Conference Paper

NREL/CP-540-46172

August 2009

Rapid Modeling of Power 

Electronics Thermal Management 

Technologies 

Preprint

K. Bennion and K. Kelly

Presented at the 5th IEEE Vehicle Power and Propulsion Conference

Dearborn, Michigan

September 7–11, 2009
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1 

Rapid Modeling of Power Electronics Thermal 

Management Technologies 

Kevin Bennion and Kenneth Kelly

National Renewable Energy Laboratory

Center for Transportation Technologies and Systems

Golden, Colorado, USA

Abstract—A methodology was developed to rapidly evaluate 

trade-offs associated with alternative packaging configurations 

and thermal management technologies for power electronics 

packaging. The methodology includes the integration of available 

experimental correlations, computational fluid dynamics results, 

parametric 3D finite element analysis (FEA) thermal models, and 

established heat exchanger analysis techniques. The parametric 

3D FEA model enables sensitivity studies related to the power 

module package configuration and cooling technologies. This 

paper focuses on the study of alternative cooling technologies as 

they are applied to a fixed power module package. The 

methodology is applied to a double-sided power module package 

for several alternative cooling technologies. 

Keywords – power electronics, thermal management, thermal 

control, thermal resistance, hybrid electric vehicle, electric vehicle, 

fuel cell vehicle

I. INTRODUCTION

Electric drive systems, which include electric machines and 

power electronics, are a key enabling technology for advanced 

vehicle propulsion systems that reduce the dependence of the 

U.S. transportation sector on petroleum. However, to penetrate 

the market, these electric drive technologies must enable 

vehicle solutions that are economically justifiable. As critical 

components of the electric drive system are made smaller, 

lighter, and more cost effective, heat removal becomes an 

increasing challenge. 

To support the development of environmentally friendly 

transportation technologies that reduce petroleum consumption, 

the U.S. Department of Energy (DOE) supports research and 

development in electric drive systems through the Vehicle 

Technologies program. The Advanced Power Electronics and 

Electric Machines (APEEM) R&D team develops power 

electronics technologies applicable to hybrid electric vehicles 

(HEVs), plug-in hybrid electric vehicles (PHEVs), fuel cell 

vehicles, and electric vehicles. The APEEM team includes

members at DOE and researchers at the national laboratories 

who work through the FreedomCAR and Fuels Partnership 

with the U.S. automotive industry. As part of this partnership, 

DOE’s National Renewable Energy Laboratory (NREL) leads 

R&D activities in the thermal control of advanced power 

electronics and electric machines.

To successfully integrate advanced power electronics 

concepts into vehicle applications, the thermal limitations of 

the semiconductor devices must be addressed. Critical 

semiconductor components such as insulated gate bipolar 

transistors (IGBTs) and diodes are sensitive to temperature and 

must operate below fixed temperature limits of 125°C to 150°C 

for silicon devices [1]. The vehicle must achieve its 

performance requirements as it operates within this thermal 

limitation while meeting the goals for system reliability, cost, 

volume, and weight. The APEEM program emphasizes the 

importance of cost, weight, volume, efficiency, and reliability 

for the widespread use of electric drive systems in vehicle 

applications [1], and each of these areas has a direct relation to 

the thermal control of electric drive systems. 

The design of the thermal control system is directly tied to 

the packaging configuration of the semiconductor devices and 

the overall vehicle thermal management system. The direct link 

between semiconductor packaging and thermal control is 

critical [2,3] in that conventional methods to increase power 

output by increasing silicon area are too expensive in terms of 

cost and volume. For this reason alternative methods are 

needed to improve heat removal of innovative high power 

APEEM systems. Also, thermal control is tied to the overall 

vehicle thermal management strategy. For example, one 

research goal under the FreedomCAR and Fuels Partnership is 

to use only one cooling loop for an HEV [1,4]. A single 

cooling loop cools the power electronics, electric machines, 

and internal combustion engine. The power electronics inverter 

would thus have to tolerate inlet coolant temperatures (Tc,i) up 

to 105°C [1]. 

Addressing goals associated with performance, cost, 

weight, and volume requires investigation into multiple thermal 

control technology pathways that involve the vehicle 

propulsion configuration (such as HEV or PHEV), component 

use (vehicle drive cycles), cooling system configuration, power 

electronics package configuration, and heat exchanger design. 

