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ABSTRACT
One of the deliverables for the GM/DOE sponsored
EcoCAR2 competition involved the design and validation of
an energy storage system (ESS) that could withstand 20g of
acceleration in the longitudinal direction, 20g of acceleration
in the lateral direction, and 8g of acceleration in the vertical
direction with a minimum safety factor of 2, in the event of a
crash. Rose-Hulman Institute of Technology (RHIT) elected
to base their energy storage system off of A123 battery
modules (7×15s2p) and components. The design included a
thermal analysis for various drive cycles and a mechanical
analysis of the enclosure built to support and protect the
battery modules. The thermal analysis investigated passive
cooling versus active cooling and, after identifying active
cooling as the best strategy, an appropriately sized cooling
loop was developed. The mechanical analysis involved the
use of Siemens NX7.5 to develop CAD models for the ESS
enclosure components. These enclosure components included
the battery base plate (interface that allows for module
mounting), the battery casing (surrounds the modules) and
the battery top plate (results in a completely contained
system). The mechanical analysis of the ESS also required
the use of NASTRAN to perform finite element analysis
(FEA) on the module-to-base plate mounting and the pack-to-
vehicle mounting. The pack-to-vehicle mounting was
achieved by mounting the ESS to a subframe that will be
welded into the vehicle.
INTRODUCTION
Utilizing data, technical support, and physical components
from A123 Systems, a Lithium-Ion Energy Storage System
(ESS) was developed. This paper looks at the components of
the design of a Lithium Ion Energy Storage System including,
thermal management, safety considerations and verification
through finite element analysis.
THERMAL ANALYSIS AND DESIGN
To begin investigation of the thermal aspects of the proposed
A123 7 × 15s × 2p battery pack, it was first necessary to
obtain the thermal capacity of the pack. Utilizing
manufacturer provided data for temperature rise with respect
to maximum continuous discharge, a first principles approach
was used to estimate the thermal capacity as a function of
temperature [1]. An average thermal capacity of 102,000 J/°C
was used while the extreme values will be investigated via a
sensitivity analysis [2].
Given a thermal capacity estimate and assuming an insulated
model, the amount of heat the battery pack can retain for a
specified temperature change was estimated by integrating
Equation 1. Thus,
Equation 1
The maximum recommended battery temperature is 50°C
with de-rating occurring between 50 °C - 60 °C [2]. At 60°C,
the battery control system shuts down the battery pack.
Assuming the battery starts at a room temperature (R.T.) of
21 °C (70 °F) and soaks in the sunshine for a substantial
amount of time, a temperature differential of 24 °C can be
reached. This temperature differential yields a maximum heat
energy retention value of 2,450 kJ. It is important to note that
the competition “room temperature” value can vary widely
from as low as 15 °C (assuming the vehicle soaks outside,
overnight, at the Milford Proving Grounds (MPG)) to as high
as 39 °C (should the vehicle get stuck soaking outside during
the day at the Desert Proving Grounds (DPG)). These ranges
will be discussed further in the sensitivity analysis.
Knowing the amount of heat the battery can retain while
staying within defined operating limits enabled the team to
Design of a High Voltage Lithium Ion Energy
Storage System
2013-01-0564
Published
04/08/2013
Laura Nash, Jonathan Nibert, Zachariah Chambers and Marc Herniter
Rose-Hulman Institute of Technology
Copyright © 2013 SAE International
doi:10.4271/2013-01-0564
THIS DOCUMENT IS PROTECTED BY U.S. AND INTERNATIONAL COPYRIGHT.
It may not be reproduced, stored in a retrieval system, distributed or transmitted, in whole or in part, in any form or by any means.
Downloaded from SAE International by Laura Nash, Wednesday, March 27, 2013 01:49:36 AM

