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Numerical Study of Oil Jet Cooling in Electric
Traction Motors with Hairpin Windings
Waruna Maddumage
School of Physics, Engineering
and Computer Science
University of Hertfordshire
Hatfield, UK
w.maddumage@herts.ac.uk
Alasdair Cairns
Department of Mechanical Materials
and Manufacturing Engineering
University of Nottingham
Nottingham, UK
alasdair.cairns1@nottingham.ac.uk
Amin Paykani
School of Physics, Engineering
and Computer Science
University of Hertfordshire
Hatfield, UK
a.paykani@herts.ac.uk
Abstract—Hairpin winding technology, combined with direct
oil jet impingement cooling, is a viable solution known to increase
volumetric power density and efficiency in the next generation
of traction motors. However, the coolant fluid interaction with
the complex winding geometry has not yet been fully examined;
specifically, with the use of high-fidelity CFD simulations. Thus,
the present work investigates the radial and axial oil impinging
jets in a hairpin winding motor using multi-phase simulation. The
study first analyses power losses and temperature distribution
of the motor under the Worldwide Harmonised Light Vehicle
Test Procedure (WLTP). Next, the performance of axial and
radial jet impingement is numerically analysed by considering
the fluid flow. Finally, the two configurations are compared in
terms of their oil film formation rate. The results indicate that
the maximum power losses observed in typical driving conditions
are considerably lower than the maximum losses predicted for
the complete operational region of the motor. Moreover, the axial
impinging jet shows a higher oil film formation rate compared
to a radial jet impingement configuration within the examined
conditions.
Index Terms—Permanent magnet motor, oil cooling, hairpin
end winding, numerical modelling, jet impingement, thermal
design.
I. INTRODUCTION
The main requirements for electric traction motors such as
higher volumetric power densities, reduced weights, and lower
costs have led to the development of electric motors with
increased current densities as well as high electromagnetic
and thermal loads compared to typical motors. Heat generation
from the electric motor is mainly due to energy losses in the
windings and iron core. High thermal loads in the motor can
damage the winding insulation of the armature and demag-
netize the field magnets of the rotor in permanent magnet
motors [1]. Additionally, the electric resistivity of the copper
windings increases with temperature, resulting in an increase
in thermal energy generated by the Joule effect.
There are two main research directions for heat dissipation:
low-loss motors and thermal management systems. Designing
a low-loss motor to reduce heat generation is difficult to
implement due to inherent losses and technical limitations.
A practical approach to dissipating heat is using an accurate
thermal management system. Over the last decade, various
forced liquid cooling methods have been utilised in electric
vehicle (EV) traction motors. In early EVs such as the 2010
Toyota Prius and 2012 Nissan Leaf, housing cooling jackets
were employed to cool the traction motor as part of the thermal
design [2]. But the current generation of EVs is heavily
moving towards high power density, high current, and high-
efficient traction motors. Using only a cooling jacket to achieve
the desired level of thermal management is unrealistic as it
would require a larger motor housing, resulting in a motor with
higher volume and weight. Moreover, internal heat transfer
from the rotor and stator to the outer cooling jacket using
conduction alone is not highly efficient. Therefore, modern EV
traction motors increasingly utilise advanced cooling methods
or hybrid thermal designs, particularly when considering the
stator.
An interesting recent thermal design concept is end-winding
cooling. End-windings are generally identified as machine
hotspots and effective thermal management is critical [2]–[4].
Particularly, when considering high-speed and high-density
traction motors with hairpin winding stators, effective thermal
management is crucial. Hairpin winding motors are becom-
ing increasingly attractive for EV applications due to their
numerous advantages, such as increased slot fill factor of
up to 0.75 compared to the 0.4-0.5 achievable with typical
windings [5], simpler manufacturing process, greater current
density, superior thermal performance, shorter end-winding,
reduced DC electrical resistance, and cheaper production time
and cost [2]. As a result, automobile manufacturers, such
as Chevrolet-Volt, have already started applying this winding
design to their latest line of electric motors [6].