A systems approach is necessary to investigate the multiple 

options and trade-offs associated with the items mentioned

above. Analysis techniques are needed that allow researchers to 

quickly investigate the system-level impacts of potential 

technologies and evaluate trade-offs to understand the design 

space. This paper outlines one such method for investigating 

the system-level performance of power semiconductor package 

configurations and heat exchanger technologies.

The authors would like to acknowledge the support provided by Susan 

Rogers, Technology Development Manager, Advanced Power Electronics and 

Electric Machines, DOE Vehicle Technologies Program.
2 

II. APPROACH

The APEEM program area within the DOE Vehicle 

Technologies Program is currently developing a suite of 

advanced thermal control technologies, including single-phase 

and two-phase jet impingement, air cooling, low thermal 

resistance semiconductor packaging structures, and improved 

thermal interface materials [5]. Because of the number of 

available packaging and thermal control options, a method for 

analyzing the system thermal performance was developed. The 

developed methodology, general in nature, allows the system 

performance characterization of multiple package and heat 

exchanger technology options within a consistent framework. 

The multistep approach for this work is summarized in Fig. 1. 

The process involves a merging of techniques to characterize 

heat transfer technologies and fluid flow with techniques that 

characterize the thermal performance of packages for power 

electronics applications. The end result is the ability to evaluate 

the total system thermal performance. This allows the 

evaluation of multiple heat exchanger technologies applied to 

power semiconductor package configurations. The intent of 

this paper is to provide an example of the analysis techniques 

on a power semiconductor package of interest.

A. Package Selection and Thermal Characterization

First (as shown at the bottom of Fig. 1), a double-sided

package for a power semiconductor module was selected that 

consisted of one IGBT and one diode [2,3]. This package is 

shown in Fig. 2; this package was selected because it is used in 

a commercial HEV and allows cooling on both sides of the 

package. A double-sided package was of interest because 

previous analysis showed the potential benefit of doublesided cooling for power semiconductor packages [6]. The 

Figure 1. Thermal system analysis process.

package was also of interest because it was part of an APEEM 

benchmarking effort conducted by Oak Ridge National 

Laboratory (ORNL) [7]. Fig. 3 shows an exploded view of the 

layers used as part of the thermal analysis described in this 

report. Table I provides the material information associated 

with the layer labels in Fig. 3 based on information provided by 

ORNL through its benchmarking work and on [2,3]. 

The CAD software “SolidWorks” was used to construct a 

three-dimensional solid model of the package. The geometry 

was then imported into the software “ANSYS Workbench” for 

thermal analysis using the finite element analysis (FEA) 

methodology (see Fig. 4). For this analysis, a constant IGBT to 

diode loss ratio of 3:1 was used (see Figs. 4, 5, and 6). The 

IGBT and diode heat loads were applied with volumetric heat 

generation for each device. A convective heat transfer 

coefficient was applied to the top and bottom surfaces in Fig. 4

Figure 2. Double-sided package showing side view of IGBT and diode with 

a 50 mm by 30 mm heat exchanger base area. 

Figure 3. Double-sided package showing side view of IGBT and diode.

TABLE I 

MATERIAL KEY

Index Material X [mm] Y [mm] Z [mm]

