Cranfield mixte hybrid

GROUP DESIGN
PROJECT REPORT
A report detailing the work undertaken
to complete the Group Design Project 2017
Cranfield Mixte Hybrid
Edited by
BORJA BALLESTER SESÉ
JOSEP M CARBONELL OYONARTE
ROWAN CARSTENSEN
CARLOS GIGOSOS TAMARIZ
GUILLAUME GIRARD
KYLE MIDDLETON
SAM RAWCLIFFE
Cranfield University
Advanced Motorsport Engineering MSc
2017
Group 1

Group 1 10/05/17 Group Design Project Report
Declaration
We, the authors, herewith declare that we have produced this study without the prohibited assis-
tance of third parties and without making use of aids other than those specified. Reference and
acknowledgment, where necessary, has been made to the work of others. This work has not previ-
ously been presented in identical or similar form to any other British or foreign examination board.
Borja Ballester Sesé:
Carlos Gigosos Tamariz:
Guillaume Girard:
Josep M Carbonell Oyonarte:
Kyle Middleton:
Rowan Carstensen:
Sam Rawcliffe:
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List of Figures
1 The two circuits the car must compete on for the project . . . . . . . . . . . . . . . . 3
2 Mass sensitivity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6
3 BSFC sensitivity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6
4 CD sensitivity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7
5 Isometric view of the car geometry . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8
6 Heave sensitivity study results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
7 Pitch sensitivity study results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
8 Pressure field on the car surface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11
9 Cooling ducts’ streamlines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12
10 Wheels wakes’ streamlines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 13
11 Powertrain of the Cranfield Mixte Hybrid car . . . . . . . . . . . . . . . . . . . . . . 14
12 AVL Boost engine model and results . . . . . . . . . . . . . . . . . . . . . . . . . . . 16
13 Octane number required by the engine . . . . . . . . . . . . . . . . . . . . . . . . . . 16
14 Test bed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
15 Generator characteristics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
16 Simulation Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19
17 Simulation Results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20
18 Figures demonstrating the oil cooling system fitted to the car . . . . . . . . . . . . . 22
19 Sensitivities of Vertical Tyre Load . . . . . . . . . . . . . . . . . . . . . . . . . . . . 25
20 Camber and roll center curves vs. chassis roll . . . . . . . . . . . . . . . . . . . . . . 26
21 Front tyre camber curve from caster and kingpin angles . . . . . . . . . . . . . . . . 27
22 Required ride rate for downforce and dynamic ride heights . . . . . . . . . . . . . . . 27
23 A quarter car Simulink model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29
24 Front damper coefficient and frequency sweep . . . . . . . . . . . . . . . . . . . . . . 29
25 Rear damper coefficient and frequency sweep . . . . . . . . . . . . . . . . . . . . . . 30
26 Avon F3 pacejka model coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31
27 Sandwich construction illustration . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32
28 Three point bend testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
29 Three point bend test and simulation results . . . . . . . . . . . . . . . . . . . . . . . 34
30 Roll structure load application zones and safety area . . . . . . . . . . . . . . . . . . 34
31 Illustration of the main roll structure . . . . . . . . . . . . . . . . . . . . . . . . . . . 35
32 Main roll structure simulation and results . . . . . . . . . . . . . . . . . . . . . . . . 36
33 Front roll structure simulation and results . . . . . . . . . . . . . . . . . . . . . . . . 36
34 Energy storage structure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37
35 Energy storage crash box simulated acceleration results . . . . . . . . . . . . . . . . . 38
36 Composite chassis render . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38
37 Torsional stiffness target, simulation results and maximum load of the chassis . . . . 39
38 Maximum strain in the chassis under prescribed loads . . . . . . . . . . . . . . . . . 39
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39 Load applied to the rear wishbone under worst case conditions as shown in the full
vehicle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40
40 Effective stress in the wishbone under maximum cornering . . . . . . . . . . . . . . . 41
41 Chassis manufacturing techniques . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42
42 Motor model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47
43 Honda GXH50 datasheet extract . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 116
44 Circuits designed in the test bed . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 117
45 Testbed Honda GXH50 test results . . . . . . . . . . . . . . . . . . . . . . . . . . . . 118
46 Alternator efficiency curve [23] . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 119
47 Comparison of the model and testbed engine’s data . . . . . . . . . . . . . . . . . . . 119
48 Capacitor Sizing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 132
49 Fluid boundary layer on a flat plate [24] . . . . . . . . . . . . . . . . . . . . . . . . . 133
50 The inputs for the Simulink model used in the cooling estimates . . . . . . . . . . . . 135
51 Three point bend test specimen sample dimensions . . . . . . . . . . . . . . . . . . . 142
52 Three point bend test experimental results . . . . . . . . . . . . . . . . . . . . . . . . 142
53 Three point bend results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 143
54 Joining specimen diagrams . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 144
55 Adhesive in peel testing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145
56 Bolted configuration in lap shear testing . . . . . . . . . . . . . . . . . . . . . . . . . 145
57 Bolted configuration in lap shear testing . . . . . . . . . . . . . . . . . . . . . . . . . 146
58 Adhesive in peel simulation comparison . . . . . . . . . . . . . . . . . . . . . . . . . 146
59 Roll structure load application zones and safety area . . . . . . . . . . . . . . . . . . 147
60 Main roll structure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 148
61 Main roll structure simulation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 148
62 Main roll structure simulation and results . . . . . . . . . . . . . . . . . . . . . . . . 149
63 Front roll structure simulation image . . . . . . . . . . . . . . . . . . . . . . . . . . . 149
64 Front roll structure simulation and results . . . . . . . . . . . . . . . . . . . . . . . . 150
65 Energy storage crash box assembly . . . . . . . . . . . . . . . . . . . . . . . . . . . . 151
66 Energy storage crash box kinetic energy and acceleration simulation results . . . . . 152
67 Energy storage crash box placement inside the chassis . . . . . . . . . . . . . . . . . 152
68 Bottom chassis portion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 153
69 Top Chassis portion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 153
70 Seat chassis portion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 153
71 Section view of combined parts of the chassis . . . . . . . . . . . . . . . . . . . . . . 154
72 Chassis loading applications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 154
73 Maximum strain in the chassis under prescribed loads . . . . . . . . . . . . . . . . . 154
74 Maximum strain at critical points in the chassis under prescribed loads . . . . . . . . 155
75 Torsional stiffness target, simulation results and maximum load of the chassis . . . . 155
76 Final3PtBend . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 156
Advanced Motorsport Engineering MSc iv Cranfield University

77 Load applied to the rear wishbone under worst case conditions as shown in the full
vehicle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 157
78 Load applied to the rear wishbone under worst case conditions . . . . . . . . . . . . 157
79 Eﬀective stress in the wishbone under maximum cornering . . . . . . . . . . . . . . . 158
80 Nose cone as placed in the vehicle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 158
81 Nose cone simulation results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 187
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List of Tables
1 Sensitivities with respect to the total time . . . . . . . . . . . . . . . . . . . . . . . . 5
2 Results of the wing and no-wing cases. . . . . . . . . . . . . . . . . . . . . . . . . . . 8
3 Results of the two configurations tested. . . . . . . . . . . . . . . . . . . . . . . . . 9
4 Results of the definitive configuration (H50R0.5) . . . . . . . . . . . . . . . . . . . . 11
5 Main engine geometries . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15
6 Effects of the engine switching regime on fin dimensions and weight . . . . . . . . . . 23
7 Final Fin Configuration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23
8 Initial Conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24
9 Suspension Set-up 1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
10 Suspension Set-up 2 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
11 Damper Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 30
12 Groups 1 structures material choices . . . . . . . . . . . . . . . . . . . . . . . . . . . 138
13 Toray T700-SC-12K mechanical properties [29] . . . . . . . . . . . . . . . . . . . . . 140
14 SHD MT510 resin system mechanical properties [22] . . . . . . . . . . . . . . . . . . 140
15 Airex C71.75 core mechanical properties [27] . . . . . . . . . . . . . . . . . . . . . . . 140
16 Aluminium 5182 mechanical properties [30] . . . . . . . . . . . . . . . . . . . . . . . 141
17 Properties of manipulated isotropic material used to model the composite structure . 143
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Keywords
Series Hybrid, Testbed, Boxer, Structures, Composites, Finite Element Analysis, CFD, Drag, Eﬃ-
ciency, Spring, Damper, Anti-roll, Hillclimb, Simulation, Supercapacitor, DeltaWing, Star-CCM+,
Aerodynamics, Reynolds Number, Boundary Layer, Forced Convection.
Advanced Motorsport Engineering MSc vii Cranﬁeld University

