TEAM 18 – BULLZ RACING

ENGINEERING DESIGN
PRESENTATION
Engineering Design Presentation link:
https://youtu.be/02nZJ0ZYF1M
CONTENTS Page

1. General Overview 1
2. Suspension 2
3. Steering 10
4. Braking System 18
5. Powertrain 19
6. Chassis 33
7. Aerodynamics 39
8. Ergonomics 46
9. Electrical and Electronics 47

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GENERAL OVERVIEW
BZR-02 is our 2nd ever car. Our broad design objectives have been to design a lightweight,
cost-effective, agile and fun car. With this car, we have sought to resolve certain issues that
our previous car had. We have incorporated the use of engine modelling in our powertrain
design, reduced the weight of the car overall by around 20kgs, and have laid special emphasis
on backing all our numbers and figures with rigorous analysis.

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SUSPENSION
Introduction
BZR-02 uses an Unequal double A-Arm Double wishbone suspension with direct acting
springs and dampers. Our objective has been to build a simple, light and nimble car.
We first started by defining our basic dimensions. We decided to adopt a wheelbase of
1525mm so as to build a car that would be as responsive in cornering as possible. A short
wheelbase would allow the car to be shorter, and the chassis relatively light and stiff. We
analysed the lateral acceleration and yaw velocity response for a bicycle model of the car (the
simple linear tyre model was replaced with the tyre brush model) to confirm that a shorter
wheelbase would result in a more responsive car while cornering.
The tyres used are 20/7.2-13" Avon 14254S tyres.

Simulink Model used for the bicycle model of the car

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 The step input steer angle that was applied while determining the wheelbase

Lateral acceleration for a car with 1525mm (left) and 1625 (right) wheelbase

Yaw rate for a car with 1525mm (left) and 1625 (right) wheelbase

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The trackwidth was determined keeping in mind the lateral forces that the car would be
subjected to. We also had to ensure that the trackwidth wasn’t too high as that would require
the car to move a larger distance from side to side while going through slaloms. Ultimately, it
was decided to adopt a trackwidth of 1200mm. CG height was taken to be 300mm based on
the CG location of our previous car and the heights commonly reported by other teams.
Wheelbase 1525mm
Trackwidth (front and rear) 1200mm
CG height 300mm
Vehicle mass (inclusive of driver) 300kg

Ride rates and Roll rates

We picked a front ride frequency of 3Hz and a rear ride frequency of 3.2Hz. This will
provide us adequate stiffness in ride, the front is not equal to the rear. This will prevent the
case of pitching of the car in ride. We want the car to bounce over bumps and not pitch as this
is detrimental to the driver and the control and stability of the vehicle.
Front:rear mass distribution = 55:45
Ride Frequency 2.5Hz
Motion Ratio 0.77
Front Ride Rate 14.8 N/mm
Front ride specifications
Ride Frequency 2.8Hz
Motion Ratio 0.77
Rear Ride Rate 18.57 N/mm
Rear ride specifications
Formulae used for ride and roll rates (Source: OptimumG Tech-Tips):

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With a target roll gradient of 1deg/g, we get the following results:
Roll Rate 747520 N-mm/deg
Front Spring Roll Rate 186040 N-mm/deg
Rear Spring Roll Rate 233370 N-mm/deg
Required Anti-Roll Bar Rate 328120 N-mm/deg
Front ARB Rate 143467 N-mm/deg
Rear ARB Rate 184653 N-mm/deg

Understeer Gradient
We decided to build an understeering car, with a target understeer gradient of approximately
+0.2 to +0.5, and a roll gradient of about +1 deg/g.
We used the following expression to calculate the axle stiffness:

Where,
C12 = front axle stiffness

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Cα = tyre cornering stiffness; obtained from tyre data
Cγ = camber stiffness of tyre; γ = camber angle
ɸ = roll angle
δɸ = roll steer
he= CG vertical height from roll axis
L = wheelbase
f = CG to front axle distance; b = CG to rear axle distance
CTOT = roll stiffness

The value of axle stiffness was then used to calculate the understeer gradient of the car. The
roll rates obtained from the previous calculations were input, and the value of camber change
rate and toe change due to roll were varied with an aim of hitting our target understeer
gradient. We minimised the contribution of bump steer to the understeer gradient value due to
other adverse effects that large toe change can bring in the handling of the car.
Several iterations were carried out and the results are as follows:
Front Camber Change Rate -1°/inch
Rear Camber Change Rate -1°/inch
Front Toe Change Rate -0.2°/inch
Rear Toe Change Rate +0.2°/inch

The resulting understeer gradient achieved is +0.26

Suspension Geometry
The Front View Swing Arm Length (FVSA) was calculated for the camber change rate of
-1°/inch and was found to be 1455.14mm using the formula:
1
𝐶𝑎𝑚𝑏𝑒𝑟 𝐶ℎ𝑎𝑛𝑔𝑒 𝑅𝑎𝑡𝑒 = ( )
𝐹𝑉𝑆𝐴

Front View Swing Arm Geometry

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 Parameters Front Suspension Rear Suspension

Camber Gain -1°/inch -1°/inch

Roll Center Height 57.6 mm 57.6 mm

Front View Swing Arm Length 1455.14mm 1455.14 mm

Scrub Radius 25.4 mm 25.4 mm

Steering Axis Inclination 2° --

Motion Ratio (defined as Spring Travel/ 0.77 0.77

Wheel Travel)

For the packaging within the rim, the interior dimensions of the rims were taken as the
baseline within which the rest of the components were packaged.
Finally, the steering links were added and numerous iterations were carried out on Adams
Car so as to minimise bump steer and reach the target values.