In practice, oil jet impingement cooling is often coupled
with hairpin winding motors to cool the end-winding directly.
Toyota introduced the hairpin end-winding and direct oil
impingement cooling technology in their current Toyota Prius.
Toyota replaced the old P410 engine with P610, which uses
hairpin end-windings and oil impingement cooling. The new
design offers a smaller motor with greater efficiency than the
P410. The new architecture has reduced the motor size by
35% and increased power density by 36% with 20% fewer
losses [7]. The combination of hairpin end-windings and oil
impingement cooling as a cooling system offers a possibility
for developing new EV motor designs with a more compact
structure and greater power density.
Direct oil impingement cooling offers several advantages
compared to other cooling methods such as housing jacket and
slot channel cooling. It has been shown to improve axial heat
transfer in the slot, achieve efficient heat extraction, and reduce
the heat transfer distance from the end-windings without
causing additional electromagnetic fluctuation or electric noise
[2], [8]. These benefits make it a preferred choice, especially
for traction motors.
The direct oil impingement cooling of hairpin end-windings
has attracted numerous studies in the past few years as a
promising cooling approach for EV traction motors. Liu et
al. [9] analysed the oil jet cooling of hairpin winding using an
oil jet distribution bar mounted on top of the end-winding.
The oil jet distributor was designed with a series of jet
nozzles to impinge oil onto the end-winding. It was shown
that the proposed configuration limits heat extraction to the
upper regions of the winding. In another study by Liu et
al. [10] a model calibration workflow for determining the heat
transfer coefficient of oil jet cooling with a hairpin winding
is proposed. The model can be used for cooling performance
estimation of similar setups. In a study conducted by Gronwald
et al. [11] the use of flat jet nozzles for end-winding cooling
is examined with empirical models and experimental data.
The empirical model proposed is capable of estimating heat
transfer between the oil jet and winding with a maximum error
of 6%. In a recent study by Chen et al. [12] the cooling effect
of a single jet nozzle placed on top of the end-winding was
investigated using three different oil flow configurations. These
configurations included the generic end-winding and two other
configurations where an external and internal surface was
constructed along the end-winding to facilitate coolant flow.
Results showed that the internal surface configuration provided
the highest average and maximum heat transfer. However,
these attempts mostly followed an experimental approach. To
the best of the authors’ knowledge, none of the attempts
proposed high-fidelity numerical models with non-simplified
geometries of the hairpin end-winding to study multi-phase
jet impingement designs. Furthermore, existing attempts do
not consider the effect of axial jet impingement configuration.
The aim of the present study is to investigate the impact of
oil jet impingement with axial and radial configurations using
numerical modelling techniques. Firstly, an internal permanent
magnet (IPM) motor is chosen as a case study. Secondly, the
power losses of the traction motor and temperature distribution
of the electric machine are analysed. Thirdly, CFD numerical
simulations are conducted for axial and radial jet impingement
thermal designs. Finally, two thermal designs are evaluated
based on the surface area covered by injected oil. The study
provides valuable insights by examining power losses on a
real-life drive cycle for a traction motor, proposing a high-
fidelity modelling methodology for multi-phase jet impinge-
ment, and the comparison of fluid flow with the two proposed
thermal configurations (i.e., axial and radial impinging jet).
II. CA SE S TU DY
A. Motor specifications
The study aims to analyse the thermal performance of an
Internal Permanent Magnet (IPM) synchronous machine with
a maximum speed of 12000 rpm and power of 150 kW. The
motor uses a 6-layer hairpin winding for the stator with an
8-pole and 48-slot topology.
B. Existing thermal design
The original thermal design of the electric machine consists
of an axial housing jacket and shaft cooling. The hollow shaft
and housing jacket are cooled with an inlet temperature of
65◦C. Fig. 1 shows the axial view of the existing cooling
system.