A Heat sink: Al 30 50 0.6

B TIM: Grease 21 50 0.1

C Substrate: SiN 21 50 0.3

D TIM: Grease 21 50 0.1

E Heat Spreader: Cu 21 35 1.85

F1, F2 Solder 10.6, 7.1 10.6, 7.1 0.076

G1,G2 Spacer: Cu 10.6, 7.1 10.6, 7.1 1.14

H1, H2 Solder 10.6, 7.1 10.6, 7.1 0.076

I1, I2 IGBT, Diode: Si 12.7, 10 12.7, 10 0.145

J1,J2 Solder 12.7, 10 12.7, 10 0.076

K Heat Spreader: Cu 21 40 1.85

L TIM: Grease 21 50 0.1

M Substrate: SiN 21 50 0.3

N TIM: Grease 21 50 0.1

O Heat sink: Al 30 50 0.6
3 

to generate the desired net UA (heat transfer coefficient times 

area) at a specified fluid temperature. To determine the 

maximum allowable IGBT heat flux for a given UA value, an 

iterative approach was used within “DesignXplorer” of 

“ANSYS Workbench.” The volumetric heat load of the IGBT 

was determined based on the targeted maximum temperature 

limit of the IGBT device. This temperature limit is referred to 

as the maximum junction temperature (Tj) throughout this 

paper. Once this maximum allowable heat load was 

determined, the process was performed again for another UA 

boundary condition. Two junction temperature limits of 150°C

and 175°C were used for the analysis. In addition, two coolant 

temperatures were used (70°C and 105°C). The 70°C coolant 

temperature is based on current coolant loops used in HEV 

applications, and 105°C is based on the upper temperature limit 

target from DOE’s APEEM R&D team [1]. Fig. 4 shows a 

temperature profile of the package for a sample heat load 

condition, while Fig. 5 shows the heat flow paths through the 

asymmetric package construction. Fig. 6 shows the temperature 

path through the package for the IGBT, where the large 

temperature differences occur across the thin thermal interface 

material (TIM) layers.

Figure 4. Sample of temperature distribution on ANSYS FEA model 

(coolant temperature 70°C, IGBT junction temperature 150°C, net effective 

UA applied to each side 10 W/K, IGBT heat flux 170.7 W/cm2

, IGBT to diode 

loss ratio 3:1).

Figure 5. Heat flow paths from ANSYS FEA model (coolant temperature 

70°C, IGBT junction temperature 150°C, net effective UA applied to each 

side 10W/K, IGBT heat flux 170.7 W/cm2

, IGBT to diode loss ratio 3:1).

Figure 6. Staight line IGBT temperature profile through package to the heat 

sink base from ANSYS FEA model. Note the z axes scale is magnified to 

show detail (coolant temperature 70°C, IGBT junction temperature 150°C, net 

effective UA applied to each side 10 W/K, IGBT heat flux 170.7 W/cm2

, 

IGBT to diode loss ratio 3:1).

The parametric ability of the FEA thermal model allows for 

sensitivity studies associated with multiple aspects of the 

package design and not just the heat exchanger performance. 

For this demonstration, the package performance curves were 

generated using different material properties for the TIM. A 

baseline TIM material was selected from information generated

as part of NREL’s characterization of TIM thermal 

performance [8]. To see the impact of improvements to the 

thermal interfaces, a 5X and 10X improvement in the interface 

thermal resistances were considered (Table II). 

B. Heat Exchanger Characterization − Finned

As shown at the top of Fig. 1, another step in the process 

required characterizing the thermal performance of various 

thermal control technologies. This can be performed through 

computational fluid dynamics (CFD) analyses, analytical 

solutions, and experimental correlations. Each of these methods 

will be discussed in this paper. 

The first selected thermal control technique approximates 

the actual finned heat exchanger used in a vehicle application. 

The approximated geometry is shown in Fig. 7, and the key 

parameters are listed in Table III. The fluid flow assumptions 

are highlighted in Table IV. The coolant fluid was assumed to 

be a 50/50 (by mass) mix of water ethylene glycol with 

properties based on [9]. The actual power electronics system 

includes 12 stages of power modules placed between 13 

cooling channels running between the inlet and outlet 

manifolds. For this analysis, we assumed that each stage 

received the same coolant flow on each side of the package so 

the total flow rate was divided by 12 instead of 13. 

TABLE II

TIM PROPERTIES

Thermal 

Conductivity 

(W/m-K)

Thickness 

(mm)

Thermal 

Resistance 

(mm2-K/W)