Multiple iterations were carried out on Adams Car

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Component Design
All components are designed for the maximum possible loads that a particular wheel can
experience. The front loads have been calculated for the case of the outer wheel in cornering,
while the car brakes. Similarly the rear loads have been calculated for the case of the outer
wheel in cornering, while the car accelerates. The lateral and longitudinal forces acting on the
wheel from the road were calculated after applying weight transfer (lateral as well as
longitudinal) and the tyre’s friction coefficient.
● For the wheel axle, the maximum bending moment is found to occur at the centre of
the outer bearing.

The resulting maximum bending

moment was found to be around
615Nm. The axle has been
designed with an internal diameter
of 25mm and an outer diameter of
35mm.
The front hubs have been designed
with a factor of safety of 2. They
are made of Aluminium 6061-T6
and are to be CNC machined.
Axle load calculations Front Hub
(Source: Race Car Design
by Derek Seward)
● For the uprights, the loads exerted by the inner and outer bearings while cornering and
braking/accelerating were calculated using the equation below
Similarly the forces exerted by the inner bearing

were also calculated. The bearing forces were

applied on the upright models and several iterations
were carried out to minimise the weight and
maximise the stiffness of the uprights. The front and
rear uprights weigh 843.19 grams and 502.25 grams
respectively
Upright load calculations
(Source: Race Car Design by
Derek Seward)

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 Front uprights – stress and deformation plots

Rear uprights – stress and deformation plots

● The A-arms were designed in

accordance with the loads calculated
as shown.
The A-arm tubes were made of AISI
1020 steel.
A-arm Load Calculations
(Source: Race Car Design by Derek Seward)

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 STEERING
Introduction
Ackermann Steering Geometry has been selected for our car because of the following
reasons:
❖ It will perform better than other steering geometry in low speed conditions.
❖ Simple geometry to construct and fabricate.
❖ The geometry induces less wear and tear of tires

SPECIFICATIONS
Front Track Width 1200 mm
Rear Track Width 1200 mm
Kingpin Centre to Centre Distance 1131.46 mm
Wheelbase 1525 mm
Kingpin Inclination 20
Scrub Radius 25.4 mm
Caster Angle 60
Toe Angle 00
Gross weight of car 300 kg
Weight Distribution – Front : Rear 50:50

Steering Geometry

Computer Aided Design (CAD)

The CAD models of Steering System components were made using SolidWorks 2020. The
complete assembly of the Steering System was also done using the same software.

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 Front, Side and Top Views of
Steering System

Specifications
COMPONENTS SPECIFICATIONS Quantity WEIGHT
1 Steering Wheel Diameter – 10” 1 1 kg
2 Quick Release Mechanism Material – Aluminium Alloy 1 0.3 kg
Operation – Pull to Release
3 Steering Column 1 Diameter – 20 mm 1 0.1 kg
Length – 110 mm
Material – Round Aluminium
Alloy
4 Ball Bearing Inner Diameter – 20 mm 2 0.14 kg
Outer Diameter – 47 mm
Width – 12 mm
Max. Static Load – 5000 N
Max. Dynamic Load – 9300 N
5 Bearing Support Material – Sheet Metal 1 0.12 kg
Thickness – 3 mm
6 V Column Support Outer Diameter – 0.5” 1 Welded to
Thickness – 1.5 mm Chassis
Material – AISI 1018 Steel tube
7 Universal Joint Diameter – 20 mm with Splines 2 0.7 kg
Material – Mild Steel
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 Working Angle – Less than 45°
8 Steering Column 2 Diameter – 20 mm 1 0.16 kg
Length – 180 mm
Material – Round Aluminium
Alloy
9 Rack and Pinion Gear Rack Length – 14” 1 1.2 kg
Pinion PCD – 28 mm
Max. Lateral Load – 1800 N
Material – C40
Height of Rack – 2”
Width of Rack – 3”
10 Rack Bracket Material – Sheet Metal 1 0.36 kg
Thickness – 4 mm
11 Rod End Bearing Bore Diameter – 8 mm 4 0.15 kg
Width – 8 mm
Type – Male Steel on Steel
Max. Static Load – 10000 N
Max. Dynamic Load – 5500 N
12 Tie Rod Diameter – 20 mm 2 0.64 kg
Length – 370 mm
Material – Round Aluminium
Alloy
13 Fasteners M8x40 Allen Bolt & Nut 8 0.210 kg
M6x30 Allen Bolt & Nut 1 0.012 kg
M4x30 Allen Bolt & Nut 3 0.014 kg
Table- Specifications of Components

Calculations
ACKERMANN ANGLE
Ackermann Angle, β = tan-1(KPC−C/2 (WB)) where, KPC-C is Kingpin Centre
to Centre Distance
⇨ β = 20.458 0 WB is Wheelbase of the vehicle

STEERING ARM PARAMETERS

We have, Steering Arm Length, SA = 110 mm
Hence, x – component of SA = 38.25 mm
And, y – component of SA = 103.13 mm Distance of Rack from Front Axle

LENGTH OF TIE RODS

Length of Tie Rods, LT = (KPC−C /2)−(Rack Length/2)−( x component of SA)
⇨ LT = 349.677 mm

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ACKERMANN ROLLING CONDITION
Based on the Track Data of FSAE Autocross Germany and data from the rulebook, it was
decided that Ackermann Rolling Condition should satisfy when the vehicle undergoes a 15 m
radius turn.