assess thermal response on a drive cycle. For the Argonne
National Laboratory (ANL) 4-cycle and 0-60 mph drive
cycles the Rose-Hulman Institute of Technology (RHIT)
proposed vehicle architecture incurred the following battery
heat generation rates and total heat energies [Table 1- Drive
Cycle Analysis
Table 1. Drive Cycle Analysis
The ANL 4-Cycle data is especially critical to RHIT because
it can be used to predict performance for the most critical
competition event - the Emissions and Energy Consumption
(E&EC). The E&EC includes distances (legs) of 20, 40, and
100 miles which are designed to showcase both vehicle
modes, charge-depleting and charge sustaining (CD and CS
respectively). With a 4-Cycle charge depleting range of 36.1
miles (based on the ANL 4-Cycle), the following heat
retentions for the E&EC event are presented in Table 2-
E&EC Heat Energies.
Table 2. E&EC Heat Energies
It is assumed that each leg of the E&EC will be started with a
fully charged battery. If not, the heat retention will be less
because the CS mode induces less heat than the CD mode.
Therefore, the values presented represent a worst-case
scenario for heat retention values for the E&EC event.
The 0-60 mph event is very important to RHIT because the
outstanding acceleration specified by the Vehicle Technical
Specifications (VTS) arises from the engine and motor
operating in parallel for limited bursts; therefore, having a
fully charged battery is critical to this event. While CS heat
retention values are presented, this event will only be run in
CS mode as a last resort. With a heat retention value of 12 kJ
per run in CD mode, the maximum number of passes before
de-rating occurs, can be predicted.
To look briefly at the sensitivity of the data, Table 3-ESS
Temperature Rise is presented below for a range of thermal
capacities and battery initial temperatures (i.e. room
temperatures):
Table 3. ESS Temperature Rise
From the sensitivity analysis, it is clear that the worst case
scenario (lowest thermal capacity, highest ambient
temperature) will put the battery into a de-rating situation, 10
km, 4- Cycle schedule, while, the best case scenario (highest
thermal capacity, lowest ambient temperature) shows only a
6.5 °C temperature rise over the CD 4-Cycle. This suggests
that the 100 mile E&EC leg could be completed without
entering a derate condition. The burst of high power
conditions for the 0-60 mph acceleration causes a very small
amount of heat to be generated and a near negligible rise in
temperature.
From the results, it appears that passively cooling the ESS
would be acceptable for nearly all competition events and
temperatures excluding the desert hot soak extreme.
However, these results assume that the battery pack has been
given sufficient time to passively cool back to room
temperature. Realistically, the team could go from fully
heating the battery with the 0-60mph competition to the first
leg of E&EC with only a few hours in between. Therefore, an
estimate of passive cooling time was required.
Neglecting radiative effects and assuming natural convection,
we have Equation 2:
Equation 2
where the heat transfer coefficient (h) ranges from 1- 20
W/(m2 K) for pure natural convection with air [1]. With a
surface area of 1.24 square meters the time required to cool
the pack from 45 °C to a variety of ambient room
temperatures for the three thermal capacities and two heat
transfer coefficients are presented below.

Examining the results of, it is clear that passively dissipating
heat from the battery will have a negative impact on the
performance at competition events due to the extreme amount
of time required to cool the pack back to ambient
temperature. Additional concerns arise from on-road
convective heating of the battery due to the high track
temperatures at the DPG, which can exceed 130 °F. While
insulating the battery could mitigate this additional source of
battery heating, it would essentially eliminate passive heat
rejection between events. Thus, some form of active cooling
is required.
Following A123 recommendations that air-cooling the pack
is slightly better than passive cooling, a liquid-to-air heat
exchanger will be employed. A preliminary design review
with A123 recommended the cooling plates be mounted on
the bottom of the modules and for RHIT to utilize a
conductive isolation pad between the plates and modules to
maximize surface contact. To investigate the pad and cooling
plate system, a thermal circuit analysis will be performed to
estimate the amount of heat that can be dissipated by the
system. A quasi-steady state first principal analysis will be
used to estimate the temperature rise of the liquid.
For the ANL 4-cycle schedule, the previously described
cooling system was modeled in Simulink and loaded with the
heat generation data.
After the team selected active liquid cooling for the ESS, a
cooling loop was developed which uses a water/glycol mix.
The ESS cooling loop (Appendix A), consists of the ESS's
internal cooling plates, a sealed reservoir/overflow tank, a
drain point, inline pump, and liquid-air heat exchanger. The
cooling loop will be fully self-contained in the rear of the
vehicle, co-located with the ESS.
The ESS cooling system will use a water pump. In order to
select a pump, the pressure drop across the cooling plates and
radiator need to be quantified to ensure an appropriate
pressure could be attained by the pump. The system was
designed for a 2 gallon per minute (GPM) flow rate, at which
the cooling plates each have a 7 psi drop, and the radiator a 3
psi pressure drop. With 7 individual cooling plates in series,
the total cold plate drop is ∼49 psi. Adding in the radiator,
the overall loop pressure drop is ∼52 psi. See Appendix B.
Liquid Cooling System Modeling
In order to properly size components for the cooling system
and verify it would keep the ESS temperature within
acceptable limits, a model of the previously derived cooling
system was generated in Simulink. The top-level model, as
shown in Appendix C, simulates the connection between the
ESS's cooling plates and the liquid-air heat exchanger. The
exit fluid temperature value is passed between the
components, while the initial ESS and fluid temperatures are
inputs along with the ESS heating power and the ambient air
temperature.
Looking into the ESS cooling plate model (Appendix C), the
thermal circuit as defined in Appendix A is modeled. The
module thermal capacity that was derived earlier was used in
the model, and the thermal resistances of both the cooling
plates and thermal interface pad were taken from the
manufacturer's technical specifications [4]. The heat
generation data was again based upon an A123-supplied
figure of 97% efficiency combined with the average power
discharge.
Since the thermal resistance of the cooling plates is a function
of the liquid flow rate, and the radiator's thermal resistance is
a function of both liquid and air flow rates, both of these
parameters were swept to aid in identifying optimal operating
points. To run the sweep, a worst-case operating scenario was
established where the initial battery and fluid temperatures
were 41 °C and the ambient air temperature (the air being
drawn across the radiator) was 38 °C. These values are based
upon reference material citing the average summer high for
Yuma, Arizona as 41°C and assume that the vehicle would
have equalized to that value, and the ambient air would be
cabin air, cooled to 3° C below outside temperature by the
HVAC system.
This worst-case scenario was then run with a heating value
from a 100% State of Charge (SOC) depletion charge-