Motor housing
Shaft cooling
Housing
Water jacket
End-windings
Shaft
Rotor
Stator core
Fig. 1. Schematic of the existing cooling strategy.
C. Proposed jet impingement design
The present study proposes two jet impingement cooling
configurations as outlined below. Both thermal designs consist
of 12 impinging oil jets around the end-winding. A simple
orifice jet nozzle is used with lubricating synthetic oil as the
coolant liquid. A recent study by Chen et al. [12] has shown
that the use of restriction surfaces around the end-winding
can significantly increase the heat extraction capability with
jet impingement cooling. Therefore, both proposed thermal
designs in this study use fluid flow restriction surfaces to
improve the contact area between the fluid and winding.
1) Axial jet impingement: The proposed axial jet impinge-
ment configuration is shown in Fig 2. The fluid flow is
restricted using two flow restrictions surfaces: internal and
external surfaces.
Stator core
End-windings
Oil jet
Motor housing
External
surface
Internal
surface
Fig. 2. Schematic of the proposed axial jet impingement cooling strategy.
2) Radial jet impingement: The proposed radial jet im-
pingement configuration is shown in Fig 3. The flow is
restricted using an internal surface.
End-windings
Oil jet
Motor
housing
Internal
surface
Stator core
Fig. 3. Schematic of the proposed radial jet impingement cooling strategy.
III. MOD EL LI NG A PP ROAC H
In this research work, we used ANSYS Motor-CAD v15
and Fluent 2022 R1 to analyse the power losses of the electric
machine and predict the cooling performance of the proposed
impinging jet thermal designs. The details of the modelling
approach are given in the following sections.
A. Electromagnetic analysis model
The electromagnetic characteristics of the selected IPM
machine are analysed using Motor-CAD software prior to
investigating the proposed thermal design. The analysis of
the electromagnetic behaviour is used to predict the power
losses of the electric machine and determine the thermal
characteristics under operation.
The power losses of the machine are calculated using the
maximum torque per amp (MTPA) control strategy consider-
ing several current levels with a maximum value of 500 ARMS
for the speed range 0 to 12000 rpm. The power losses of the
motor are examined to predict the thermal characteristics of
the stator end-windings. The motor power losses are analysed
over the WLTP drive cycle and an operating point is selected
for detailed analysis. In calculating the power losses of the
electric machine, the present study assumes that the airgap
windage and bearing frictional losses as negligible.
Before analysing motor power loss, the Motor-CAD model
used is analysed, especially the methods related to stator power
losses. The AC copper losses of the stator were calculated
using the hybrid finite element (FE) model [2], [3] for the drive
cycle analysis. The model uses a combination of analytical and
finite element methods to predict AC copper losses. Prior to
calculations, the model was calibrated using the inbuilt full
FEA model. A lumped parameter thermal network (LPTN)
model was used to calculate the thermal behaviour of the
machine. The reliability of the 3D LPTN model results of
the stator slots was examined using the in-build FE model.
B. Numerical model
We carried out the numerical modelling of the proposed
thermal design using ANSYS Fluent commercial CFD soft-
ware package. In the present study, the computational domain
is restricted to a 30◦stator section of the electric machine.
The fluid flow in the motor is limited to the end-winding by
using flow restriction surfaces. Thus, only the end-winding and
a section of the stator wall are considered for the numerical
model. We did not consider variables such as motor rotation
and rotor surfaces.
Considering the random nature of the fluid flow, a transient
calculation is conducted. The two phases, air and oil are
modelled as a multi-phase flow using the volume of fluid
(VOF) method. Fluid flow is solved using the realizable k−ε
turbulence model with the scaled wall function enabled. A
poly-hexcore mesh with boundary layers is adopted for the
study. Moreover, to reduce the number of cells in the mesh,
automatic mesh adaption is used. Numerical calculations are
performed with Intel Xeon Gold 6130 CPU processors. The
core hours used for the simulations of axial and radial jets are
30,249 and 31,452 core hours, respectively.