Track Data of FSAE Autocross Germany

Based on Track Data, R = 15 m = 15000 mm

Weight Distribution (Front : Rear) = 50 : 50

⇨ CG location from rear = 0.5 x 1525

⇨ CG location from rear = 762.5 mm

Now, R1 = √ (RCG2+762.52)
⇨ R1 = 15019.367 mm
⇨ R1’= 14453.637 mm

For Inner and Outer Wheel Angles to produce 15m radius turn under Rolling Condition are
respectively,
Θi = atan (WB/R1’)
⇨ Θi = 6.023 0

From Rolling Condition,

Θ Θ
⇨ Θo = 5.588 0

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LEAST TURNING RADIUS
Inner Lock Angle, Θi = 35 0 (Assumed)
⇨ Θo = 24.74 0
From Figure,
⇨ Ri = 2658.756 mm
⇨ R0 = 3643.956 mm
⇨ R1 = 2743.655 mm
And,
⇨ RCG = 2847.64 mm

This is the least turning radius of the vehicle.

Least Turning Radius

STEERING EFFORT – CORNERING AND BRAKING

MOMENT DUE TO VERTICAL FORCE
Assume that the vehicle is taking a right turn experiencing 1.5g lateral acceleration with steer angle of
35o . It has suddenly undergone a deceleration of 2.5g.
Hence, Vertical Force = Normal Force + Longitudinal Load Transfer ± Lateral Load Transfer
Now, Normal Force = Weight Distribution × Weight of the vehicle/ 2
⇨ Normal Force = 735.75 N
Now, Longitudinal Load Transfer = Mass of the vehicle × Deceleration × CG Height / (2 × Wheel Base)
⇨ Longitudinal Load Transfer = 723.6884
Now, Lateral Load Transfer = Lateral Acceleration × Weight Distribution ×Mass of the vehicle ×CG /Height
Track Width
⇨ Lateral Load Transfer = 551.8125 N
Therefore, FZL = 735.75 + 723.6885 + 551.8125 and FZR = 735.75 + 723.6885 - 551.8125
⇨ FZL = 2011.251 N FZR = 907.626 N

Moment due to Vertical Force, ⇨ MV = 916.149 N mm

MOMENT DUE TO LATERAL FORCE

ML = (FYL + FYR) r tan(ν)
⇨ Force on Front Axle = Weight Distribution * Weight of the vehicle + Longitudinal Load Transfer
⇨ Force on Front Axle = 2918.877 N
Therefore, Effective Mass = Force on Front Axle/9.81 = 297.54 kg

Now, Lateral Force = 𝐸𝑓𝑓𝑒𝑐𝑡𝑖𝑣𝑒 𝑀𝑎𝑠𝑠 𝑥 𝐿𝑎𝑡𝑒𝑟𝑎𝑙 𝐴𝑐𝑐𝑒𝑙𝑒𝑟𝑎𝑡𝑖𝑜𝑛 / 2 ⇨ FYL = FYR = 2187.157
Moment due to Lateral Force was found to be, ⇨ ML = 116885.5936 N mm

MOMENT DUE TO TRACTIVE FORCE

MT = μ (FZL - FZR) d where, μ is coefficient of friction between tyre and road = 1.5
FZL is Vertical Force on left wheel = 2011.251 N
⇨ MT = 42048.113N mm FZR is Vertical Force on right wheel = 907.626 N
d is Scrub Radius = 25.4 mm

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STEERING EFFORT
Total Torque, T = MV + ML + MT ⇨ T = 159849.85 N mm

Force on Steering Arm, FSA = T / (SA ×cos (β)) ⇨ FSA = 1549.917 N

Force on Rack = FSA = 1549.917 N

Torque on Pinion, TP = FSA * Radius of Pinion ⇨ TP = 21698.842 N mm

Torque on Steering Wheel = Tp = 21698.842 mm

Hence, Steering Effort = Tp/Diameter of Steering Wheel ⇨ Steering Effort = 85.428 N

TURNS LOCK TO LOCK

⇨ Turns Lock to Lock = Lock to Lock Steering Wheel Angle360
⇨ Turns Lock to Lock = 135 × 2/360 = 0.75

STEERING RATIO
⇨ Steering Ratio = Steering Wheel Angle / Inner Wheel Angle
⇨ Steering Ratio = 135/35 ⇨ Steering Ratio = 3.86:1

STEERING GEOMETRY SPECIFICATIONS

Geometry Ackermann
Type Rack and Pinion
Mount Distance 79.64 mm
Universal Joints 2
Ackermann Angle 20.35
Steering Arm Length 110 mm
C factor 109.82 mm per turn
Percentage Ackermann 97.6 %
Tie Rod Length 349.67 mm
Rack Length 14”
Pinion Pitch Circle Diameter 28 mm
Inner Wheel Angle 35
Outer Wheel Angle 24.74
Minimum Turning Radius – CG 2.8476 m
Turns Lock to Lock 0.75
Steer Ratio 3.86 : 1
Steer Wheel Diameter 10”
Maximum Steering Effort 85.428 N

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Computer Aided Engineering (CAE)
MULTI BODY DYNAMICS (MBD)
The geometric model of the Steering System was modelled and simulated using MSC Adams
2020 Student Edition.

