -0.04 -0.25 Table 2: Fan pressure rise and torque
-0.28
CONCLUSION The extensive use of CFD for fan design has been extended to the full fan systems with both the external flow field through the fan blades and the internal flow field under the fan hub through the electrical motor. Previous 6
9
work had involved automatic grid templates for three kinds of topologies, respectively the in-duct configuration, the virtual test rig and the rotor-stator configuration. The latter included some new relevant local effects as the detailed tip clearance labyrinth and the electrical motor presence. Based on this upgraded topology, a new automatic grid template for the under-fan-hub configuration has been developed which exactly fits in the previous template. It is designed to cover a wide range of rib design with high sweep and potentially various profiles and stacking.
4.
5.
6.
Four different large rib designs with increasing sweep and therefore geometrical difficulty were used to test this new grid capability. They involved the same fan used in previous studies. No stator blades have been included to limit the model size. All four geometry was meshed successfully with an equivalent grid quality of about 600000 nodes designed to capture the main relevant flow features.
7.
8.
A second step was to develop a simple realistic electrical model that bore most of its aerodynamic characteristics. A best fit of the available measured pressure drop was then supplied to the flow solver as a porous medium external subroutine. The flow simulations were then easy and fast enough to provide a detailed investigation of the four different rib shapes and to compare then to the initial small rib design. Gains of about 25% in flow rates with respect to the reference fan system were estimated for each large rib case at a single operating condition. That confirmed the trend observed in earlier temperature measurements made in climate cells in key locations within the electrical motor. The numerical comparison of various fan system internal configurations is now effective at an early stage of the product design process. Future work will include more elaborate rib design that will address the second pump stall issue. Additional flow rates will also be run to assess the sensitivity of the internal flow rates to the external flow rates. Finally results from [8] will be revisited to study the impact of the ribs on the wake topology of this nine-blade fan mounted in the MSU test rig.
REFERENCES 1.
2.
3.
“Thermal Analysis of Electric Motors in Engine Cooling Fan Systems,” T. Hong, M. Rakotovao, M. Henner, S. Moreau and J. Savage, SAE-2001-0110174 paper, Detroit, February 2001. “Improvement of Fan Design using CFD,” S. Moreau and E. Bennett, SAE-970934 paper, Detroit, February 1997. “CFD based Design for Automotive Engine Cooling Fan Systems,” E. Coggiola, B. Dessale, S. Moreau, R. Broberg and F. Bakir, SAE-980427 paper, Detroit, February 1998.
7
“On the use of CFD in the Automotive Engine Cooling Fan System Design,” E. Coggiola, B. Dessale, S. Moreau, and R. Broberg, AIAA 98-0772 paper, Reno, January 1998. “Latest Improvements of CFD Models of Engine Cooling Axial Fan Systems,” M. Henner, S. Kessaci and S. Moreau, SAE-02HX-36 paper, Detroit, February 2002. “Modélisation du transfert de chaleur au sein des machines électriques tournantes. Dimensionnement et Optimisation de leur système de refroidissement,” C. Vasilescu, PhD dissertation, Université Paris VI, 2002. “Unsteady Rotor-Stator Interactions in Automotive Engine Cooling Fan Systems,” M. Henner, M. Stanciu, S. Moreau, S. Aubert and P. Ferrand, Proceedings of the ISUAAAT 2000 Conference, Ecully, September 2000. “Evaluating CFD Models of Axial Fans by Comparisons with Phase-Averaged Experimental Data,” D. Neal, S. Moreau, M. Henner and J. Foss, VTMS5-01-89 paper, Nashville, May 2001.