The oil jet impingement nozzle is modelled with a diameter
of 3 mm, where the oil jet is injected at an initial velocity of 20
m/s. A synthetic lubricant oil (BP Turbo oil 2389) as suggested
in [13] is used as the coolant for the proposed thermal design
analysis.
Fig. 4 shows the computational domains of the proposed
two oil jet cooling configurations. At this point of the study,
only the nozzle in the 12 O’clock position is examined. Since
the nozzle is situated in the topmost position, fluid flow from
the remaining nozzles into the defined computational domain
is not considered.
Fig. 4. Computational domain of the proposed oil jet cooling configurations:
(a) axial impinging jet; (b) radial impinging jet.
In the axial jet model, two sides are defined as outlets to
ensure realistic fluid flow while all other surfaces are defined
as walls. However in the radial model, three sides: two sides
and the front wall is defined as outlets, and the remaining
surfaces are defined as walls.
IV. POW ER L OSS ANA LYSIS
The cooling strategy for the electric machine is designed
for automotive applications, here and thus, its power losses
on a typical drive cycle are an essential factor to consider.
To understand the operational behaviour of the electric motor
when used in an EV, we implemented a vehicle model using
the GM Chevy Bolt as the reference. The existing 150 kW
permanent magnet drive motor of the Chevy Bolt is replaced
with the selected IPM motor. The vehicle model is simulated
over the WLTP class 3 drive cycle. Parameters selected for the
longitudinal vehicle models can be found in [14].
Fig. 5 shows the power losses predicted when simulated
over the drive cycle. Two points in the drive cycle are ob-
served: maximum total power loss point and maximum stator
copper loss point. For clarity, only a section of the drive cycle
is shown. Maximum loss points are shown by the red vertical
dash lines.
1500 1550 1600 1650 1700 1750 1800
Time (s)
0
500
1000
1500
Power loss (W)
Total Stator copper Stator iron Max. total loss
(a)
800 850 900 950 1000
Time (s)
0
500
1000
1500
Power loss (W)
Total Stator copper Stator iron Max. copper
(b)
Fig. 5. Power losses of the motor on WLTP class 3 drive cycle: (a) Maximum
total power loss; (b) Maximum stator copper loss.
The stator copper and iron losses are the most prominent
for the drive cycle operating points. Specifically, the stator
copper losses and iron losses accounted for 36% and 58% of
the total losses respectively. We observed the operating point
with the highest total power losses at 1567s, which did not
coincide with the point of maximum stator copper loss. The
latter occurred at 962s and accounted for 74% of the total
losses.
The thermal behaviour of the end-winding is mainly gov-
erned by the stator copper losses. Thus, stator copper power
losses corresponding to the operating points of the motor when
run on the WLTP class 3 drive cycle are analysed. Fig. 6 shows
the copper losses of the electric machine over its torque curve.
The red triangles indicate the operating points of the motor
within the stator copper loss map.
12600
10800
9000
7200
5400
3600
0
Speed (rpm)
Shaft torque (Nm)
350
300
250
200
150
100
50
0
102008200620042002200
(2518 rpm, 126 Nm)
200
Copper power loss (W)
1800
Fig. 6. Copper power losses of the motor (Drive cycle operating points are
shown in red triangles).
The stator copper losses are found to increase with higher
motor torques and speeds, as shown in the results. However,
the motor operating points during the drive cycle are below
34% and 68% of maximum motor torque and speed, respec-
tively. The maximum stator copper loss point is predicted at
34% and 20% of maximum torque and speed, respectively.
Since the drive cycle operating points are predicated on low
to medium regions of the torque map, the stator copper losses
are significantly lower than the maximum copper losses of the
motor. Compared to the maximum copper loss of the motor for
its entire operational region, the maximum copper loss point
of the drive cycle is predicted at only 4.4%.