Table of Nomenclature
L/D Aerodynamic efficiency -
L Characteristic length m
θ Diffuser angle degrees
CD Drag coefficient -
µ Friction coefficient -
A Frontal area m2
H Height of the car m
CL Lift coefficient -
Nu Nusselt Number -
Pr Prandtl Number -
hrc Roll centre height mm
ω Specific turbulence dissipation rate s−1
ms Sprung mass kg
hms Sprung mass height mm
k Thermal conductivity W/mK
ε Turbulence dissipation rate m2/s3
mu Unsprung mass kg
r Wheel radius mm
l Wheelbase mm
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Table of Abbreviations
BSFC Brake Speciﬁc Fuel Consumption
CAD Computer Aided Design
CAN Controller Area Network
CF Carbon Fiber
CFD Computation Fluid Dynamics
CoG Center of Gravity
CP Center of Pressure
DC Direct Current
ECU Engine Control Unit
F3 Formula 3
FE Finite Element
ICE Internal Combustion Engine
MSA Motor Sports Association
PU Polyurethane
RANS Reynolds Averaged Navier-Stokes
RON Research Octane Number
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Contents
1 The Challenge 3
2 Simulations and Sensitivities 5
3 Aerodynamics and Computational Fluid Dynamics 7
3.1 Heave and pitch sensitivity studies . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9
3.2 Final configuration results . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
4 Powertrain 14
4.1 Internal Combustion Engine . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15
4.2 Generator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
4.3 Energy Storage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18
4.4 Electric motors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18
4.5 Controllers and electrical power transformers . . . . . . . . . . . . . . . . . . . . . . 18
4.6 Simulations and optimisation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19
5 Cooling 21
5.1 Effect of Engine Switching . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 23
6 Vehicle Dynamics 24
6.1 Sensitivity Study . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24
6.2 Geometry . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 24
6.3 Non-Linear Spring Rates . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26
6.4 Anti-Roll System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
6.5 Damper Settings and Sweeps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
6.6 Tyres . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 29
7 Structures and materials 32
7.1 Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32
7.2 Testing and simulation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
7.2.1 Three point bend test . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
7.2.2 Joining . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
7.3 Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
7.3.1 Roll structure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
7.3.2 Energy storage structure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37
7.3.3 Chassis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 37
7.3.4 Rear wishbone . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40
7.4 Chassis Manufacturing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40
7.4.1 Ply layup design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 40
7.4.2 Tooling and manufacturing design . . . . . . . . . . . . . . . . . . . . . . . . 41
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8 Summary 43
8.1 Further Work . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43
A Presentation of the Simulink Model 47
B Aerodynamics and Computational Fluid Dynamics (CFD) Appendix 58
C BMW HP2 Datasheet 108
D Testbed validation 116
E Fuel Flow Sensor Appendix 120
F Hybrid Component Datasheets 127
G Skeleton Capacitor Calculation Sheet 132
H Cooling Appendix 133
I Vehicle Dynamics Appendix 136
J Structures Appendix 138
K Meeting Minutes Appendix 160
L Predicted Gantt Chart 188
M Actual Gantt Chart 189
N Review 1 190
O Review 2 205
Advanced Motorsport Engineering MSc xi Cranﬁeld University

Advanced Motorsport Engineering MSc Group Design Project
Abstract
The project described in this report was undertaken by Group 1 of the Advanced Motorsport
Engineering MSc Group Design Project 2017. The task set was to design a hybrid hillclimb car
to descend and then reascend the Goodwood and Shelsley Walsh hillclimb courses with a limited
amount of fuel. This report will show the design decisions taken to reach a final design, which
was able to do a total lap time of 87.52 seconds at Goodwood and 48.17 seconds at Shelsley
Walsh.
A final render of the Cranfield Mixte Hybrid
Advanced Motorsport Engineering MSc 1 Cranfield University

Acknowledgements
The team Cranﬁeld Mixte Hybrid would like to acknowledge the assistance provided by
Dr. Konstantinos Karantonis from Siemens Industry Software CD Ltd. His provision of Star-
CCM+ licenses to the team enhanced the approach for the Computational Fluid Dynamics
analysis. Our special appreciation to Mr. Ben Bowlby, creator of the Nissan DeltaWing, who
did not doubt to advise us with design approaches. His knowledge of the car concept was very
useful for the development of the project. We would also like to express our gratitude to Mr.
Timoteo Briet, whose technical advice with Star-CCM simulations and meshing operation were
profoundly helpful.
Relating to the structures element of this course, the team would also like to thank Jim Hurley,
and Ben Hopper, for their assistance in the composite material manufacturing and testing
portion of the project. Finally, we would like to oﬀer our special thanks to the whole faculty
of the MSc in Advanced Motorsport Engineering, with special mention to Dr. Kim Blackburn
and Dr. James Brighton who assisted us technically and academically during the Group Design
Project.

1 The Challenge
This project was undertaken by Group 1 of the Advanced Motorsport MSc Group Design Project,
where four teams were set the same design challenge in the motorsport context. The challenge was
set to design a car to be run on two hillclimb courses, Goodwood in Sussex and Shelsley Walsh in
Worcestershire. The car must descend the hill and then turn around and ascend it again in the
shortest possible time, not including the turnaround time at the bottom of the hill.
(a) A schematic of the Shelsley Walsh Hillclimb Course (b) A diagram of the Goodwood Hillclimb
as used during the Goodwood Festival of
Speed
Figure 1: The two circuits the car must compete on for the project
The project however became greatly more relevant to the contemporary motorsport context in
the realm of the powertrain rules. The car must start at the top of the hill, only with a limited
amount of fuel on board. 0.3kg of 95 RON petrol is allowed, unless a mechanical connection be-
tween engine and wheels was maintained above 20kph. Such designs were permitted to use use an
additional 0.1kg of petrol.

Further to this, only one internal combustion engine was allowed (albeit with a free number
of cylinders and displacement), however an unlimited number of electrical machines were allowed,
as was any kind of electrical storage technology. This group project therefore became especially
relevant to the modern motorsport landscape, where increased pressure to keep motorsport road
relevant has led to the addition of electrical energy generation and propulsion systems. It became
clear that the use of a conventional powertrain with no electrical augmentation would lead to a car
that either completed the course very slowly, or more likely would be unable to complete the duty
cycle at either hillclimb.The energy storage system must have the same energy in it at the start as
it does at the end, so any energy used to propel the car during the run must have been generated
during the run. All stored energy must be made safe within 20s of the car stopping also.
The use of active aerodynamic devices was prohibited, and the wheels and tyres had to be left
visible from a side elevation. The car had to comply with the hillclimb rules given by the Motor
Sports Association, and also adhere to the safety regulations also set out by the MSA.

2 Simulations and Sensitivities
The first step of the project was the evaluation of the performances of the car in both of the circuits,
in order to know the strengths and weaknesses of the different car departments. For that purpose,
a Matlab code was created and developed, simulating the powertrain system and optimising the
use of it within the circuits. The code had some parameters implemented, such as the mass of the
vehicle, the brake specific fuel consumption, drag and lift coefficients, or some wheel parameters
such as rolling resistance, the longitudinal µ or the maximum lateral acceleration. The results of
the simulations were the lap times, the engine switch (where the lift and coast strategy was defined)
and the optimum engine power required for that particular configuration.
Table 1: Sensitivities with respect to the total time
Mass 0.53 s per 10 kg
BSFC 1.09 s/(g/kW.h)
CD 2.53 s/(0.01*CD_unit)
In order to create a path to follow in terms of design efforts, a sensitivity study was done to
some of the parameters of the code: mass, BSFC, CD, CL, frontal area and µ. The last three
parameters did not show a significant sensitivity in the simulation results, while the first three gave
results worthy of further study. The results of these sensitivity studies on mass, BSFC and CD can
be seen in Figures 2, 3 and 4, respectively. The impact of these parameters in the laptime is shown
in the Table 1

Figure 2: Mass sensitivity
Figure 3: BSFC sensitivity

Figure 4: CD sensitivity
3 Aerodynamics and Computational Fluid Dynamics
The aerodynamic design target was to minimise drag as a result of the drag sensitivity study: a
reduction of 2.53 seconds of the total time by every 0.01 reduction of CDA product. Consequently,
the aerodynamic concept was similar to the original DeltaWing, one of the most aerodynamically
efficient race cars ever built [1]. Extended explanation of this design is given in the Aerodynamics
and Computational Fluid Dynamics (CFD) appendix. The centre of pressure (CP) of the car was
targeted to be at the same location than the centre of gravity (CoG) of the car, 70% of the wheelbase.
The external aerodynamic design of the race car is shown in Figure 5. A flat floor and diffuser
were placed between the rear wheels, creating downforce in a very efficient way. The rear wheels
were covered to reduce drag. In addition, the engine cover was designed to enhance flow attachment
and minimise the wake of the car, specifically reducing form drag. Form drag is generated by the
appearance of separated flow regions on the geometry. A wing was placed at the top of the main
roll structure (further referred as "mid-wing"), placed as high as possible as allowed by regulations
to ensure freestream conditions. The mid-wing increased the downforce created by the car with
little increase in drag, as shown in Table 2. There is a further explanation of the geometry in the
Aerodynamics and Computational Fluid Dynamics (CFD) appendix. As a result, the car was less
sensitive to underbody aerodynamic variations when going through a bump or a kerb, which the
car would experience during its duty cycle. However, the most important effect of the mid-wing
was the change in the CP position it produced, as seen in Table 2.
After undertaking the turbulence model choice and mesh sensitivity study (which are explained
in detail in the Aerodynamics and CFD appendix), all the CFD simulations results were obtained
using the following conditions:

Table 2: Results of the wing and no-wing cases.
Cases L/D CLA (m2) CDA (m2) CP position (% front) Frontal area (m2)
With mid-wing 1.48 0.84 0.57 22.8 1.02
Without mid-wing 1.27 0.62 0.49 7.87 0.93
Figure 5: Isometric view of the car geometry
• Rotating wheels
• Moving ground at a speed of 30 m/s
• K-ε realizable RANS turbulence model
• Steady ﬂow conditions
• Meshes of average 14 million cells with 10 control volumes and a control surface
• A y+=37 value and 10 prism layers with geometric progression from the car’s surface
• Density= 1.225 Kg/m3
• Airﬂow speed= 30 m/s
• The simulation domain enclosed half of the car, common procedure in similar CFD analysis
when the testing geometry is symmetric with respect to its half plane [2] [3]. The domain
dimensions were 7H in front of the car (being H the height of the car in meters), 19H behind
the car (in order to allow the wake to fully develop), 7H to the side of the car and 5H above
the car [4].

Two different configurations of the car were tested: the first configuration was set up with a ride
height of 40 millimetres and 0 degrees of rake angle (H40R0) and the second configuration was set
up with a ride height of 50 millimetres and a rake angle of 0.5 degrees (H50R0.5), both inside the
limits imposed by the regulations. The results of both configurations are shown in Table 3.
Table 3: Results of the two configurations tested.
H40R0 1.48 0.84 0.57 22.8 1.02
H50R0.5 1.88 1.10 0.58 27.04 1.01
The second configuration showed an improvement in aerodynamic efficiency (L/D) of 27.4%, an
increase in CLA of 31.5% while the CDA only increased by 3%. More important was the improvement
in the location of the centre of pressure, going from the original 22.8% at the front to the 27.04%
with the new configuration, much closer to the 30% targeted.
3.1 Heave and pitch sensitivity studies
Heave and pitch sensitivity studies were done for each simulation in order to assess the performance
of the car in a typical hill climb race environment. The results are shown in Figure 6(a), Figure
6(b), Figure 7(a) and Figure 7(b). In this case, drag variations were not plotted as they were totally
negligible.
(a) Downforce variations resulted from the heave sen-
sitivity study
(b) Centre of pressure position variations resulted from
the heave sensitivity study
Figure 6: Heave sensitivity study results
An important result was that, from Figure 6(a), it could be postulated that the aerodynamic
behaviour of the second configuration (H50R0.5), unlike the first configuration, helped the car to

avoid hitting the ground as, when the car gets closer to the ground, it loses downforce, helping it to
return to its standard position. Nevertheless, total downforce was not sensitive to heave movements,
with only small variations (2%) due to ride height, while the variation of downforce from the static
ride height of the first configuration was 11%.
On the other hand, Figure 6(b) shows the variation in the CP position due to different ride
heights. The variation in CP with the H50R0.5 configuration was larger than the total downforce
variation, being 6% in this case, but it is clear that the variation was almost linear (R2= 0.982),
making it more controllable.
(a) Downforce variations resulted from the pitch sen-
sitivity study
(b) Centre of pressure position variations resulted from
the pitch sensitivity study
Figure 7: Pitch sensitivity study results
The results show a much more significant dependency on pitch angle in terms of total downforce,
with variations of 11% for the H50R0.5 configuration corresponding to tests at every 0.2 degrees,
while the H40R0 configuration had a variation of 14% at that point. Figure 7(a) shows these
variations. In this case, the variations in CLA were caused by the different performance of the
diffuser and the mid-wing at each pitch angle, as the angle of attack of the latter was also modified
when pitching, but the total downforce variations were also driven by variations in the frontal area
of each configuration.
As far as the CP is concerned, Figure 7(b) shows that the car was much more sensitive to
pitch movement than to heave, with variations of 12.5% from the static H50R0.5 configuration CP’s
longitudinal position.
3.2 Final configuration results
As a consequence of more desirable static, heave and pitch sensitivity studies results, the H50R0.5
configuration was selected as the final configuration. Its results are shown in Table 4. Figure 8(a)
and Figure 8(b) show the pressure field on the car surface.

Table 4: Results of the definitive configuration (H50R0.5)
H50R0.5 1.88 1.10 0.58 27.04 1.01
(a) Upper part of the pressure plot on the surface of the car
(b) Lower part of the pressure plot on the surface of the car
Figure 8: Pressure field on the car surface
Figure 9(a) and Figure 9(b) show front and rear cooling ducts, which take air from stagnation
points and direct it to low pressure zones, in both cases reducing drag and assisting with the
front wheel wake cleaning. There is an important feature to highlight at this point. This feature
is that the simulations were undertaken with static wheel rims, imposing tangential velocities at

wheels’ surfaces in order to simulate wheel rotation. It is for this reason that front cooling duct’s
streamlines were less chaotic than expected through the wheel. Real streamlines would present
much more interference with the wheel rim.
(a) Front cooling duct’s streamlines
(b) Rear cooling duct’s streamlines
Figure 9: Cooling ducts’ streamlines
A crucial aspect of the particular design of the car was the interaction between wheels’ wakes
and diffuser. Figure 10(a) and Figure 10(b) clearly show the high interaction of the separated flow
coming from the rotating wheel with the diffuser. One can therefore conclude that more diffuser
strakes would help to reduce this unwanted interaction, even though this would not be permitted
by the regulations.
An area of further work would be roll and yaw sensitivity studies, even though it is thought
that the car is less roll sensitive than an standard design provided its smaller floor area, so that
contribution to the total downforce would be lower.

(a) Wheels wakes’ streamlines under the car
(b) Wheels wakes’ streamlines around the car
Figure 10: Wheels wakes’ streamlines
To validate these studies, the car should be tested in a wind tunnel or actual track tests. The
wind tunnel tests would have to be done with a rolling ground and rotating wheels in order to
reproduce underﬂoor and wheel wakes accurately enough, increasing the cost of the wind tunnel
testing.

4 Powertrain
The powertrain system of this car consisted of a series hybrid system, with the following components,
as seen in Figure 11(a): Internal Combustion Engine (ICE), generator, supercapacitors, 2 in wheel
electric motors for the front, 2 more electric motors for the rear and the rest of electrical control
components. The main power source of this powertrain system was the 300g of fuel that the
regulations allow [5], as there was no mechanical connection between the engine and the wheels.
This deployed power would be partially recovered by a regenerative braking system, storing this
energy in supercapacitors. An image of the powertrain design of the Cranfield Mixte Hybrid car is
shown in Figure 11(b).
(a) Series Hybrid schematics [6] (b) Drawing of the powertrain parts of the car
Figure 11: Powertrain of the Cranfield Mixte Hybrid car
The decision of choosing a series hybrid instead of a parallel hybrid powertrain, or even a series-
parallel hybrid propulsion system stands with the team’s philosophy of making the car as light and
efficient as possible. Connecting the ICE to the wheels, added mechanical components such as a
differential, a gearbox, a clutch, driveshafts and propshafts. Most of them were metallic components
that could add at least 20 or 30 kg each. Taking into account that some of these components are
complementary, they could add more than 100kg to the total mass of the car, which is not desired.
The design of the powertrain components is described below, following the line from the main power
source to the wheels.