Steering Arm Length Analysis on Adams

The main goal of simulation using MSC Adams is to find the appropriate Length of Steering
Arm such that Ackermann Rolling Condition is satisfied when the vehicle undergoes a 15 m
radius turn. This was based on FSAE Germany Autocross Track data and data from
Rulebook. Rack Travel and Pinion rotation required to obtain necessary wheel angles are also
found using the result sets of simulation. Actual Lock Angles are also determined from the
results sets. Several iterations were performed to obtain the same.

Ideal vs Actual Behavior of Steering System Deviation from required Ackermann

After the simulation, the result sets were exported to Microsoft Excel for further study. A
graph was plotted between Outer Wheel Angle Vs Inner Wheel Angle. The Ideal values were
obtained from the Ackermann Rolling Condition. The Actual values were obtained from
simulation result sets.

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From the graph it is clear that Ackermann Condition satisfies till Inner Wheel Angle of 6 0
which is necessary for a 15 m radius turn. Then, the Actual curve deviates from the Ideal
curve. Finally, Actual Lock Angles are obtained.

Finite Element Analysis

FEA was carried out for all components in the steering assembly. Shown below are the
results for the V-support that supports the steering wheel and connects it to the chassis.

GEOMETRY: LOADS:
Outer Diameter = 0.5” Normal Load = 2 x Arm Mass x Deceleration
Thickness = 1.5 mm = 171.675 N
Material = AISI 1018
Yield Strength = 350 MPa
Software = SolidWorks 2018 Simulation
Mesh Size = Beam Element – 1

Weight = (Steering Wheel + Quick Release + Steering Column + Bearings + Bearing Support
+ Universal joint + Fasteners) x 9.81 + Arm Weight
= 88.64 N

Own Weight = 1.4715 N

RESULTS:
Stress = 114 MPa
Deformation = 0.095 mm
FOS = 3.07

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BRAKING SYSTEM
Rotors: 220mm diameter Stainless Steel CBR 250 brake discs
Callipers: Twin piston (27.4mm diameter) Bybre floating callipers
Master Cylinders: 12mm diameter, proportioning by bias bar

Brake Calculations:
Target Deceleration is 1.5g. Assuming front: rear mass distribution of 50:50
Normal reaction force at each Front wheel
Rf = 1169.96 N
At each Rear Wheel Rr = 301.536 N
Coefficient of friction between road and tyre, μ1 =1.5
Therefore,
Traction Force on Front Wheel
Tf= Rf μ1=1754.94 N

Traction Force on Rear Wheel Tr= Rr μ1=452.304 N

Tyre Radius is 254mm

So traction Torque on Front Wheel,

Tf’= 1754.94*254= 445754.76 Nmm
Traction Torque on Rear Wheel,
Tr’=452.304 *254=114885.216 Nmm

This is the maximum torque that the wheels can support. We now calculate the braking effort
required.
∴ 2 μ2 Nf r =445.754Nm where
μ2 = 0.42 is the coeff. of friction between the pads and the disc;
r = 90mm is the effective brake disc radius, and
Nf is the normal force applied by the front calipers on the brake disc
Nf=[445.754/(2*0.42*0.09)]
Nf=5896.21 N
Similarly, Nr=1519.6428 N

The pressure in the calipers is found. Assuming only 90% of the pressure applied at the
Master Cylinder acts at the calipers, the pressures at the Caliper Cylinders are:
Pf’=55.14*105 Pa Pr’=14.21*105 Pa
Area of Master Cylinder = 176.71*10-6m2
Therefore, Force required on Front and Rear Cylinder,
Ff= Pf’*A=974.37 N Fr= P ’*A=251.104 N
Brake Pedal Leverage=5.8 Therefore, Brake Pedal Force
→

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POWERTRAIN
The powerhouse of the car is a KTM RC390 engine. It's a 4-stroke naturally aspirated single
cylinder engine. It was chosen due to its lightweight, efficient and easily modifiable for
power gains. In order to adapt the engine to our car, a final drive assembly was designed to
transmit torque to the rear wheels. Fuel metering is handled by an EFI system based around
the ECU.

1-D Engine Simulation

To study the behavior and performance of our engine, we created a 1-D model of our KTM
RC 390 engine on Ricardo WAVE.
Methodology
● To accurately model the stock engine on WAVE, and customised the model to be
compliant for the event.
● To determine the operating conditions of engine to design the intake system
● To observe how the intake runner length and plenum volume affects engine
performance
● Based on Optimum lap simulation, the range of rpm to be tuned was found, and
runner lengths and plenum volume was optimised for the same range.
● To tune the runner length and plenum volume based on Optimum lap simulation,
using the rpm range used most
● Designed and analysed air intake restrictor using cfd.

The purpose of simulating is to acquire performance data of our engine, since we haven’t
tested our engine on a dyno, and to optimize the intake system. Majority of data has been
manually measured from the KTM engine. The fuel type used for the simulation is Indolene.
Surrounding air composition was set as 21% oxygen and 78% nitrogen, 300K temperature at
1 bar pressure. All the ducts used are circular in section.To create the model, intake and
exhaust port lengths and diameters were measured. Simulation was performed from 4000 -
11,000rpm with steps of 500rpm. Temperatures of cylinder and engine data were unavailable,
so preset values were used.

Intake and exhaust port

Discharge coefficient was set to Auto, which Wave calculated based on the geometry of the
port. Discretization length was set to 2mm, which improves accuracy, with the downside of
longer solving time. Wall friction and wall transfer coefficient was left at 1. Composition
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comprises fresh air. The length of the port was calculated as the average of inner and outer
circumference.