Understanding the temperature distribution of the motor
with the existing thermal design is essential before analysing
the proposed design. Thus, the thermal characteristics of the
motor are calculated at the maximum copper loss point of the
drive cycle. Fig. 7 presents the axial schematic of the design
object and temperature behaviour with the existing thermal
design.
Temp. (OC)
111.0
103.6
96.2
88.8
81.4
74.0
Fig. 7. Axial view of the temperature distribution at drive cycle maximum
copper loss point with the existing thermal design.
The electric machine is cooled via the cooling jacket and
hollow rotor shaft, providing heat extraction from the centre
as well as outside of the machine. However, the thermal
distribution analysis shows that heat extraction of the stator
end-winding needs to be improved compared to the rotor,
stator core and shaft. The highest temperature values were
predicted for the stator winding. The average temperature
of the end-winding is 8% and 12% higher than the average
temperature of the stator core and rotor magnets respectively.
V. PRO PO SE D TH ER MA L DE SI GN A NALYSI S
Heat extraction from the hairpin winding is carried out by
directly injecting oil onto the end-winding. Jet impingement
cooling is known to provide excellent heat extraction capa-
bilities. For instance, oil jet cooling on a continuous surface
is capable of achieving a heat transfer coefficient of 8000
W/m2k with a jet velocity of 7.5 m/s [15]. But, the hairpin
end-windings have inherent air gaps and a complex geometry,
which considerably decreases the ability of the impinging jet
to wet the target surface. For instance, a radial jet without
any flow restrictions is capable of only achieving an average
heat transfer coefficient of 300 W/m2k [12]. The thermal
performance of the impinging jet is influenced by the flow
fluid flow pattern on the target surface. Thus, understanding
oil film formation on the end-winding is critical for the thermal
designs of hairpin motor jet cooling.
In the present study, two jet configurations, namely axial
and radial impinging jets are analysed considering the oil
flow characteristics. The heat transfer capability of the oil is
proportional to the contact area of the oil and the coil [16].
An accurate prediction of the thermal extraction capability of
different jet impingement configurations can be achieved by
analysing the oil and coil surface area.
The oil jet considered is a simple orifice nozzle with a 0◦
spray angle. Thus, the oil jet will concentrate on the winding
surface directly below the nozzle. A high inlet velocity of
20 m/s is used to increase the splash of the impinging jet.
Fig. 8 shows the two impinging jet configurations at 0.003s
flow time. The oil flow is shown using an iso-surface defined
at 0.5 oil volume fraction. The flow restriction surfaces and
walls are shown as transparent surfaces.
Fig. 8. Oil splash effect of the impinging jet (for clarity an early time step,
0.003s is shown): (a) axial impinging jet; (b) radial impinging jet.
The fluid film formation can be affected by the jet inlet
velocity. Previous studies have shown that a low splash effect
is present when the low inlet velocity is low, such as 3
m/s [16]. However, as shown in the figure above, the splash
effect is prominent in both configurations due to the high
inlet velocity of the impinging jet. The increased splash effect
enables the fluid flow to go through the air gaps present in the
end-winding, covering a wider surface area.
In addition to the oil film created on the winding, the oil
splash on the stator wall may contribute to the overall heat
extraction. Therefore, the oil film on the stator wall was also
examined. Fig. 9 shows the oil film layer coverage after 0.03s.
The oil film formation on the stator wall and end-winding is
shown using the volume fraction of oil found on the surfaces.
Visualising the fluid flow becomes increasingly difficult as
the fluid flow develops with time due to the complex geometry
and splash effect. To accurately compare the flow development
of the two configurations, oil film formation is quantified. An
oil film formation factor is defined by the area of the oil film
formed and the total area of the end-winding and stator wall.