4.1 Internal Combustion Engine
The engine was designed in base to the power required in the simulations of the Appendix Appendix:
Simulations. These simulations predicted an engine producing 110 kW at 7000 rpm, speed at which
the generator works most efficiently. The design of the engine performance was done using the AVL
Boost software, resulting in the characteristics seen in Table 5.
Table 5: Main engine geometries
Bore 88.8 mm
Stroke 88.8 mm
Displacement 1100cm3
Compression ratio 12:1
Intake Valve Diameter 33.6 mm
Exhaust Valve Diameter 28.32 mm
The results of the simulations produced by the AVL engine model are shown in Figure 12(a),
where a value of 112 kW and 137 Nm at 7000 rpm was achieved. As it can be seen, both power and
torque do not demonstrate a significant variation around that speed, so the fluctuations in engine
speed would not affect the performance of the system. The engine was designed to be as efficient
as possible, resulting in a BSFC of 232 g/kWh at that speed.
Figure 12(b) represents the one-dimension model designed in AVL Boost. It consists of a 2-
cylinder 4-stroke, naturally aspirated engine, with direct injection and 4 valves per cylinder. The
Vibe 2-zone model was used to represent the combustion [7], and the Patton, Nistchke and Heywood
model helped to model the friction in the engine [8].
As seen in the regulations, the fuel used in the races must be 95 Octane petrol. In Figure 13,
the minimum octane number required by the engine at each speed is shown. This proves that this
engine is under the regulation and would not have detonation at the optimum speed.
The engine was designed to be a boxer engine with dry sump and air cooled. This allows the
engine to be placed as low as possible in the car. The first stage of the project gives a BMW
HP2 engine (Appendix C) tuned and modified to achieve the power, torque and fuel consumption
required. The final phase of the project would mount a different version of the same engine with
some modifications in the crank and camshafts, removing the wet-sump and balance shaft. This
would allow the engine to be lowered in the car and to remove more than 20 kg of mass.
The process followed to create our model was validated in a test bed. It was done by obtaining
power and BSFC data from a 50cm3 engine (Figure 14(a)) and modelling it in AVL. A full expla-
nation of the testbed can be found in the Appendix D, but the comparison of power and BSFC can
be seen in Figure 14(b). The small fuel tank of 0.5 litres was placed right behind the driver, above
the engine. The main fuel line was provided with a Fuel View DFM-50 fuel sensor, whose datasheet

(a) Engine model performances (b) Engine model in AVL
Figure 12: AVL Boost engine model and results
Figure 13: Octane number required by the engine
can be found in the Appendix E.

(a) Test bed set-up (b) Test bed power matching in AVL
Figure 14: Test bed
4.2 Generator
As a series hybrid architecture was chosen, a generator was fitted to the IC engine in order to
convert all the mechanical power produced by the IC engine to electrical energy. This architecture
allowed the generator to only work at its maximum efficiency point.
The generator was done with a YASA 400 Motor Generator which has the following character-
istics:
(a) Generator Efficiency Map [9] (b) Generator Torque and power curves [9]
Figure 15: Generator characteristics
As seen in Figure 15(a), the optimum operating point of the alternator was at 7000 rpm. Nev-
ertheless, as Shelsley Walsh is shorter than Goodwood, it was decided that the generator would run

at 7200rpm at Shelsley Walsh. As shown in Figure 14, the power of the engine is higher and, as
fuel consumption is not a concern at this track, this operating point resulted in a reduced laptime.
Furthermore, as fuel consumption was not a concerned at Shelsley Walsh, the engine was running
all the time, including during the braking phases.
The Appendix F gives the full datasheet of the YASA 400 motor.
4.3 Energy Storage
The energy storage utilised super capacitors. This solution compared to the batteries allowed a
higher discharge current compared to a battery. As the two tracks were significantly different,
the choice was made to use an assembly of individual cells .This configuration permitted more
modularity by allowing to run a different number of capacitor for each track without the extra
cost of having to buy two different energy storage module. The chosen cell is a Skeleton SCA1800
capacitor (The datasheet of which can be found in Appendix F). For Goodwood, a total of 93
individual capacitors would be used, composed of three parallel strings of 31 capacitors. For Shelsley
Walsh, a total of 129 individual capacitors would be used, composed of three parallel strings of 43
capacitors. The rationale for the choice of the number of capacitors for each tracks is shown in
Appendix G.
4.4 Electric motors
The electric motors used are four YASA 400, one for each wheel. At the front, the motors were
slightly modified from the stock configuration, as those motors will be in wheel motors. This decision
was made in order to be able to deal with the packaging issues created by the narrow front, dictated
by aerodynamic considerations. An additional benefit of using five identical motors was that a least
amount of spare parts has to be carried with the car, as the team running the car would need only
one spec of motors as spare part for the motors, both front and rear, and the alternator. The torque
and power curves for the YASA 400 are shown in Figure 14.
4.5 Controllers and electrical power transformers
The electric motors were controlled with the controller provided by YASA that was powered us-
ing a regulated 400V DC electrical supply. This controller also included a CAN bus in order to
communicate with the ECU to control the torque of each motor. The electrical circuit included a
DC:DC converter to convert the voltage provided by the capacitor to a regulated 400V DC to the
controllers. In order to make the energy storage safe in less than 20 seconds, a relay was installed
inside the box in order to isolate the energy storage from any other part of the car within the
required 20 seconds. This relay was a normally open relay in order to open in case of power supply
cut. This relay was powered through an external power supply (battery), which also power a light
(green or red), installed on the roll hoop to indicate to the marshals if it is safe to touch the car.

4.6 Simulations and optimisation
The design and sizing of all the elements of the hybrid powertrain were calculated using a a quasi-
static simulation, taking into account the efficiencies of all the powertrain elements and the dynamic
behaviour of the vehicle. The inputs of the model were the radii of the corners, the altitude and
the distance from the start line for each corner. For Goodwood, the model also used the length
during which the IC engine will be running before the lift and coast periods. Those times have
been optimised using the Matlab solver, to reduce the Goodwood laptime, without consuming more
than the allowed 300 grams of fuel. At Goodwood, the four wheels motors were the only braking
elements, as the friction brakes were not used during the course in order to be able to recover as
much energy as possible. However, friction brakes were mounted at the rear to slow the car in
case of emergency. For Shelsley Walsh, the friction brakes at the rear were used to brake in order
to shorten the braking distances. For this course, being able to recover all the energy was not as
important as for Goodwood, due to the short length of Shelsley Walsh, all the recovered energy
could not be deployed. The Shelsley Walsh model also included a braking phase before the finish
line in order to be below the speed of 40m/s to be able to stop within 50m.
The results given by the simulation are:
Figure 16: Simulation Results
Appendix A presents a detailed presentation of the model.
The following figure gives the curves given by the simulation:

(a) Goodwood Simulation results
(b) Shelsley Walsh Simulation results
Figure 17: Simulation Results

5 Cooling
The cooling considerations of this vehicle were an area of innovation, and whilst not fully developed
and validated, the means by which the cooling this vehicle is achieved is one that accounts for the
cooling considerations whilst also considering the car as a whole, which must use the fuel given as
efficiently as possible.
Using the relationships described in Appendix H, the decision was taken to try and eliminate
the closed water cooling system and to instead utilise the lubrication system as a cooling system
also. With an assumption of 5kg of coolant, this saving would lead to a reduced lap time of 0.25
seconds. Also, the oil would be circulated even when the IC engine was not running. This reduced
the redundancy in the system and led to a more weight efficient design, as the oil was be lubricating
the motors while the engine was not running as well as cooling it.
A 1-D analysis was conducted using the above relationships to model a mass of aluminium to
simulate the engine block heated by a power equal to that of the power produced by the engine.
Then, assuming a convective area of 1m2, heat was transferred into the oil, and then transported to
the back of the car, where it then rejected that heat into an aluminium block at the rear of the car.
This block had a convective surface also of 1m2, allowing the transferring of heat from the oil into
the aluminium tank. From there, using the correlations available for the heat that can be dispersed
through a fin, one was able to calculate the heat dissipated into the air. It should be noted also that
owing to the one dimensional nature of this analysis that local temperatures are not accounted for.
From the simulation of the hybrid powertrain components, a timeseries of engine running times
and speed were taken into this simulation. The engine switching running times were used as the
times during which heat was fed into the aluminium block simulating the IC block, and the speeds
used to increment the air speed over the fins for each time step. Thus, the heat they could reject
into the airstream could be calculated using the correlations discussed in the appendices but also
shown here;
h =
Nuk
L
(1)
where;
Nu = 0.664 Re
1
2 . Pr
1
3 (2)
A Simulink model was constructed to simulate this, and was run with a large matrix of fin
dimensions. For a fixed length, a Matlab code then ran that Simulink model again and again
trying different fin widths, heights and numbers of fins to find combinations that would keep the oil
beneath 130oC. Then, once all of those results were stored, sweep through those results to find the
fin dimensions that met the temperature criteria with the smallest volume of fins, thus optimising
the solution for minimum weight, which as aforementioned is a sensitive parameter of the car.
Figure 18(a) shows the temperature output plots of the Simulink model with the optimal fin
dimensions. The three data sets shown are the bulk temperatures of the IC engine block, the oil