depleting mode drive cycle, since it will incur the largest
power draw. The resulting final ESS temperature versus air
and liquid flow rates is given in Appendix D-1.
Based on these surfaces, an air flow rate of approximately
480 cubic feet per minute (CFM) and a liquid flow rate of 8
liters per minute (LPM) was selected, as these points were
where the returns in terms of increased flow rates diminished
greatly.
With these flow rates specified, an analytic sweep was then
run to find the impact both the initial ESS/fluid temperature
and the ambient air temperature had on the final ESS
temperature. Both temperatures were swept from a lower
bound of 25° C (assumed R. T.) to 41° C (worst-case). Again,
the heating value was based upon a 100% SOC depletion CD
mode drive cycle. The resulting 3D plot is given in Appendix
D-2.
Based on this plot, for a worst-case scenario where both the
initial temperatures and ambient are 41° C, the final ESS
temperature is 50.1° C. For the expected case of an initial
temperature of 26° C (near R.T. soak) and 36° C ambient, the
final temperature is 41° C.
Beyond the analysis of the 100% CD draw-down, the
temperature rise during a subsequent CS operation was also
analyzed. Two distinct cases were run; the worst-case where
ambient is 41° C and the initial temperatures were the final
values from the worst-case in the prior analysis and an
assumed case where the ambient was 36° C and the initial
temperatures were the final values from the assumed case in
the prior run. In this way, the analysis would mimic actual
vehicle operation where the vehicle would operate in CD
mode for a given amount of time, then switch into a CS mode
for the remainder of the operating time. The results of the CS
mode analysis showed that in the worst-case, the ESS
temperature converged to 52° C, and in the assumed case, the
ESS temperature ended at 47° C.
The analysis results indicate that the cooling system will keep
the ESS below the thermal de-rating zone during all operation
for the assumed case, and will just barely enter de-rating for
the worst-case. Even in the worst-case scenario, the ESS
temperature stabilized at a value in the lower end of the de-
rating zone, and never rose to the absolute cutoff temperature
limit.
Additionally, it should be noted that the CD heating values
were based on a CD run that saw a 100% SOC depletion. The
100% to 0% SOC swing of that analysis is worse than the
actual operation is expected to be. In practice, upper and
lower SOC bounds will be in place which will limit the CD
operation to a narrower SOC band. This in turn will limit the
amount of time the vehicle is exposed to the higher CD
heating power, thereby lowering the effective final
temperature at the end of CD operation.
STRUCTURAL ANALYSIS AND
DESIGN
According to EcoCAR2 regulations, components of the ESS
must be analyzed under specific conditions in order for the
ESS design to be considered safe to install in the vehicle. A
structurally validated ESS involves a complete analysis of the
pack-to-vehicle mounting and a complete analysis of the
module-to-baseplate mounting with a minimum factor of
safety of 2 for all results.
ESS Mounting (Battery Pack-to-Vehicle)
The ESS will be secured to the vehicle by mounting it to a
subframe. The design material for the ESS Subframe is
1080steel, a material readily available from McMaster-Carr.
The 1080 tube steel (3/4 × 3/4 × 1/8 wall thickness) that will
be used to fabricate the ESS subframe is listed as having a
yield strength of approximately 350 MPa. This yield strength
will be used in conjunction with the analysis results to
determine if the design meets the required factor of safety of
2. The SIEMENS 7.5 NX model of the ESS subframe is
presented below. The NX NASTRAN analysis uses a
simplified beam model to represent the ESS subframe.
Figure 1. ESS subframe Bolt Pattern
Loading Conditions
The EcoCAR2 regulations call for components to be designed
to withstand a 20g longitudinal, 20g lateral, and 8g vertical
set of loading conditions unless otherwise specified. In order
to perform the finite element analysis (FEA) using NX 7.5
NASTRAN, equivalent forces needed to be calculated. For
the analysis, the battery pack mass was found by taking the
7×15s2p total mass of 137.2 kg [2] and adding in the mass of
the enclosure, which was estimated to be 49.81 kg. This
rendered an overall pack mass of 187.01 kg. This mass was