Injector
Proportional type of injector is used, since it injects enough fuel to the air charge to match the
targeted air fuel ratio. Initially the air fuel ratio was set to a constant value of 14.58. The
injector is connected to the Y-junction between intake runner and port.

Valves
The reference diameter of the intake valve was set to 36mm and exhaust to 29mm as
mentioned in the user manual. Valve profile data has been manually measured from our
engine. Since the engine was cold during measurement, hot lash of 0.15 mm on intake and
exhaust was added, to compensate for the thermal expansion of camshaft and valve. The
values were taken from the user manual.

Lift of each camshaft was individually measured for 360 degrees rotation, i.e 720 degrees
rotation of crankshaft. The valve open timing with respect to crank rotation was also noted.

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 Valve Lift in mm for 720 degree crank rotation

Cylinder and Engine Block

The cylinder data input to the model is as follows

Stock Engine Output

Torque v/s RPM plot for the stock engine modelVolumetric efficiency v/s RPM plot for the
stock engine model

Restrictor
According to the event rules, a 20mm restrictor is to be implemented before the intake
system. After literature survey, a venturi restriction with 12 and 6 degrees of convergence and
divergence angles was opted. It was observed that having a convergence angle greater than
divergence, showed minimum pressure drop across the restriction. To model the restrictor
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accurately, 3 ducts were used, viz. convergence, throat and divergence section. The diameter
of convergence was matched to the throttle body’s diameter, which is 46mm. Since all
simulations were performed for wide open throttle, throttle element was not added before the
restrictor.

The model was improvised by adding the additional ducts and giving necessary filleting at
the bends , it was also observed that the longer the diffuser( divergence part) resulted in a
better recovery of pressure drop.

Optimizing runner lengths

Runner length was calculated based on Pressure wave theory, which says that air inside the
runner reflects back and forth when the valve is closed, stacking the pressure behind the
valve. The length is tuned such that timing of valve open matches with the final wavefront.

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This phenomenon was also observed in Wave. The frequency and amplitude of the wave
depended upon length and number of reflections for the rpm it was tuned.

The runner was modelled as a duct, with diameter set to 40mm. Length was varied from
160mm to 320mm with increments of 20mm. The right end is connected to a y junction,
where the runner branches into two intake ports, one for each valve.

Plenum
The function of plenum is to recover most of the pressure from the 20mm restriction, and
even out the pulses from the engine during suction. For iterations, this was set as a multiple
of engine displacement volume, i.e 373.2 cc. Iterations were performed for 3, 4 and 5 times
displacement. Torque and volumetric efficiency was observed.

For the plenum, a duct with equal diameter and length is modelled, which was calculated
using the required volume.

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Simulation results

Brake torque values for 3,4 and 5 times engine displacement

as plenum volume v/s runner length at 8000rpm

It was observed that by increasing the volume of plenum, the volumetric efficiency increased
but the peak torque trend was shifted towards left. Hence based on the results, we opt to go
with a runner length of 220mm and 1866cc as plenum volume. Compared to last year’s
plenum, there has been an increase of 284cc, which can be fit in the chassis. The inlet and
outlet diameter of the plenum is matched to the divergence diameter and intake port
respectively.

Brake Power vs Engine rpm

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 Brake Torque vs Engine rpm

Cylinder Pressure Plot for 720 degrees of crankshaft rotation

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 Mass airflow into cylinder vs engine speed

Plenum Volumetric Efficiency v/s Engine rpm

Calculations for intake restrictor

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CFD analysis using Ansys
Mass flow rate of 0.03kg/sec was obtained from the 1-D simulation at the restrictor, which is
input as boundary conditions for the restrictor CFD analysis. Since the velocity calculated at
the throat for 0.03kg/sec is less than 0.3 mach, incompressible flow domain was considered.
For meshing, Inflation meshing was used, to improve accuracy of the computation near the
walls. Only half cross-section was modeled, since axisymmetric condition is used. Edge
sizing and body sizing was provided, specifying the number of elements to the axis, to
increase node count. Node count is 30,000. Boundary conditions - at the inlet, ambient
pressure of 101325 Pa, mass flow rate at the outlet is specified as 0.03kg/second. Solver
method is Green Gauss Node based.

Intake restrictor and plenum mesh

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Results

Static Pressure plot - maximum pressure drop at throat is 4705 Pa

Velocity Plot - Maximum velocity at throat is 83.55m/s

Drivetrain
KTM RC 390 couples a racing performance engine with a high performance sequential 6-speed
transmission. In the OEM application, this engine transmits its power to the ground through a chain
drive to a differential, a solid rear axle or a single rear tire in the motorcycle application. With a simple
sprocket output, it was decided to transmit this power to the wheels in the same fashion it would be in
an ATV. In order to maximize the handling characteristics of the vehicle, it was obvious that a
differential was necessary. Once the rear suspension type had been decided, the rear axles could be
designed. These axles will transmit torque from the differential to the rear hubs, and further to the
wheels.

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Details of the car was loaded into Optimum Lap and then ran through many iterations for different
Final Drive Ratios on same map. The optimal value was then divided by the primary gear reduction.
Ie 10.2/2.886 ~ 3.8

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After considering the transmission gear ratios, competition speeds and the torque curve of the motor,
an ideal range of final drive ratios was determined. Packaging constraints in the vehicle showed that
the largest possible sprocket would be a 49 tooth, which fell at the lower end of the ideal range. Since
a sprocket with the correct number of teeth was in stock, a simple sprocket adapter was manufactured
to adapt this part.