Base Case 1.5 0.10 66.7

5X 7.5 0.10 13.3

10X 15 0.10 6.7

TABLE III 

FIN PARAMETERS

Parameter Definition Value

a Fluid channel width 0.9 mm

b Fin or channel height 2.0 mm

t Fin thickness 0.6 mm

l Channel length 50.0 mm

Figure 7. Heat exchanger geometry for one side of the module package.
4 

TABLE IV

FLUID FLOW ASSUMPTIONS

Definition Value 

Total system volumetric flow [2,3] 12 L/min

Coolant inlet temperature 70°C

Cooling stages 12

Number of cooling channels per side 20

Channel mass flow rate 0.00043 kg/s

The finned heat exchanger performance was estimated 

using two different methods. The first involved running a CFD

simulation of a single channel; the second involved an 

analytical solution for fully developed laminar flow through a 

rectangular cross section [10]. For the CFD simulation, a model 

of a single channel was developed using the ANSYS software 

programs “Workbench” and “CFX.” An entry region was 

applied to the model to ensure fully developed fluid flow, 

although the model did not enforce a fully developed thermal 

entry region. Convergence was confirmed along with mesh 

independence. The average heat transfer coefficient over the 

total cooled surface was determined according to equation 1, 

where q is the heat transferred to the fluid, As is the total 

surface area cooled by the fluid, Ts is the average surface 

temperature at the interface, and Tref is the inlet temperature of 

the fluid. While the heat transfer coefficient was of interest, the 

parameter needed for the analysis was the overall UA of the 

heat exchanger. This can be determined from the fin surface 

temperature and area (UAs) as shown in equation 2, or from the 

base temperature and area before the fins (UAb) as shown in 

equation 3; Tb is the average base surface temperature. The 

base area corresponds to the area in the ANSYS model (Fig. 4) 

to which the UA value was applied. The UAb results were used 

for this analysis. The results were multiplied by n (number of 

channels) to get an estimate for the total heat exchanger 

performance. 

. (1)

. (2)

(3)

The second approach used to evaluate the performance of 

the finned heat exchanger was to use available analytical 

solutions for fully developed laminar flow through rectangular 

channels [10]. Using these techniques, approximate values for

the Nusselt number (NuD) and the Darcy/Moody friction factor 

(f) multiplied by Reynolds number (ReD) were estimated. For 

this analysis, a uniform surface temperature was assumed. 

Based on the geometry in Table III, NuD was fixed at a value of 

3.68 and fReD was fixed at 65. The average heat transfer 

coefficient through the channel was calculated according to 

equation 4, where k is the fluid thermal conductivity and Dh is 

the hydraulic diameter of the channel.

 . (4) 

As mentioned earlier, the effective UA of the heat exchanger 

was of interest. The UA for the full heat exchanger was 

determined by equation 5, where n is the total number of 

channels, ηo is the surface efficiency, and As is the total cooled 

surface area of the channel. The surface efficiency was 

estimated by treating the fin as a standard straight rectangular 

fin [10]. 

 . (5) 

In addition to the UA value, it was also of interest to 

evaluate the pressure drop through the channel. The pressure 

drop due to the channel is determined based on equation 6, 

where l is the channel length, r is the fluid density, and um is 

the mean fluid velocity through the channel [10, 11]. Equation 

6 considers only the channel pressure drop, but entry and exit 

loss coefficients could also be included. These were excluded 

for this analysis in order to compare the results to the pressure 

drop from the CFD results.

. (6) 

C. Heat Exchanger Characterization − Jet

The approach to characterize the finned heat exchanger 

used analytical solutions and CFD results. The third approach 

described in this paper provides an example of using 

correlations based on experimental work. For this example, 

experimental work performed at the NREL on jet cooling was

used [12]. NREL researchers characterized the performance of 

steady and self-oscillating single-phase jets in submerged and 

free surface conditions. This analysis focused on the data 

associated with the steady jet in the submerged condition. The 

target distance between the jet outlet and the impingement 

surface was 1.1 mm. The jet velocity was 2 m/s, 7 m/s, and 12

m/s. The jet diameter was fixed at 1.24 mm, the coolant fluid 

was water at an inlet temperature of 25°C, and the heated target 

area was 1.27 cm2

. 

Since the data are based on water at 25°C it was necessary 

to estimate the jet performance for a 50% by mass mixture of 

water ethylene glycol at the desired temperatures of 70°C and 

105°C. For this reason, the reported performance data were 

converted into a generalized dimensionless form based on 

Reynolds number (ReD) and Colburn factor (jH) using 

equations 7 though 9, where St is the Stanton number, Pr is the 

Prandtl number, cp is the fluid specific heat, and μ is the fluid 

dynamic viscosity. The results are plotted in Fig. 8 along with a

curve fit of the data correlating the Colburn factor to Reynolds 

number. Using the correlation, it is possible to back out 

estimates for the heat transfer coefficient with jets using water 

ethylene glycol assuming a desired jet velocity. In this analysis, 

the mean jet velocity was kept consistent with the experimental 

results. An exhaustive approach was not used to refine the jet 

correlation. The intent was to show the process and obtain 

approximate values using jets with water ethylene glycol.
5 

Figure 8. Experimental results for submerged single-phase jet (target 

distance 1.1 mm, jet diametert 1.24 mm, water inlet temperature 25°C).