The area of the oil film is calculated by plotting a volume
fraction of the oil contour on the winding and stator wall. The
oil film factor of the winding Afcoil and stator wall Af wall
are defined as,
(a) (b)
Volume fraction of oil
0.25 0.75 10.50
Fig. 9. Oil film formation on the end-winding and stator wall: (a) axial
impinging jet; (b) radial impinging jet.
Afcoil =Aoilfilm /A winding (1)
Afwall =Aoilf ilm/A wall (2)
where Awinding and Awall are the surface area of the end-
winding and stator wall within the considered computational
domain, respectively. The Aoilfilm represents the contact area
between the oil and the winding/ stator wall. The fluid flow
formation with time on the winding and stator wall is shown
in Fig. 10. To understand the oil film formation rate, the
derivative of the oil film factor is calculated at each data point.
The oil film formation rate is defined as,
rate n= (f ilmf actor n+1 −f ilmfactor n−1)/∆t(3)
The axial impinging jet shows a higher oil film coverage
factor on the winding throughout the examined time frame. At
0.3s axial jet shows a 50% winding oil film coverage compared
to 33% in the radial jet configuration. The peak oil film
formation rate on the winding with the axial jet is 93% higher
than the radial jet. In contrast, the radial jet performs better
when considering the stator wall. The peak oil film formation
rate on the stator wall is 162% higher with the radial jet
compared to the axial jet. However, to understand the cooling
performance thermal conductivity of the surfaces should also
be considered. The winding has a thermal conductivity of 401
W/mK while the stator wall thermal conductivity is only 30
W/mK. Since the thermal conductivity of the winding is 13
times higher than the stator wall, heat extraction capability
from the end-winding is significantly higher than the stator
wall.
VI. CONCLUSION
In this paper, the thermal performance of an impinging
axial and radial jet was analysed considering the fluid flow
characteristics. We proposed jet impingement thermal designs
for a 150kW IPM electric machine with a housing jacket and
shaft cooling. Before examining the proposed thermal designs,
we evaluated the power losses of the motor over a WLTP class
0 0.005 0.01 0.015 0.02 0.025 0.03
Flow time (s)
0
0.2
0.4
0.6
0.8
1
Winding - Oil film factor
0
20
40
60
80
Winding - Oil film
formation rate (s-1)
Oil film factor - Axial jet
Oil film factor - Radial jet
Oil fim fomation rate - Axial jet
Oil fim formation rate - Radial jet
(a)
0 0.005 0.01 0.015 0.02 0.025 0.03
Flow time (s)
0
0.2
0.4
0.6
0.8
1
Stator wall - Oil film factor
0
25
50
75
100
125
150
175
Stator wall - Oil film
formation rate (s-1)
Oil film factor - Axial jet
Oil film factor - Radial jet
Oil fim fomation rate - Axial jet
Oil fim formation rate - Radial jet
(b)
Fig. 10. Oil film spread on the end-winding and stator wall: (a) end-winding;
(b) stator wall.
3 drive cycle and its thermal behaviour at maximum copper
loss point.
The drive cycle operating points were observed on the low
and mid regions of the torque curve. The maximum drive cycle
copper loss was only 4.4% of the maximum copper loss of
the motor. Also, for the maximum drive cycle copper loss
point, stator winding showed the highest temperature. These
findings indicate that it is important to consider the power loss
behaviour over expected drive cycles to optimise the cooling
strategy in designing the motor thermal management system.
The increase in oil jet inlet velocity led to a higher splash
effect, resulting in an increased overall fluid film surface area.
The axial impinging jet performed better in terms of oil film
formation rate on the winding, while the radial jet showed a
higher oil film formation rate on the stator wall. But, since
the thermal conductivity of the winding is 13 times higher
than that of the stator wall, higher thermal performance can
be expected with the axial jet. Therefore, it is important
to consider both the oil film formation rate and thermal
conductivity when designing the motor cooling system.
The future research will further examine both configurations
considering the heat transfer coefficient. Also, the numerical
models will be validated using experimental data and theoret-
ical results.
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