(a) The output plot of the Simulink model for Goodwood
(b) The oil tank in situ at the rear of the car
Figure 18: Figures demonstrating the oil cooling system fitted to the car
and the aluminium oil tank at the rear of the car. The horizontal line shows the maximum oil
temperature of 130oC and the oil temperature stays just below this temperature through use of the
optimisation code. The highest temperatures experienced here on the Goodwood duty cycle is the
portion of the run as the car decelerates into the finish line and turnaround section, as the airflow
reduces and the engine is still running around this time.
The tank shown in the CAD drawing is purely illustrative of the component that would be
created should the car be brought to the market. The fins could be manufactured in a more cost
effective way, as the geometry shown in Figure 18(b) could only be made by casting or milling the

component out of a solid block that would be very time consuming and expensive. The innovation
in this area of the car was to use no water or fans to cool the car. Also, an attempt was made
to use fins in the freestream to cool the car, whilst optimising that system for its weight or more
clearly specific heat rejection. This led to a design that aimed to bring performance to the car in
an environment where every kilogram of carried weight is critical to the overall lap time.
5.1 Effect of Engine Switching
When the decision was made to increase engine power and to switch the engine on and off instead
of having the engine run at one speed constantly throughout the run, a cost benefit analysis was
undertaken to understand the effect of this change on the cooling system. Two runs were completed,
with the engine running at a constant power of 40kW and one with the switching regime at 70kW.
The cooling requirements were reduced in the switching regime, and Table 6 shows the benefit this
had on the car in terms of fin dimensions and hence weight. The engine switching regime saved
circa 10kg at the rear of the car owning to the reduction in fin size requirements.
Table 6: Effects of the engine switching regime on fin dimensions and weight
Engine Power Optimal height (m) Optimal width (m) Optimal number Weight (kg)
40kW constant 0.03 0.01 72 29.2
70kW switched 0.02 0.01 72 19.5
The cooling requirements are greater for Goodwood, so the configuration that is optimised for
this circuit will be used which is shown in Table 7. Towards the end of the project, the engine
power was increased, thus the cooling requirements for the final configuration even with a switching
regime cannot match the configurations shown above.
Table 7: Final Fin Configuration
Engine Power Optimal height (m) Optimal width (m) Optimal number
110kW 0.1 0.03 150

6 Vehicle Dynamics
The design of the suspension and handling of the race car will be discussed in the following section.
6.1 Sensitivity Study
Due to the unconventional design of a narrow front track, a sensitivity study of the weight distribu-
tion between tyres versus front track, rear track, and front/rear weight distribution was undertaken.
Some assumptions and initial conditions were required to complete the sensitivity study shown
in the following table 8.
Table 8: Initial Conditions
Lateral g, Ay 1.5
Data Front Rear
Wheel radius, r (mm) 277 287
Height roll centre, hrc (mm) 20 30
Unsprung Mass, Mu (kg) 30 15
Wheelbase, l (mm) 3000
Sprung mass, MS (kg) 400
Height sprung mass, hms (mm) 150
For the sensitivity study a tyre stiffness of 250 N/mm was chosen using examples from [10]. In
the final suspension calculations, a tyre stiffness of 166.77 N/mm was used as shown in Appendix I.
Figures 19(a), 19(b) and 19(c) show the results from the sensitivity sweeps of the front track,
rear track, and the front/rear weight distribution. Owing to the non-linear increase in front tyre
load with respect to front track and the aerodynamic advantages of a narrow front track, a final
front track width of 700 mm, rear track width of 1800 mm, and a front/rear weight distribution of
30% was chosen.
Another benefit of having a rearward weight distribution allowed for the front roll stiffness to
be influenced by the chassis torsional stiffness and rear roll stiffness which allows for softer front
spring rates, thus increasing front grip [10]. This helped to reduce the understeer tendency due to
the narrow front track as well as increasing the moment force that the front tyres generated around
the centre of gravity, further reducing the understeer of the car.
6.2 Geometry
The main target for the suspension geometry was stability, ensuring that during all dynamic forces
and accelerations the car reacts and performs in a predictable way. Controlling the camber, toe,

(a) Vertical tyre load vs. front track width (b) Vertical tyre load vs. rear track width
(c) Vertical tyre load vs. front/rear weight distribution
Figure 19: Sensitivities of Vertical Tyre Load
and scrub were secondary objectives however could not be neglected.
With stability under all dynamic forces being a key objective, good roll centre control at the
front and rear became the most important variable. To reduce the roll centre movement and control
the camber of the tyres, converging wishbones of unequal length were used. Keeping the roll centres
close to the road surface as well as reducing their movement as much as possible determined the
suspension connection points. Maximising the length of the suspension arms helped reduce the
camber change through suspension working space. This ensured the packaging of the in-wheel
motors at the front, and ensured that the scrub radius was kept to a minimum to prevent any
sudden steering force variation. Keeping the roll centres close to the road surface also reduced the
amount of jacking that the car would have experience when subjected to lateral accelerations.
Similar roll centre movement at the front and rear of the car was also desired so that the
suspension stiﬀness and thus the balance of the car is maintained throughout the dynamic movement

of the chassis. As seen in Figure 20 the roll centre movement of the front and rear mirror each other
closely. The static height of the front and rear roll centres is 28.4 mm and 32.6 mm respectively.
For the camber curves starting at a static camber of -1.5 degrees for the front and rear, the tyres
stay below 0 degrees of camber up to 2 degrees of chassis roll for the front and 2.5 degrees of chassis
roll for the rear. As a consequence of this, the suspension roll rates were required to be stiff enough
to ensure that the rear tyre camber did not go over 0 degrees of camber.
Figure 20: Camber and roll center curves vs. chassis roll
Front camber was recovered from the caster angle shown in Figure 21. Another benefit to the
caster angle is the caster trail which improves straight line stability. This was kept under 30 mm
so that the self-aligning torque of the caster trail did not overpower the pneumatic trail of the tyre
which allowed the driver to feel for the maximum peak grip.
6.3 Non-Linear Spring Rates
From [11] stating the most effective ride height for ground effect aerodynamics is 0.7h/θ (where θ
is the diffuser angle) resulting in an ideal ride height lower than the 40 mm minimum imposed, non
linear springs were used to allow the race car to quickly and easily reach the optimal aerodynamic
ride height.
A ride rate sweep was done for 0 - 2000 N downforce and 10 - 40 mm dynamic ride height as
shown in Figure 22.
Figure 22 shows that the static front and rear ride rate was 7.67 N/mm and 6.72 N/mm with
the maximum ride rate of 22.84 N/mm for the front and 30.94 N/mm for the rear.
Using these ride rates, a more conventional suspension set-up is shown in table 9.

Figure 21: Front tyre camber curve from caster and kingpin angles
Figure 22: Required ride rate for downforce and dynamic ride heights

Table 9: Suspension Set-up 1
Suspension Set-up 1
Front Spring Rate N/mm 24.62
Rear Spring Rate N/mm 30
Rear Anti-roll Bar Spring Rate N/mm 15
Front Heave Spring Rate N/mm 12.31
Rear Heave Spring Rate N/mm 16.93
6.4 Anti-Roll System
To reduce the drawback of suspension stiffness in single wheel displacement which can transfer loads
to the opposite wheel, a second vehicle set-up was proposed. A greater percentage of the suspension
stiffness was realised using the anti-roll and heave springs using a frequency limited hydraulic or
pneumatic system. Once the suspension system experiences a frequency input above a designed
limit, the force generated would bypass the anti roll or heave system thus preventing any load
transfer from one tyre to the other under single wheel bump or rebound. This allowed for a stable
aerodynamic platform while permitting softer corner spring rates which will react to road bumps
and curb strikes more efficiently [12], [13] [10] . Therefore, the second vehicle set-up is shown in
table 10.
Table 10: Suspension Set-up 2
Suspension Set-up 1
Front Spring Rate N/mm 24.62
Rear Spring Rate N/mm 30
Rear Anti-roll Bar Spring Rate N/mm 15
Front Heave Spring Rate N/mm 12.31
Rear Heave Spring Rate N/mm 16.93
6.5 Damper Settings and Sweeps
The general method to determine the optimal damping coefficients is done through physical testing
following several steps such as KONI’s published instructions [12].
Due to the lack of physical testing, a frequency and damping coefficient sweep was done using
a quarter car Simulink model shown in Figure 23. The damping coefficients for low and high speed
displacement were chosen based on the minimum root mean square vertical tyre force output, results

Figure 23: A quarter car Simulink model
shown in Figure 24 and 25.
(a) 3D plot (b) 2D plot
Figure 24: Front damper coeﬃcient and frequency sweep
From these sweeps, the following conﬁguration was chosen shown in Table 11.
6.6 Tyres
Avon moto tyre used based on rearward weight distribution of the car. Owing to the rearward weight
distribution of the car, the vertical load that the tyres would experience due to weight transfer and
aerodynamic loads, as well as the availability of tyre sizes, 180/550R13 Avon Motorsport F3 front
tyres and 250/570R13 Avon Motorsport F3 rear tyres were chosen. These tyre sizes were also able
to accommodate the in wheel electric motors being used.

(a) 3D plot (b) 2D plot
Figure 25: Rear damper coefficient and frequency sweep
Table 11: Damper Coefficients
Front Damping Coefficients Ns/m
Low Frequency High Frequency
Bump 1095 545
Rebound 2460 1230
Rear Damping Coefficients Ns/m
Low Frequency High Frequency
Bump 2700 700
Rebound 5270 2635
The Pacejka model [14] was used to help determine the balance of the car when in steady state
cornering. Using the provided Pacejka coefficients shown in Figure 26, the vertical wheel loads, and
the calculated lateral grips required, the required slip angles of the tyres were determined.
From this example, the front slip angle was 1.2 degrees larger than the rear slip angle which
correlates to a mild understeer balance of the race car.