. (7)

. (8)

. (9)

D. Heat Exchanger Effectiveness-NTU Analysis

When evaluating the performance of heat exchangers, it is 

important not to neglect the impact of the coolant fluid 

properties and flow rate. Often, simplifications are used that 

approximate the thermal resistance of the heat exchanger as 

simply 1/UA, but this can lead to inaccurate estimates of heat 

exchanger performance [13]. For this reason, the performance 

of the previously discussed cooling channels will be compared 

in terms of the effectiveness-NTU (number of transfer units) 

thermal resistance shown in equation 10, where is the heat 

exchanger effectiveness and is the fluid mass flow rate. The 

heat exchanger effectiveness is calculated from the NTU, as 

shown in equations 11 and 12 (see [10] and [13] for details). 

. (10)

. (11)

(12)

Fig. 9 compares actual thermal resistance RNTU against the 

simplified ideal thermal resistance (1/UA) across a range of 

UA values. As UA increases, the actual thermal resistance 

departs from the ideal case, and as expected, it is dependent on 

the flow rate for a specified fluid. For example, the results for 

12 L/min depart from the ideal case for UA greater than 10 

W/K.

III. RESULTS

A. Package Thermal Characterization

As shown in Fig. 1, an initial requirement to evaluate the 

system performance of a power semiconductor package is an 

understanding of the package specific thermal performance. 

The package specific thermal performance curves are shown in 

Fig. 10 for the selected package. The curves show the package 

thermal performance against a range of potential heat 

exchanger performance characteristics. Specifically, the curves 

show the maximum allowable heat flux through the IGBT vs. 

heat exchanger thermal resistance based on the specified 

temperature limits. These curves were generated by iterating on 

the volumetric heat generation of the IGBT and diode until the 

specified peak IGBT junction temperature limit was reached. A 

3:1 ratio was used for the IGBT to diode losses.

Three conditions were developed based on different 

temperature specifications for the inlet coolant and the 

maximum allowable junction temperature. The baseline curve 

shows the thermal performance, assuming a 70°C inlet coolant 

temperature and an IGBT junction temperature limit of 150°C. 

The second curve represents a 105°C coolant temperature and 

150°C junction temperature, while the third represents a 70°C 

coolant temperature and 175°C junction temperature limit. Two 

key points to notice are how the curves begin to flatten at low 

heat exchanger thermal resistance values and, as expected, the

105°C coolant temperature causes a drop in achievable heat 

flux. 

Figure 9. Comparison of actual and ideal (1/UA) thermal resistance as a 

function of UA for 50% by mass mixture of water ethylene glycol at 70°C at 

three flow rates.

Figure 10. Package specific thermal performance curves showing IGBT heat 

flux vs. heat exchanger thermal resistance at three operating conditions.
6 

In this analysis, the different temperature specifications 

were modeled in the 3D FEA model. It is also possible to 

estimate the impact of different temperature specifications 

without running all combinations, which is important in 

analyzing the impact of alternative coolant temperatures and

the effects of increasing component temperature limits. To 

determine the impact of alternative temperature specifications 

on a fixed package, the baseline model results must first be 

generated. The result for the case of a 70°C inlet coolant 

temperature and an IGBT junction temperature limit of 150°C 

(shown in Fig. 10) was selected as the baseline. For each 

operating point, it is possible to estimate the total thermal 

resistance from junction to fluid (Rtot) according to equation 13, 

where Tj is the allowable junction temperature limit, Tc,i is the 

coolant inlet temperature, and qtot is the total heat applied to the 

package (IGBT + diode). It is then possible to estimate the 

IGBT heat flux (q”IGBT) assuming different temperature 

specifications following equations 14 through 16. A fixed loss 

ratio of 3:1 between the IGBT and diode is assumed, and 

ASi,IGBT is the silicon area of the IGBT device.