Von Mises Stress plot on the large sprocket (N/m^2) and Displacement plot on the large
sprocket (mm)

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Differential
A differential is a mechanical device that allows the driven wheels of a car to rotate at
different speeds while allowing drive torque to still be applied. Many high performance cars
utilize some type of limited slip differential, which only allows a certain amount of speed or
torque difference between the two output shafts. On high lateral acceleration turns, the inside
wheel may not be able to transmit all of the torque required to accelerate the car forward. In
this case, the limited slip differential forces some of the drive torque to be transmitted to the
wheel with more grip. In some cars, it is beneficial to omit a differential all together. This is
called a solid rear axle, and is usually done for cost reduction and design simplicity.

Figure shows that in order to satisfy the turnability and handling requirements of the vehicle, a
limited slip differential is necessary. There are many different methods of resisting differential
rotation in a drive axle. The most common are clutch pack, viscous, and torsen which each
have their own characteristics under various driving conditions. The torsen differential uses an
intricate set of gears that allow different output shaft speeds, but only a limited amount of
differential drive torque. Once one wheel begins to spin faster than the other, the differential
begins to lock up. The greater the driving torque and differential speeds, the greater the lockup
of the differential. The clutch pack type utilizes clutch plates locked to the output shaft that are
forced against clutch plates locked to the differential housing. This type of differential remains
locked until a large enough torque differential is applied between the output shafts. A clutch
pack LSD benefits from the relatively easy replacement of worn out clutch packs, and the
ability to adjust the torque at which the differential breaks free.

Chain Drive
After the differential, axles and engine had been aligned in the CAD model, the chain drive
could be designed. The chain must be aligned between the engine and differential sprockets.
Another issue that arose was the alignment of the engine with the rear frame tube. The height
of the engine sprocket and the size of the final drive sprocket placed the chain just tangent to
the rear frame tube. In order to keep the chain from running along the tube, the engine mounting
member had to run through the gap enclosed by the chain drive. A 520 chain with pitch of
15.875 mm was selected after considering the tension induced in the chain due to the delivered
torque at larger sprocket.

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Half Shafts
In the TRX350 application, the differential is set slightly off centre to allow for the drive shaft
to pass from the rear to the front of the engine. This means the two output shafts are different
lengths. The shock torque was considered for the design of the half shafts, and the material
used was AISI 4340 annealed Steel. The size of this shaft was chosen based on the maximum
amount of torque that needs to be transmitted and the method of manufacturing for reliability
and weight savings.

Shock Torque = F.O.S * Primary Reduction*Gear Reduction*Final Drive Ratio*Engine

Torque = 648.53 N.m

At τ = 168 MPa,
Diameter of Half Shaft = 26.99 mm ~ 27 mm.

Displacement plot on the half shafts (mm)

Von Mises Stresses plot on the half shaft (N/m^2)

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CHASSIS
Design Goals
The necessary criteria that have to be meet to have a high-performance reliable racing car is
 Maintain Low Center of Gravity
 Use the material that's cost effective and has reasonable manufacturing costs
 Minimize the weight to stiffness ratio
 Create a solid base chassis to evolve on for years to come
 Aesthetically pleasing design
 Implementing right manufacturing technique

Design Procedure
Material Selection: Selection of the material plays a very important role in order to have a
high-performance reliable racing car it should be able to withstand high forces, temperature
and also have enough stiffness to absorb vibrations without undergoing failure. Material
property has to be seen in order to meet the above requirement. A tubular space frame chassis
was chosen over a monocoque chassis despite being heavier because, its manufacturing is
cost effective requires simple tools and damages to the chassis can be easily rectified. The
two very commonly used materials for making the space frame chassis are Chromium
Molybdenum steel (Chromoly) and SAE-AISI 1018. Both these materials were analyzed for
different parameters and finally decided on to use AISI 1018 for making the tubular space
frame chassis
SAE 1018 grade steel is better in terms of Thermal properties but weaker than Chromoly in
terms of strength. But the main priority of our design was torsional stiffness, light weight and
low cost. The density of both the materials is almost the same, as well as the young’s
modulus. This means that the stiffness of the frame would be equal if it was constructed using
either material. Also, the cost of Chromoly 4130 Steel is more than the cost of AISI 1018, as
well as the fact that normalizing procedures are required to increase the weld strength of
4130. Due to these factors and the priority of the team being maximum utilization of scarce
resources, AISI 1018 was chosen. The only negative in using AISI 1018 would be the lower
yield strength, however it could be compensated by a better design with increased
triangulation.

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CAD Modelling
Software used: SOLIDWORKS
Front and rear sections: The initial design begins of chassis begins from marking the
suspension hard points and creating the front and rear suspension box then all other
components must work around these points. Figure 1 shows these nodes, joined by lines to
make the points easier to see. To allow room for fixings, the A-arms do not mount directly to
the nodes but are is mounted as close as is practical so that minimal bending is introduced
into the chassis. Minimizing potential deflection is essential for each suspension mounting as
this ensures the wheel does not move and change its geometry under load. . If the wheel’s
geometry were to change under load the suspension may become difficult to tune and
optimize as the wheel would move away from the position which provides the best grip for
the tire

Suspension Boxes
Roll hoops: With the suspension pivot and spring locations now defining front hoop and
main hoop is done keeping the rules in mind figure 2 shows the roll hoops design. Front hoop
height 575mm main hoop height 1080.05