Figure 26: Avon F3 pacejka model coeﬃcients

7 Structures and materials
The design, optimisation and validation of the structural components of the vehicle will be discussed
in the section which follows. Whilst this section provides a broad overview of the materials and
structures used in the design of this vehicle, greater detail on this section can be found in Appendix
J.
7.1 Materials
A categorisation of two groups of structures types was implemented as a result of requirements and
cost:
1. Components of critical stiffness and quasi-static load bearing requirements which compromise
cost and cover a significant portion of the volume of the vehicle.
2. Components of which were critical in dynamic, energy absorption, requirements, were cost
effective and cover a small portion of the volume of the vehicle.
The first group of structures demanded materials of high specific stiffness and strength with a
compromise to cost. The combination of materials chosen is presented in Figure 27 below.
Figure 27: Sandwich construction illustration
The carbon fibre pre-impregnated (CF prepreg) laminate consisted of Toray T700-SC-12K plain
weave and SHD MT510 resin system while the polyurethane (PU) foam core was Airex C71.75.
The second group of structures demanded materials which possessed dynamic properties which were
well understood and validated. This allowed for accurate modelling and a safety factor to be applied
to meet the requirements, thus, Aluminium 5182 was chosen.
The joining of various structures were accomplished through the use of Araldite 420 adhesive and
bolted configurations depending on the load application they were subjected to.
The data sheets and reference validated properties of the above materials can be found in Appendix
J along with material choice optimisation and explanation.

7.2 Testing and simulation
Samples were manufactured to allow for finite element (FE) validation and thus component opti-
misation.
7.2.1 Three point bend test
The sandwich construction composite was tested in a three point bend test rig according to [15]
and [16] to obtain the stiffness and strength properties of the composite. Figure 28(a) and 28(b)
display the test sample and the test rig set up. The layup configuration can be found Appendix J.
Figure 29 depicts the experimental data against the simulated data of composite and manipulated
isotropic models.
(a) Test sample (b) Test rig setup
Figure 28: Three point bend testing
It was shown that the stiffness was accurately modelled using composite inputs, however, the
yield strength and onset of material softening was under predicted. Manipulation of an isotropic
material allowed for accurate properties (stiffness, yield strength and failure) of the structure to be
modelled.
7.2.2 Joining
Adhesive in peel and bolt in lap shear joint failure testing and simulation results can be found in
Structures and materials appendix.
7.3 Components
7.3.1 Roll structure
The main and front roll structures were designed to meet the regulations as defined in [17], to
protect the driver in the event of a roll-over incident and constructed as per Figure 27. The roll

Figure 29: Three point bend test and simulation results
structure load application zones can be seen in Figure 59.
Figure 30: Roll structure load application zones and safety area
The main roll structure can be seen in Figure 60 and its simulation results in Figure 62 - the
construction of which is deﬁned in Appendix J. The aerofoil which is placed ontop of the main
roll structure is joined in such a way that, in a roll-over scenario, the aerofoil will break oﬀ from
the main roll structure and not impede its performance. The front roll structure simulation results
shown Figure 64.

Figure 31: Illustration of the main roll structure

Figure 32: Main roll structure simulation and results
Figure 33: Front roll structure simulation and results

7.3.2 Energy storage structure
To safely house the capacitors required, an energy storage structure was designed which could
withstand the regulated impact and insulate the electrical energy through the use of an ethylene
propylene diene monomer rubber inner liner. This regulation avoids scenarios like those in [18] and
[19] experienced by Tesla which produced catastrophic results due to unconsidered impact scenarios
on the energy storage systems. Figure 34(a) depicts the structure while Figure 34(b) depicts the
structure at final impact. Figure 35 shows the simulation results and that all the energy was
absorbed. The construction and design of which can be found in Appendix J.
(a) A render of the energy storage box (b) Simulation render results
Figure 34: Energy storage structure
7.3.3 Chassis
The composite chassis (constructed as per Figure 27 and illustrated in Figure 36) provides a torsional
stiffness of 1.6kNm/deg and a lateral stiffness of 62N/m which allow the handling of the vehicle to
be adjusted. The torsional stiffness was at least 1.3 of the maximum roll stiffness of the suspension
(1272Nm/deg) as proposed by JC Dixon [20]. The torsional stiffness simulation results can be seen
in Figure 75 which depict an optimisation of number of plies per side of the core and thus weight.
Figure 38(a) and Figure 38(b) show the maximum strain in the chassis under the two loading
conditions.
A weight saving of 16% was made as per torsional stiffness configuration to the target stiffness
with a resultant weight of 43.1 kg as well as keeping the maximum strain below 0.01 which was found
to be the onset of yielding in the three point bend test. The stress concentrations can be reduced
by increasing ply overlap and producing geometry with less change in direction and sharp edges.
The final specific modulus and strength of the composite resulted in 23.1 MPa/kg/m3 and 0.081

Figure 35: Energy storage crash box simulated acceleration results
Figure 36: Composite chassis render
MPa/kg/m3 respectively. Sample calculations and details of the layup can be found in Appendix
J.

Figure 37: Torsional stiﬀness target, simulation results and maximum load of the chassis
(a) Torsional load of 2kNm (b) Lateral load of 3kN
Figure 38: Maximum strain in the chassis under prescribed loads

7.3.4 Rear wishbone
The analysis of the rear wishbone was done so in order to ensure that the component could withstand
the highest lateral acceleration of the vehicle. The analysis utilised a maximum lateral accelera-
tion of 2G, assumed a 70% rearward weight distribution and that all the load on the rear was
transferred to the outer tyre. This would impart a moment as shown in Figure 77 and 79 of 4635
N where the wishbone analysed is highlighted in red. The construction of the rear wishbone was
[+45◦, 0/90, +45◦, 0/90, core]S
Figure 39: Load applied to the rear wishbone under worst case conditions as shown in the full
vehicle
The final effective stress plot is that shown in Figure 79 and clearly shows the stress is below
the yield stress (21MPa) of the material. It should be made aware that the hydrostatic stress is not
accounted for.
7.4 Chassis Manufacturing
The manufacturing design process of the chassis is detailed below due to the combined significance
of design and manufacturing in producing a structurally sound the chassis. In practice, the follow-
ing two sections were not separated and designed in conjunction with one another. Figure 41(b)
illustrates the design.
7.4.1 Ply layup design
The laminate design consisted of a number of factors namely, but not limited to:
• Manufacturing process-ability for the laminator to place the ply as required. This linked to
the drape-ability of the fabric and thus the choice for plain weave material.

Figure 40: Effective stress in the wishbone under maximum cornering
• Optimised structural properties in the desired directions. Torsional and lateral stiffness of the
chassis demanded a laminate which possessed fibres in the +45◦ and along the longitudinal
axis directions.
• Consideration of overlaps and joins reduced stress concentrations in tight corners such as the
front and rear portions of the chassis.
• Targeted weights and thicknesses drove the choice for four plies per side of the core.
A detailed plybook was constructed to allow the laminator to understand and perform the layup
easily.
7.4.2 Tooling and manufacturing design
Tooling design was based on, but not limited to:
• The shape of the component dictated where mold joins were due to tool machining and
component demolding draft angle limitations.
• Understood the number of chassis to be produced and thus chose a recyclable and easily
machinable tooling material of high density PU model board.
The other choices made regarding manufacturing were;
• The use of an autoclave process for improved ply consolidation, void content and resin volume
fraction [21].
• Hand layup due to the bespoke nature of the chassis.

• A cure cycle ramp and dwell, time and temperature as per [22].
• Joining of the top, bottom and seat sections of the chassis (Illustrated in Figure 41(b)) to one
another through the use of a tongue overlap as shown in Figure 41(a).
• Joining of the other vehicle components using either adhesive or bolted conﬁgurations as per
the preferred loading on that joint - such as bolted wishbone mounts.
(a) Tongue join technique used for the
three sections of the chassis
(b) Separately manufactured components of the chassis
Figure 41: Chassis manufacturing techniques

8 Summary
The car proposed by this project is one that has capitalised on numerous innovations, from the
unique cooling concept to the overall aerodynamic direction of the vehicle. The vehicle has been
simulated to complete the task set whilst adhering to all of the rules imposed, by maximising
performance in the areas of the car that have the greatest impact on the overall laptime. All of
these innovations deviate from what has been done many times before, and it is believed that these
innovations are what will set this vehicle apart from its competition.
8.1 Further Work
While this report shows a project that accomplished the body of what it set to achieve, there are
still areas of the project requiring further work. With greater resources, one would have explored
the aerodynamics in greater detail, in particular the use of Reynolds Stress Models to further
improve the CFD accuracy. In the realm of vehicle dynamics, the performance of the car would
be improved by the utilisation of the frequency-dependent anti-roll system proposed. Cooling is an
area of further work also, to optimise the system for maximum heat rejection for minimum weight.
In powertrain, the innovative engine design proposed would be an area of further innovation and
greater performance. Finally, the simulations developed especially for this project may be improved
to take into account more variables to better estimate the ﬁnal overall laptime of the vehicle.