. (13) 

. (14) 

. (15) 

. (16) 

Fig. 11 shows a comparison of the results. The estimated

values based on equations 13 through 16 match the actual 

simulated results from the 3D FEA model. The agreement 

between results gives confidence in the ability to investigate 

multiple temperature specification options without the need to 

run numerous 3D FEA simulations.

It is also important to evaluate the impact of potential 

changes to the package configuration. For example, it is also 

possible to look at the sensitivity of performance to other 

factors, such as material selection. As an example, the TIM 

thermal properties were adjusted as shown in Table II. Fig. 12

shows the results. For high heat exchanger thermal resistance 

values (Rhx), the heat exchanger is the bottleneck, and package 

thermal improvements become less effective. The curves show 

that different package configurations with different package 

thermal resistance characteristics approach the same value of 

heat flux as Rhx increases. It is necessary to decrease Rhx to get 

the benefit of improved thermal performance packages with 

reduced thermal resistance. As Rhx decreases, the package 

curves level off at different rates based on the package thermal 

performance.

B. Heat Exchanger Characterization

With the information generated for the package thermal 

performance, the next step required understanding the 

performance of the potential heat exchanger technologies. As 

highlighted previously, this work looked at the performance of 

an aluminum fin heat exchanger and a jet impingement heat 

exchanger. Table V highlights the performance of the finned 

heat exchanger for the operating conditions in Tables III and 

IV. Table V compares the results of the CFD analysis with the 

analytical solution. The results are relatively close given the 

assumptions applied to the analytical solution. Specifically, the 

difference in the heat transfer coefficient is expected because 

the analytical solution assumed a fully developed thermal 

region and uniform temperature. However, the pressure drops 

are close because both solutions are based on fully developed 

flow through the channel.

Figure 11. Comparison of simulated and estimated performance curves for 

alternative temperature specifications.

Figure 12. Comparison package performance with improved thermal interface 

material. 

TABLE V 

FIN RESULTS COMPARISON

Parameter CFX Analytical Value

h (W/m2-K) 1877 1241

nhA (W/K) 10.89 7.2

UAb (W/K) 10.38 6.98

∆P (Pa) 267 271

 (W/K) 0.11 0.16
7 

The estimated convection heat transfer coefficients for the 

jet impingement cooling are provided in Fig. 13 and were 

generated following the previously developed correlation (Fig. 

8). Fig. 13 shows the estimated performance of a submerged 

steady jet at 70°C and 105°C using a 50% by mass mixture of 

water ethylene glycol. The experimental convection heat 

transfer coefficient values were based on an average 

temperature over a heated area, so the results would not be the 

same if the temperatures were averaged over a different area. 

The specified convection coefficient values are tied to a 

specific area. 

For this analysis, it was assumed that the jet would not 

provide a high convection coefficient over the entire heat 

exchanger surface used in Fig. 2 and Fig. 4 (30 mm by 50 mm). 

For this reason, the selected convection coefficient for one jet

was assumed to be over a 1.27 cm2 area. Because of this 

smaller area, two jets were applied to each side of the package, 

which doubles the total system flow rate if the flow rate 

through a single jet was kept constant at 0.5 L/min. This is 

shown in Table VI for the assumed system flow rate. Using 

1.27 cm2 represents the low-end thermal performance, because 

additional area on the heat exchanger would experience some 

cooling. For this reason, a doubling of the effective cooling 

area was also considered with the intent to show the potential 

range of performance. A more detailed investigation would 

need to be performed to determine the actual area averaged 

convection coefficient using jet cooling for this specific 

package since, for jet cooling, the local convection coefficient 

decreases away from the jet impingement area. 

Table VI shows the performance of the jet cooling at the 

two effective area assumptions. One can see that the idealized 

thermal resistance (1/UA) is close to the effectiveness-NTU 

thermal resistance even at the higher UA values because of the 

higher flow rate. This is confirmed by Fig. 9. The results are 

shown at an inlet coolant temperature of 70°C and 105°C.