Roll Hoops
Side impact structure: Rule says that the vehicle needs to have at least 3 side impact
structure members for the protect driver from side on collision The FSAE rules require a side
impact structure to be present in the frame to protect the driver in the event of a side on
collision. It consists of two horizontal members and one diagonal member; the side impact
structure connects the front and rear roll hoops. Figure shows side impact structure. The
upper member must be at a height of 240mmto 320mm above the lowest chassis inside point
between front and main hoop. This height corresponds to the height of the upper chassis
members in the front and rear sections so the upper side impact member can simply connect
these sections. Lowest point of lowest member of sis is 49mm

Side Impact Structure

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Front bulkhead: Front
bulkhead is designed with a
proper dimension to ensure that
the IA of required dimension
fits and all the necessary
conditions are meet. Height
304.8mm Width 304.80mm.
Figure 4 shows the front
bulkhead

Design Iterations
Several iterations were carried out to meet the design criteria and the flaws were addressed in
each successive iteration. After around 20-25 iterations we arrived with the current design
and the best design was selected which was light weight and had required stiffness
1. 2. 3.

4. 5. 6.

Figure 6. Design Iterations

Finite Element Analysis
Software Package: SOLIDWORKS finite element analysis package was utilized to design
and optimize the chassis. Many different chassis designs were analyzed using the same
flexure and torsion tests in order to minimize the amount of frame members used while still
maintaining proper stiffness. The torsion analysis was performed by fixing These steps are
carried out in iteration to get the chassis with good performance characteristics.
For torsional analysis, the rear suspension box is fixed and applying a couple to the front
suspension box. The flexure analysis was performed by fixing the front suspension box and

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the rear suspension box and applying distributed load along the bar at the base of the main
roll hoop.
Mesh Specifications
Beam mesh model; 20mm element size; No. of elements =1669
Finer mesh is created using mesh control tool
Torsional stiffness
Initial conditions
 Moment arm= L=600 mm
 Constraints: front Pick-up points
 Force applied F = 500
Observations
 Total displacement(Y) = 4.078+4.037 = 8.115 mm
 Maximum stress = 9.495e+07 N/m^2
 FOS 3.97

Calculations
Angular displacement,
𝑌 8.115
𝜃 = tan−1 ( ) = tan−1 ( )
2𝐿 2 × 600
𝜃 = 0.3875°
Torsional stiffness, K, 𝐾 =
𝑀𝑜𝑚𝑒𝑛𝑡 𝑎𝑝𝑝𝑙𝑖𝑒𝑑 500×1.2
=
𝐴𝑛𝑔𝑢𝑙𝑎𝑟 𝑑𝑖𝑠𝑝𝑙𝑎𝑐𝑒𝑚𝑒𝑛𝑡 0.3875

K = 1548.38 Nm/deg
Specific Stiffness = 51.85
Nm/deg/kg (taken mass of the
chassis, m = 29.86kg)
Results
 Torsional stiffness, K = 1548.38
Nm/degree
 Specific Stiffness = 51.85
Nm/deg/kg Displacement under Torsional Loading
 FOS = 3.97

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Torsional Strength
 Initial conditions
 Applied force = 3G at front pick-up
point
 Constraints: Rear Pick-up points
Observations and results
 Maximum stress = 2.86e+08 N/m^2
 Maximum displacement = 19.36
mm
 FOS obtained = 1.22

Stress Under Torsional Loading

Static verticle bending

Initial conditions
Load: weight of all components at the respective point of contact
Results
 Maximum stress = 1.085e+08 N/m^2
 Maximum displacement = 1.11 mm
 FOS obtained = 3.2

Stress Under Vertical Bending Load

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 FEA RESULTS

Sl. Loading Condition Load Load Max. stress Max. FOS

No. (G) (N) (MPa) Deformation
(mm)
1 Acceleration 0.7G 3233 38.54 0.942 9
2 Braking 2.5G 7131.84 68.52 1.67 5.1
3 Cornering 1.5G 4775 54.54 0.53 6.4
4 Vertical Bending - 1G 2686.7 108.5 1.11 3.2
sag
5 Torsional strength 3G 7946.1 288.6 1.9 1.2

Force calculations
Note: For body force calculations mass of the car is taken as 270 kg
1) Acceleration (max. acceleration is assumed to be 0.7G)
Net force = weight of the car + inertia force due to acceleration
F = √(270 × 9.81)2 + (270 × 0.7 × 9.81)2
= 3233.15 N
2) Braking (Max. retardation is assumed to be 2.5G)
Net force = weight of the car + inertia force due to retardation
𝐹 = √(270 × 9.81)2 + (270 × 2.5 × 9.81)2
F = 7131.84 N
3) Cornering (Max. cornering force is assumed to be 1.5)
Net force = weight of the car + inertia force due to cornering
𝐹 = √(270 × 9.81)2 + (270 × 1.5 × 9.81)2
F = 4775 N
Final Design Summary
Finalized Design
The final design meets all our requirements its high stiffness to weight ratio, low cost, easy to
manufacture, highly reliable and meets. It makes the most of the members required by the
rules, requiring few additional members to support loads and add triangulation. The design
utilizes the required battery firewalls as a structural component to make the most of the
weight that they add to the chassis. The weight of the
complete space-frame is 29.72kg. with centre of
gravity 377mm. The design includes good driver
ergonomics and it is easy for the driver to climb in
and out of the car due to the low roll mounted hoop
supports. The chassis is 2428.43mm long, 656.64mm
wide, ground clearance 49mm and the roll hoop is
1092.7mm tall. Torsional stiffness of chassis is
1548.38 Nm/deg.
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AERODYNAMICS
Goals
 To reduce drag on the vehicle as much as possible.
 Integrate a diffuser to increase downforce at the cost of minimal drag.