then used in Equation 3, along with a value of 9.8 m/s2 for
gravity, to find the translational forces imparted by each
module for the 20g and 8g cases, using respective values of
20 and 8 for the value of C.
Equation 3
20g Longitudinal Case
In order to account for the 20g longitudinal applied force
(Fpack) being applied at the center of the battery pack we
need to find the equivalent longitudinal forces that the 12 bolt
holes of the subframe will experience (Fpack,eq).
Setting up an equivalent system and rearranging, we obtain
Equation 4.
Equation 4
Using Equation 3 we can calculate the value of Fpack.
Using the mass of the pack from above as 187 kg,
acceleration due to gravity as 9.8 m/s2 and C as 20 we obtain
Fpack. Plugging this into Equation 4, we obtain Fpack,eq.
A summary of the calculated longitudinal forces is given in
Table 4- Pack Mounting Longitudinal Forces.
Figure 2. NX NASTRAN Longitudinal Loading
Additionally, since Fpack acts at the center of gravity (CG) of
the battery pack and not in line with the bolt holes in the
subframe, a moment is induced by Fpack which must be
accounted for by vertical equivalent forces at each of the bolt
holes.
Figure 3. ESS Longitudinal Loading
Figure 4. Longitudinal loading distances and equivalent
forces
Using Figure 3- ESS Longitudinal Loading and Figure 4-
Longitudinal loading distances and equivalent forces, an
equivalent system was set-up as given in Equation 5.
Equation 5
We can solve for Fm,eq knowing: h = 132mm, l = 273mm, x1
= 58mm, x2 = 174mm, and Fpack = 36,652 N. Therefore, the
equivalent force due to the induced moment at each bolt hole
is 2,395 N. This force is applied to each bolt hole in the
appropriate vertical up/down direction as shown in Figure 5 -
NX NASTRAN Moment Induced Equivalent Forces.
Table 4. Pack Mounting Longitudinal Forces

Figure 5. NX NASTRAN Moment Induced Equivalent
Forces
20g Lateral Case
The lateral case was approached in the same manner as the
longitudinal case where an equivalent system was set up to
determine the lateral force at each bolt hole.
Setting up an equivalent system and rearranging, we obtain
Equation 6.
Equation 6
Using Equation 3 we can calculate the value of Fpack
Given the values are the same as the longitudinal case we
obtain Fpack = 3,054 N.
Figure 6. NX NASTRAN Lateral Loading
Additionally, since Fpack acts at the CG of the battery pack
and not in line with the bolt holes in the subframe, a moment
is induced by Fpack and must be accounted for by vertical
equivalent forces at each of the bolt holes.
Figure 7. ESS Lateral Loading FBD
Figure 8. Lateral loading distances and equivalent forces
Using Figure 7 - ESS Lateral Loading FBD and Figure 8-
Lateral loading distances and equivalent forces, an equivalent
system was set-up as given as shown in Equation 7.
Equation 7
We can solve for Fm,eq knowing: h = 132mm, w = 423mm, a
= 105mm, b = 174mm, and Fpack = 36,652 N. Therefore, the
equivalent force due to the induced moment at each bolt hole
is 1,205 N. This force is applied to each bolt hole in the
appropriate vertical up/down direction as shown in Figure 9 -
NX NASTRAN Moment Induced Forces.
Figure 9. NX NASTRAN Moment Induced Forces
Loading and Constraints
Since the geometry of the ESS Subframe is based on a fully
enclosed rectangle, the entire outer perimeter was constrained
to account for the fact that when the part is integrated into the