C. System Performance

With the generated package performance curves and the 

completed heat exchanger analysis, it was possible to combine 

the two results to the look at the total system performance as 

shown in the general process diagram of Fig. 1. This was done 

by overlaying the heat exchanger performance in terms of its 

effectiveness-NTU thermal resistance on top of the generated 

package thermal performance curves, as shown in Fig. 14. This 

allowed a comparison of the combined system and an 

understanding of how improvements in the heat exchanger 

performance affect the system performance. Fig. 14 highlights 

four package performance curves. The first is the baseline 

system with the base TIM, inlet coolant temperature of 70°C, 

and a maximum junction temperature of 150°C. The second 

curve is the same package with an inlet coolant temperature of 

105°C. The third curve is the same as the baseline system, 

except the TIM thermal resistance is reduced by a factor of 10. 

Finally, the fourth performance curve included the 10X 

improved thermal interface material and an inlet coolant 

temperature of 105°C. The heat exchanger performance is 

shown in Fig. 14 by the vertical lines. Three representative heat 

exchangers are shown. The first is the double jet at the lower 

limit of the effective heat exchanger area. As can be seen, its 

Figure 13. Comparison of convection heat transfer coefficient vs. flow rate for 

water at 25°C and jets using a 50% by mass mixture of water ethylene glycol 

at 70°C and 105°C.

Figure 14. System performance showing package thermal performance curves 

with estimated heat exchanger performance. 

TABLE VI 

JET RESULTS ESTIMATE

Two Jets per Side

Two Jets per Side

2X Area 

Enhancement

System flow [L/min] 24 24 24 24

Tc,i [°C] 70 105 70 105

Flow per Jet [L/min] 0.5 0.5 0.5 0.5

h [W/m2-K] 30000 37000 30000 37000

Cooled Area [m2] 2.53E-04 2.53E-04 5.07E-04 5.07E-04

UA [W/K] 7.60 9.37 15.20 18.75

1/UA [K/W] 0.132 0.107 0.066 0.053

 [K/W] 0.140 0.115 0.074 0.062

 

thermal performance is close to the baseline finned heat 

exchanger at a 70°C inlet coolant temperature. Although not 

shown, the thermal resistance of the finned heat exchanger at 
8 

105°C is similar. Finally, the third heat exchanger shows the 

potential of a two times improvement in the heat exchanger 

area for the double jet system.

In the case of the baseline system, one can see that the 

benefit of increasing the heat exchanger performance beyond 

the fin design is of limited value because of the limitation 

related to the package thermal performance. This is because of 

the leveling off of the curve at lower Rhx values. Fig. 14 also 

shows the impact of increasing the coolant temperature. While 

the system at 70°C coolant is capable of operating at about 165 

W/cm2

, this is decreased to 93 W/cm2 when forced to operate 

with a coolant temperature of 105°C (a drop of 44%). This 

assumes similar performance for the finned heat exchanger at 

70°C and 105°C. As shown in Fig. 14, the package with the 

improved TIM is able to take advantage of a higher 

performance heat exchanger. 

Finally, since the selected package is part of an existing 

HEV application, the estimated system performance from this 

analysis was compared against publicly available information. 

For this comparison, the total thermal resistance (Rtot) from 

junction to fluid was calculated using equation 13 with the fluid 

temperature at 70°C and junction temperature at 150°C. The 

total thermal resistance at the estimated finned heat exchanger 

thermal resistance (0.11 K/W) is 0.219 K/W. For comparison, 

the published maximum thermal resistance of the package is 

0.203 K/W [2,3]. This is within 8% of the published value.

IV. CONCLUSIONS

This paper describes a method of integrating techniques for 

characterizing power semiconductor thermal performance and

heat exchanger technologies. The methodology enables an 

analysis of the system performance and the trade-offs of a 

combined cooling system. A production power semiconductor 

package capable of double-sided cooling was selected to 

demonstrate the process and analysis techniques. Package 

specific thermal performance curves were generated comparing 

the achievable IGBT heat flux to a specified heat exchanger 

thermal resistance operating at specified temperatures. Methods 

for using fluid and thermal data from analytical solutions, CFD 

results, and experimental correlations were illustrated, with 

their application to standard heat exchanger analysis 

techniques. The importance of not relying on the ideal thermal 

resistance (1/UA) based only on convection was also 

demonstrated. Even for fully developed laminar flow, the 

selection of the mass flow rate is important in terms of heat 

exchanger effectiveness, which affects the error of the 

simplified 1/UA thermal resistance model. 