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A Presentation of the Simulink Model
Engine
The engine is modelled using the engine model of Simulink. In parallel of the engine, a friction
module has been placed to simulate the engine internal frictions. For the different parameters, the
model use the values given by the AVL model of the engine. The engine is controlled by putting
full throttle during the desired periods determined to maximise performance. The torque of the
alternator is then controlled using a closed loop to keep the engine at the desired speed.
Energy Storage
The energy Storage is modelled using a Capacitor that has the equivalent values of the arrangement
of ultra-capacitors. The value of the capacitor voltage sensor is send to the ECU in order to control
the power delivered to the four motors to be sure to never go under the voltage that is in the
capacitor at the start of the run.
Motors
The electric motors are modelled as a system that convert mechanical power into electrical power
(generator or regenerative braking) or electrical power into mechanical power (motor). The model
include as well losses that ore proportional to the squared torque. The k coefficient has been
determined in order to match the performances of the model to the datasheet of the motors. The
following figure shows how the motors were modelled:
Figure 42: Motor model
Vehicle Dynamics
For the modelling of the car, the default vehicle dynamic module of Simulink was used. The
limitation of this module is that it does not take into account the lateral load transfer, which can
limit the precision of the modelling for the corners. The model does not take into account the

downforce produced by the car so a vertical force, was added to all the wheels, representing the
aero load produced by the car regarding its speed.
Controls
The controls aim is to give to each motors the target Torque, the throttle to the IC engine and the
torque that the generator has to produce to maintain the engine at the optimum speed while being
at full throttle. The model includes a basic modelling of an ABS and traction control, which is
controlled independently for each wheels. The inputs of the model are the four wheels speeds, the
four wheels longitudinal slip ratio, the capacitor Voltage, the throttle and brake command from the
driver, and the rotational speed of the IC engine. The outputs are the torque each motor has to
produce, the torque that the alternator has to produce and the throttle command for the engine.
The torque for each motor is the product of the maximum torque that the engine can produce
(regarding its rotational speed), the traction control (or ABS) command, and a factor according to
the battery charge. The traction control and ABS works according to the same principle: if the
slip ratio of the tire is over the given constant value (Which correspond to the slip ratio that gives
the optimum traction), the system will produce an output of 0, which will cut the torque from this
motor until the slip ratio comes back under the target value. All those calculations are done using
the absolute value, which allows the same system to work as a traction control and as an abs as
far as the regenerative braking is concerned. For the capacitor voltage control, when the capacitor
voltage is under a threshold, it multiplies the output torque by a factor between 0 and 1 in order
to not consume more energy than the energy produced by the generator.
To control the torque applied to the engine, the torque applied by the generator will be controlled
using a feedback loop in order to maintain the engine speed equal to the fixed target speed.
Driver
The virtual driver that controls the motors is given a curve of speed vs distance while braking which
is created using the same model with the driver only braking from a predefined speed, until the
vehicle stop. The driver is also given the position of the corners (in term of distance) and their
radius. The model calculate the speed in this corner using a fixed maximum lateral acceleration
given by the vehicle dynamic simulations. The model calculate if it can use the throttle of if it must
brake. For Goodwood, the model also takes into account the input values telling him for how long
the engine must run between the corners before using lift and coast periods to save fuel.
Turnaround
The turnaround procedure is different for both track because of the difference of distance to brake
for the two tracks and the higher importance of recovering as much energy as possible at Goodwood.
At Goodwood, the regenerative braking is enough to stop the car in less than 100m so the driver

just use the regenerative braking to transform the kinetic energy of the car into energy that will
be stored into the capacitors. At Shelsley Walsh, in order to be able to stop the car in less than
50m, the driver will have to brake using all the regenerative power available and the friction brakes.
However, in order to be able to stop the car in less than 50m, considering that the car does not
have friction brakes at the front, the maximum speed crossing the finish line must not be higher
than 40m/s. Therefore, the simulation include a “virtual corner” at the finish line that force the
driver to slow down to 40m/s at the finish line.

Detailed Simulation results
Simulations results:
Goodwood
Downhill
Laptime: 44.076s
Total IC engine running time: 20.9992s
Total fuel consumption: 126.5325g
Figure 1 Goodwood downhill simulation results

Turnaround
Final Capacitor Voltage: 75.5591V
Braking distance: 98.0335m
Figure 2 Goodwood turnaround simulation results

Uphill
Laptime: 43.448s
Figure 3 Goodwood uphill simulation results

Shelsley Walsh
Downhill
Laptime: 24.235s
Figure 4 Shelsley Walsh downhill simulation results

Turnaround
Final Capacitor Voltage: 115.6775V
Braking distance: 49.8002m
Figure 5 Shelsley Walsh turnaround simulation results

Uphill
Laptime: 23.932s
Figure 6 Shelsley Walsh uphill simulation results

Summary of the results:
Goodwood Downhill laptime: 44.076s
Goodwood Uphill laptime: 43.448s
Goodwood total laptime: 87.524s
Shelsley Walsh Downhill laptime: 24.235s
Shelsley Walsh Uphill laptime: 23.932s
Shelsley Walsh total laptime: 48.167s
Maximum Speed: 206.2294 km/h
Maximum Front Wheel Speed: 2556.38 RPM
Maximum Rear Wheel Speed: 1979.346 RPM
Maximum Goodwood Capacitor Voltage: 81.5442 V
Maximum Shelsley Walsh Capacitor Voltage: 121.7819 V
Goodwood total engine running time: 49.2834 s
Goodwood total fuel consumption: 297.0507 g
Shelsley Walsh total engine running time: 48.1677 s
Shelsley Walsh total fuel consumption: 298.8128 g

B Aerodynamics and Computational Fluid Dynamics (CFD) Ap-
pendix
1. Race car aerodynamic design
The main target of the design was minimum drag due to the CD sensitivity study results: a
reduction of 2.53 seconds in the two tracks added time would be possible for each 0.01 CDA reduction.
The design had to be coherent between all its areas, finding the right balance and using solutions that
improve more than one aspect of it at the same time.
After an extensive review of current designs, there was a clear referent in terms of efficiency:
the Nissan Zeod RC and its predecessor, the DeltaWing. The interesting feature of the Zeod RC was
not only its innovative aerodynamics, but also that it was a hybrid car similar to the concept of focus in
this project. Ben Bowlby, the original DeltaWing designer, mentioned “We wanted to make a car that
is twice as efficient in every way, it should use half the fuel, cost half as much, use half the engine power
and weigh half as much yet still go as fast or even faster than a current Indycar” during the car’s launch.
The key points sought in the Cranfield Mixte Hybrid design were adapted from the one of an Indycar
to the requirements that a hill climb car has. The aerodynamics of the DeltaWing and its effects on the
dynamics and packaging of the car were a crucial aspect of the design, and they are thoroughly
explained in the aerodynamic appendix.
DeltaWing genesis and characteristics
The DeltaWing concept was a Ben Bowlby’s idea, as it is explained in [1]: “The triangular
profile —known as a delta wing planform— is common among Top Fuel dragsters and land-speed
record cars. But those machines race only in a straight line. If they had to turn at high speed, wouldn't
they just topple over like a little kid on a tricycle? As Bowlby thought more deeply about the issue, he
realized that the problem with most three-wheelers was not the number and arrangement of the wheels.
It was the disastrously high centre of gravity. So he conducted an experiment. He bought a pair of radio-
control cars, modified one to run with a single, centred front wheel, and tested them both on a frigid
winter night on the suburban streets around his home in Zionsville, Indiana. The battery-powered three-
wheeler, with its low centre of gravity, turned just fine. In fact, it cornered at higher speeds than the
four-wheel version. Later, back at the Ganassi shop, computer simulations showed that a full-size car
built on the same template should turn just as well.
[…] Under Indy rules, a vehicle with three wheels isn't even considered a car. But Bowlby
discovered that two small front wheels placed side by side would corner nearly as well as a bigger one.
This still allowed him to minimise drag by shaping the car around a narrow nose. Tiny tires also initiated
a cascade of design changes that progressively reduced the weight of the car. Smaller wheels meant
smaller brakes and suspension components, which meant a smaller engine, which meant a smaller
gearbox, which meant a smaller chassis, and so on. When Bowlby ran the numbers, he figured that his
car could lap at competitive speeds with a puny four-cylinder engine. He'd set out to design a car that
would show off the skill of its driver. He ended up engineering the most efficient race car ever.”