vehicle, all sides will be welded to the existing vehicle
structure.
20g Longitudinal Case
A total of 24 forces were applied to the ESS Subframe for the
longitudinal case. Twelve of these forces were applied in the
longitudinal direction (Fpack,eq) and twelve were applied in
the vertical direction(Fm,eq) to account for the moment
induced by Fpack being applied at the center of the battery
pack.
Each of the 12 bolt holes in the subframe experience
longitudinal forces of 3,054 N and vertical forces of 2,395 N.
The subframe was constrained and loaded accordingly and a
representation of the applied forces for the longitudinal case
are given in Figure 10- ESS Subframe 20g Longitudinal
Loading.
Figure 10. ESS Subframe 20g Longitudinal Loading
20g Lateral Case
A total of 24 forces were applied to the lateral loading case
for the ESS subframe. Each of the 12 bolt holes experience a
lateral force (Fm,eq) of 3,054N and a vertical force (Fm,eq) of
1,205 N. The subframe was once again constrained and
loaded accordingly and a representation of the applied forces
for the lateral case are shown below.
Figure 11. ESS Subframe 20g Lateral Loading
8g Vertical Case
In the vertical case the only force to be applied to each of the
12 bolt holes is Fpack/12. This is once again calculated using
Equation 3. In this case there is no induced moment because
there is no perpendicular distance.
Figure 12. ESS Subframe 8g Vertical Loading
FEA Results and Discussion of Results
All three cases (20g-longitudinal, 20g-lateral, and 8g vertical)
were analyzed in NASTRAN according to the loading
conditions discussed earlier. A summary of maximum stress
and maximum deflection for each case is presented below.
Table 5. Summary NX NASTRAN results
As shown above, the highest elemental stress was seen in the
20g-longitudinal case at 141 MPa. Given the yield strength of
the chosen material (350 MPa), a stress of 141 MPa meets the
required factor of safety of 2. The maximum deflection of
1.41 mm was also seen in the 20g longitudinal case. This
amount of deflection is well within the acceptable range of
values. NASTRAN results are shown below.

ESS Mounting (Battery Module-to-
Enclosure Base Plate)
In addition to performing analysis at the pack-level, FEA was
also performed on the locations where the battery modules
bolt to the enclosure base plate. The purpose of this analysis
was to verify that the module's mounting would not fail in the
event of the 20g and 8g loads the pack would be subjected to
in the pack-to-vehicle mounting analysis.
Loading Conditions
As previously stated, the EcoCAR2 regulations call for
components to be tested under a 20g longitudinal, 20g lateral,
and 8g vertical set of conditions unless otherwise specified.
In order to perform the FEA, equivalent forces needed to be
calculated. For the analysis, the module masses were found
by taking the supplied total mass of 101 kg for the 7×15s2p
configuration and dividing by seven to yield a per-module
mass of 14.42 kg. This mass was then used in Equation 8
along with a value of 9.8 for g to find the translational forces
imparted by each module for the 20g and 8g cases, using
respective values of 20 and 8 for C.
Equation 8
Since each module is secured with 4 bolts, it was assumed
that the force was distributed evenly amongst them, and so a
per- bolt force was calculated as Fmodule/4. A summary of
the calculated forces is given below in Table 6- Module
Mounting Translational Forces.
Table 6. Module Mounting Translational Forces
To illustrate the loading of the baseplate due to the module, a
free-body diagram of the module-baseplate (profile view)
loading is shown in Figure 13- Module Mounting Free-Body
Diagram