It is important to note the importance of matching the heat 

exchanger thermal performance (Rhs) with the package thermal 

performance. The optimal heat transfer mechanism is highly 

dependent on the package configuration, which leads to the 

need to evaluate the package and heat exchanger as an 

integrated system. The combined analysis ensures that one 

area of the thermal control system is not overdesigned, adding 

unnecessary cost, weight, and volume to the system. The work 

described in this paper illustrates the need for an integrated 

system analysis on new innovative power semiconductor 

package configurations, and proposes techniques for 

understanding the trade-offs associated with an integrated 

system. 

ACKNOWLEDGMENT

We thank Tim Burress of Oak Ridge National Laboratory 

for providing information about the selected package 

construction. We also acknowledge the support of Sreekant 

Narumanchi and Lawrence Chaney of NREL. 

REFERENCES

[1] Electrical and Electronics Technical Team Roadmap, FreedomCAR and 

Fuels Partnership, November 2006, 

www.eere.energy.gov/vehiclesandfuels/pdfs/program/eett_roadmap.pdf. 

[2] Y. Sakai, H. Ishiyama, and T. Kikuchi, “Power control unit for high 

power hybrid system,” SAE 2007 World Congress, Detroit, MI, April 

16-19, 2007, SAE Paper 2007-01-0271.

[3] H. Yasui, H. Ishiyama, M. Inagaki, K. Mamitsu, and T. Kikuchi, 

“Power control unit for high power hybrid system,” Proceedings of the 

EVS 23 conference, Anaheim, CA, Dec. 2-5, 2007.

[4] J. Hsu, R. Staunton, and M. Starke, “Barriers to the Application of HighTemperature Coolants in Hybrid Electric Vehicles,” Oak Ridge National 

Laboratory Technical Report ORNL/TM-2006/514, 2006.

[5] K. Kelly et al.,“An overview of thermal control technologies for cooling 

electric vehicle power electronics,” Proceedings of the EVS 23 

conference, Anaheim, CA, Dec. 2-5, 2007.

[6] M. O’Keefe and K. Bennion, “A comparison of hybrid electric vehicle 

power electronics cooling options,” Vehicle Power and Propulsion 

Conference, Sept. 9-12, 2007.

[7] T. Burress et al.,“Evaluation of the 2008 Lexus LS 600H Hybrid 

Synergy Drive System,” Oak Ridge National Laboratory Technical 

Report ORNL/TM-2008/185, 2009.

[8] S. Narumanchi, M. Mihalic, K. Kelly, and G. Eesley, "Thermal interface 

materials for power electronics applications," ITHERM 2008, May 28-

31, 2008. 

[9] K. Alshamani, “Equations for physical properties of automotive 

coolants,” SAE 2003 World Congress, March 3-6, 2003, SAE Paper 

2003-01-0532.

[10] F. Incropera and D. DeWitt, Fundamentals of Heat and Mass Transfer, 

4th ed. John Wiley and Sons, New York, 1996.

[11] B. Munson, D. Young, and T. Okiishi, Fundamentals of Fluid 

Mechanics, 3rd ed., John Wiley and Sons, New York, 1998.

[12] S. Narumanchi, K. Kelly, M. Mihalic, S. Gopalan, R. Hester, and A. 

Vlahinos, "Single-phase self-oscillating jets for enhanced heat transfer,"

Semi-Therm 2008, March 16-20, 2008.

[13] R. Moffat, "Modeling air-cooled heat sinks as heat exchangers," SEMITHERM 2007, March 18-22, 2007.
F1147-E(10/2008) 

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Rapid Modeling of Power Electronics Thermal Management 

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14. ABSTRACT (Maximum 200 Words)

A methodology was developed to rapidly evaluate trade-offs associated with alternative packaging configurations and 

thermal management technologies for power electronics packaging. The methodology includes the integration of 

available experimental correlations, computational fluid dynamics results, parametric 3D finite element analysis (FEA) 

thermal models, and established heat exchanger analysis techniques. The parametric 3D FEA model enables 

sensitivity studies related to the power module package configuration and cooling technologies. This paper focuses 

on the study of alternative cooling technologies as they are applied to a fixed power module package. The 

methodology is applied to a double-sided power module package for several alternative cooling technologies.

15. SUBJECT TERMS

power electronics; thermal management; thermal control; thermal resistance; hybrid electric vehicles; electric vehicle; 

fuel cell vehicles 

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