Figure 1. Nissan DeltaWing front axle [2]
As a result of the narrow front track, which can be seen in Figure 1, and a streamlined
bodywork, the reduction in drag was so great that the car could reach 200 mph with only 300 hp.
The small front tyres were one of the crucial aspects in the design of the DeltaWing. During the
car’s development, no company was making a racing tyre with an only 10 centimetres wide tread. The
rear tyres were no that problematic since their size was not much smaller than a typical LMP2 rear tyre.
The solution to these problems came from Michelin, after the creation of a special narrow front tyre for
this car, as shown in Figure 2.
Figure 2. Nissan DeltaWing front tyres [2]

The fulfilment of all the previous targets had to be checked with the actual performance of the
car during a race. Nissan representatives compared the fuel consumption of the DeltaWing car to an
LMP2 car. As [3] explains, the DeltaWing used only 55 percent of the amount of fuel used by the
Conquest Racing P2 prototype, which was also equipped with a Nissan V8 engine, with the two cars
doing very similar lap times.
As far as weight is concerned, the minimum of an LMP2 car were 900 kg, while the DeltaWing
weighted 465 kg. Furthermore, regarding the tyres, the DeltaWing managed to do four consecutive
stints without changing them, being their inspection from Michelin engineers the only reason to finally
change them. The low stress induced on the tyres enabled a much longer duration than those of the
rivals. An important consequence of the low tyre wear was that less tyre sets were used per race, thus
reducing the running costs of the car.
Figure 3 shows a clear force diagram of the car:
Figure 3. Nissan DeltaWing force diagram [4]
By looking at Figure 3, the weight distribution of the car could be obtained: 28.5% at the front
and 72.5% of the weight at the rear. In addition, the downforce distribution was 75% at the rear, as most
of the downforce was created at the region between the driver and the rear axle. However, the ideal
aerodynamic distribution would be the same than the weight distribution, something difficult to achieve
without wings.
These numbers also helped to understand why the front tires could be so small and, at the same
time, allowed to turn: the distance between the front tyres and the Centre of Gravity (CoG) was
considerably large. In that way, in order to have an understeer gradient (K) close to zero (so the car
steering behaviour is as neutral as possible), the lateral stiffness of the front tyre should be lower than
usual, as it is shown in equation 1:
𝐾 =
𝑚
𝐿
(
𝑏𝐶 𝑅 − 𝑎𝐶 𝐹
𝐶 𝑅 𝐶 𝐹
) (1)
In order to obtain K close to zero, 𝑏𝐶 𝑅 must be in magnitude similar to 𝑎𝐶 𝐹, being b the distance
between the CoG and the rear axle and a the distance between the CoG and front axle.

DeltaWing aerodynamics
The DeltaWing aerodynamics were mainly based on the exploitation of ground effect and the
underbody optimisation. The original DeltaWing did not use any type of wing, resulting in a decrease
of car’s drag. The only external aerodynamic device implemented was a vertical fin, likely devoted to
increase stability at high speeds.
The main source of downforce was what [5] names as the Twin vortex underbody downforce
system – BLAT (Boundary Layer Adhesion Technology). This is an old technology invented in 1980
by Dan Gurney and All American Racers engineers John Ward and Trevor Harris, as Gordon Kirby
explains in [6]. They were the designers of the unique Eagle Indy cars that used this system.
Gordon Kirby explains the BLAT as:
“Instead of ducting air underneath the car into an enclosed venturi or underwing in the
traditional ground-effect method, the BLAT concept generated twin vortexes through its underbody and
the trailing edge of the rear bodywork, aided by a very efficient exhaust cooling system. The routing of
the engine exhaust system added further energy and downforce to the airflow, and variants of this
concept have been used off and on in Formula 1.”
Ben Bowlby commented that “We’ve adapted the area, ratio and vortex-spilling side profile of
the ’81 Pepsi Challenger to the DeltaWing.” One of the keys of this system is that it was less sensitive
to ride height variations, one of the problems of usual wing cars. “The car pulled the best numbers we’d
seen in the first run with the BLAT stuff with the car running at high ride height. At low ride height it
wasn’t as efficient but it had a lovely characteristic inasmuch as the downforce increased as the car
rose higher off the road and when the car was car going through the corners at lower speed it was
making the maximum downforce.”
Regarding the drag force, the no utilisation of wings helped on reducing it. Nonetheless, the
most relevant aspect in its reduction was the utilization of such narrow front tires, being them typically
the main source of drag of race cars. Apart from reducing aerodynamic drag, small wheels also result
in a lower rolling resistance. The front narrow track and the streamlined body contributed to achieve a
low CD, and a smaller frontal area than one from similar vehicle, also contributing to drag reduction
(due to the way it is calculated).
Although the CD is unknown, an approximation of the product of CL*A could be calculated
from Figure 3, resulting in a value of 2.4.

Aerodynamic development
The original DeltaWing design had various evolutions during its life as a race car. The first
evolution was the Nissan Zeod RC (Figure 4). It presents a closed cockpit and a more streamlined
bodywork, with clear modifications at the front of the car:
Figure 4 . Nissan Zeod RC [7]
There were also cooling intakes over the rear wheels, placing the radiators in a very unusual
location. The other clear modification was a bigger shark fin, in this case used in Le Mans, where high
speed stability was essential. However, the external aerodynamics were not further modified, with
wings still not being used.
At the same time, Panoz developed the Step 2 of the original DeltaWing, the coupe DeltaWing.
The high downforce kit developed for tight tracks included a front splitter with twin endplates, rear
wing mounted between the rear wheels and the engine cover, and a gurney flap at the diffuser (Figure
5 and Figure 6).

Figure 5. Coupe DeltaWing front splitter [8]
Figure 6. Coupe DeltaWing high downforce kit [8]
A fin was also used in some races, called “dolphin fin”, acting as a stabilizer for the rear. It is
shown in Figure 7.

Figure 7. Coupe DeltaWing’s dolphin fin [9]
Hill climb design
The hill climb car’s design was based on the DeltaWing and had to be modified both for the
fulfilment of the different rules of this class and also to account for the different requirements a hill
climb car had compared to a high-speed track car. The overall design and features of the car are exposed
below.
A first important characteristic of the design was the use of covered front and rear wheels. The
wheels of an open-wheel car create approximately 40% of its drag [10]. Thus, covering them would
contribute to a high reduction of the overall car’s drag.
The car used a series hybrid system. The energy obtained from the internal combustion engine
was converted into electric energy and stored until demanded. In addition, the braking was done by the
electric motors, recovering kinetic energy and contributing to improve the car’s performance. There
was one electric motor per wheel, enabling the car to recover energy from all the wheels, but also
allowing the implementation of torque vectoring and four-wheel drive.
The front motors were placed in-wheel in order to enhance packaging, even though the
unsprung mass of these wheels was increased. The no utilization of front friction brakes left room to
these motors. A duct for each wheel was placed from an opening at the front of the nose of the car and
pointing to each motor (Figure 8 and Figure 9), in a similar way as the brake cooling ducts are used in
usual cars. The opening was placed in the zone of maximum pressure, helping to reduce the total drag
and increasing the cooling effect of this particular flow. Furthermore, the airflow exiting this duct was
used to “clean” some of the wake of the front wheels, an issue very important in this specific case as it
will be explained in detail later.

Figure 8. Interior of the front cooling duct
Figure 9. Outlet of the front cooling duct
The zone in between the coverings of both front wheels was reduced in order to decrease the
CD, keeping a more streamlined shape. The external design of that area was done together with the
packaging development of the front axle in order to find the best compromise.
The cockpit design was oversized, since it was an area where freedom was needed to allow the
placement of all the safety systems and the driver without important constraints. The same case
happened at the engine cover (Figure 10): a very streamlined shape was the main target, similar to
current F1 cars, in order to reduce the CD and minimise the area of low back pressure. Nevertheless, at
the same time, some freedom was required to pack all the hybrid system inside. This region of the car
was the result of an iterating process between external design and packaging. However, more iterations
would yield better performance.

Figure 10. Engine cover
The area of the engine cover at the rear of the cockpit featured two inlets at each side: one was
the engine intake, and the other was the rear motors cooling duct (in each side). The intake had the same
area than the intake ports of the engine, which was simulated with AVL Boost. The high pressure of
the zone where the intakes were, worked in increasing the density of the air, improving the volumetric
efficiency and the performance of the engine. On the other hand, the rear cooling duct (Figure 11) went
through the engine compartment and came out at the back of the car, dragging air from the area of high
pressure (cockpit wall) and fulfilling the area of low pressure, reducing the wake and the drag of the
car. The duct exited through the oil tank, cooling both the oil and the motors.
Figure 11. Rear cooling duct
The exhausts were placed on the floor of the car, pointing tangentially to the engine cover
surface (Figure 12). In this way, using the Coanda effect, the airflow surrounding this region was
reenergized, helping to keep it attached and reducing the wake, even though in this design the amount
of exhaust gases were very small due to the capacity of the engine.

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