Figure 13. Module Mounting Free-Body Diagram
The CG was assumed as the geometric center of the module,
and Fmodule acts at that point. Since the module is
constrained where it bolts to the baseplate at B1 and B2, a
moment Mmodule is induced. The magnitude of Mmodule is
given by Equation 9, where ‘h’ is the distance from the
mounting point to the module CG.
Equation 9
Since the actual loading of the baseplate will occur at the bolt
mounting locations, Mmodule must be accounted for in the
loading. This is done by inducing a moment, Mload as shown
in Figure 14- Module Mounting Induced Moment.
Figure 14. Module Mounting Induced Moment
The moment will be induced by placing forces at the bolt
mounting locations. The magnitude of Mload based on these
forces is given by Equation 10
Equation 10
Since Mload must equal Mmodule, rearranging and solving
for Fload gives Equation 11.
Equation 11
Since from the side view, B1 and B2 both account for 2 bolts
each, a per-bolt force is found by dividing Fload by 2. The
same approach may be used to find the induced moment
forces for a side-loading of the module by substituting the
value of L for the width instead. A summary of the moment-
inducing forces is given in Table 7- Module Mounting
Moment Forces. From the provided module drawings the CG
height, h, is taken to be 121.52 mm, the center-to-center
lengthwise bolt spacing, L, is 259.11 mm, and the center-to-
center widthwise bolt spacing, W, is 164.81mm.
Loading and Constraints
Each of the module mounting areas were loaded with both the
translational and moment-inducing forces from the previous
section. Figure 2.38 shows an example of the loading for the
20g longitudinal case. For the lateral case, the Fbolt forces
were applied along the y-axis, and the moment-inducing
forces were also relocated so as to induce the moment in the
proper direction. In the case of the 8g vertical test, the Fbolt
forces were applied in the Z-axis, and no moment-inducing
forces were applied.
Figure 15. Longitudinal Loading
Example Module Loading
The actual loading for the 20g longitudinal case is given
below in Figure 16. The aforementioned forces are shown in
red, with the constrained geometry in blue. The constrained
portions were the 12 M10 bolt holes used for the pack
mounting bolt, and the 26 M5 bolt holes that are used to
mount the baseplate to the case.
Table 7. Module Mounting Moment Forces

Figure 16. ESS Baseplate NASTRAN Longitudinal
Loading
20g Longitudinal Loading
Results
The results from the FEA runs show a maximum stress
occurring in the 20g longitudinal case, with a peak stress of
114.10 MPa. The highest stresses for the 20g lateral and 8g
vertical cases were 90.13 and 30.76, respectively. The stress
gradient overlaid onto the model for the 20g longitudinal case
is shown below in Figure 17- NX NASTRAN Longitudinal
Results
Figure 17. NX NASTRAN Longitudinal Results
The resulting maximum stresses and their related factors of
safety are summarized in Table 8- Stress Summary. The
intended build material for the baseplate is quenched and
tempered 1080 steel, which has a yield strength of 350 MPa.
Overall, the summary shows that the lowest factor of safety
for the module-to-baseplate mounting using 4130 steel is
8.58, and for 1018 steel is 2.49. This shows that the baseplate
exceeds the minimum factor of safety of 2 as required by
EcoCAR2 regulations with both the intended 4130 steel, and
the alternative 1018 steel.
Table 8. Stress Summary
CONCLUSION
In conclusion, ‘Design of a High Voltage Lithium Ion Energy
Storage System’ demonstrated that the Energy Storage
System Rose-Hulman will implement into their modified
2013 Chevy Malibu will need to be actively liquid cooled in
order to meet competition requirements. Once collected data
revealed the need for active liquid cooling of the ESS, an
appropriate liquid cooling system was selected. Additionally,
this paper presented loading conditions, constraints, and
results for NASTRAN run FEA simulations which verified
that both the ESS subframe (welded into the vehicle) and the
ESS pack (securely bolted to subframe) could withstand the
required 20g longitudinal, 20g lateral, and 8g vertical loads.
REFERENCES
1. Cengel, Yunus. Heat Transfer: A Practical Approach.
New York, NY: McGraw-Hill, 1998.
2. Rutkowski, Brian, “Battery Sub-System Design
Specification: Interface Control Document,” A123
SYSTEMS, 2011.
3. SAE International Surface Vehicle Recommended
Practice, “Recommended Practice for Packaging of Electric
Vehicle Battery Modules,” SAE Standard J1797, Reaf. June
2008.
4. AAVID THERMALLOY http://www.aavid.com/product-
group/liquidcoldplates
CONTACT INFORMATION
Laura C. Nash
nashl@rose-hulman.edu
Jonathan W. Nibert
Nibertjw@rose-hulman.edu
Marc Herniter, Ph.D.
Herniter@rose-hulman.edu
Zachariah Chambers, Ph.D.
Chambez@rose-hulman.edu

ACKNOWLEDGMENTS
General Motors, Argonne National Laboratories, and the US
Department of Energy

APPENDIX A - COOLING LOOP
APPENDIX B - COOLING PLATE DATa
APPENDIX

APPENDIX C - SIMULINK MODELS

APPENDIX D
Figure 1. ESS Cooling System Flow Rate Sweeps

Figure 2. ESS Cooling System Ambient and Initial Temperature Sweeps
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