Nosecone
The nosecone is the aerodynamic bulkhead area in the front of the driver’s feet. Its function is
to deflect the air into the required places which would optimize the working of other
aerodynamic devices in the car such as sidepods or undertray. Its design promotes undertray
and diffuser’s airflow optimization, promoting reduced drag and increasing downforce
potential.

Diffuser
The undertray is not only the largest aerodynamic component on a FSAE car, it’s also the
most aerodynamically efficient, producing nearly 9 times more downforce per unit of drag
than a rear wing. It is able to take advantage of the phenomenon of ground effect. The
objective of the diffuser is to slow the flow back down again and to give the used air flow
from the undertray of the car as much possible space to exit from the rear end. This means
that if the air can escape more easily from under the car, then more air at faster velocities can
flow under the undertray of the car creating a lower pressure and therefore higher downforce.
The diffuser increases in volume along its length, creating a void that has to be filled by the
air passing under the body.

Design Considerations
The following parameters must be optimized in the design phase:
1. Airflow: As mentioned above the nosecone must deflect the air to the required places.
Primarily the nosecone must direct air both to the sides to the radiators and down to the
undertray to increase downforce. The designed undertray must be integrated with the
nosecone to optimize the intake into the venturi tunnels created by the diffuser.
2. Drag: The nose cone must have a low value of Cd (coefficient of drag). This can be
achieved by having smooth variation in curvature, minimizing stagnation pressure when
the airflow makes contact with the nose cone.
3. Downforce: Usually nose cones with low drag generate a small amount of lift. Nosecones
which produce downforce on the other hand usually have larger Cd values. Hence a
compromise has to be made between the two.
4. Weight: It is essential to minimize the weight of the nose cone as much as possible. This
can be achieved by making as compact a nose cone as possible and having flat surfaces
wherever possible. The material used for the nosecone and undertray also make a
difference as carbon fiber weighs less than glass fiber which in turn weighs less than
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 aluminum. But they are also inversely more expensive so as a middle ground glass fiber is
chosen.

Design Methodology
The design is largely CFD driven. One must make approximatemodels of the nosecone and
diffuser and subject them to CFD Simulation and make changesbased on these results while
applying the knowledge of aerodynamics.

DESIGN
Nosecone
There exist mainly 3 types of nosecones.
 Horizontal nose cone: The horizontal nosecone uses a horizontal edge to deflect the
air downwards. This is mainly useful to feed the undertray of the car. Since our team
is not employing an undertray. This type of a nosecone would not be useful.

Horizontal nosecone used by TU Graz

 Vertical nosecone: A vertical nose cone uses a vertical edge to deflect the air towards
the sidepods, very little downforce would be produced in such a nosecone. However,
presence of a vertical edge would also increase the drag

Vertical nosecone used by Lancia motorsports

 Angled nosecone: An angled nosecone is an improvement upon the vertical nosecone
type which gives has a slightly angled edge than a vertical edge. This reduces the drag
as well as reduces the lift generated. This type of nosecone was found to be most
suitable for the car. Design iterations were based on this nosecone.

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 Final nosecone CAD model
The nose once was made as compact as possible by conforming closely to the chassis. Angled
nosecone was the design principle chosen. The purpose of this design is to use again a
vertical edge in order to “split” the incoming air to the sides for the sidepods and use the
angle of that edge in order to reduce the pressure of the stagnation point and give more space
to the underside of the nosecone to “feed” the undertray. Various parameters were
approximated and the final values were determined by CFD Simulations in addition to
looking at various teams which have successfully participated in Formula student
competitions. The top surface of the nosecone also has peaked lines and a cockpit deflector to
increase downforce, as shown in figure.

Undertray
The next step after the selection of the final nosecone is the creation of a CFD model of
undertrays that were designed. Two main types of undertray are designed and simulated at
the speed of 80km/h while the ride height from the ground is 35mm for all cases. The aim of
all designs is to achieve the highest amount of downforce, while keeping drag as low as
possible. The weight of each undertray is also taken under consideration in order to estimate
the aerodynamic efficiency of each type comparing to its mass.
1. Single diffuser: The first undertray model has a large single diffuser at the rear of the
undertray. Its dimensions are restricted from the rules which set a standard safe distance
from the wheels. The maximum available space for the angle of the diffuser at the rear
side of the chassis is also restricted from the bulk-head. This first iteration was just a
design exercise and was hence set as a placeholder from which to improve on with the
next iteration.

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2. Side Diffusers: The next type of undertray is completely different than the previous
designs. It consists of two side diffusers which have an angle of 17.5° each. This gave
more space for the central diffuser, which is basically a curvature fillet with an angle of
12°, with 7 vertical flaps.

However, at that case, the angle of the side diffusers is restricted from the suspension
dumpers that are located just above the diffusers’ exits. To overcome this problem and
increase the downforce without increasing the diffusers’ angles, the diffuser tunnels are
designed in such a way that they can take advantage of the Venturi effect in two
dimensions, vertically and horizontally. As it is shown from the floor plan of the
undertray, the shape of the diffuser tunnels is similar to that of a Venturi tube.

CFD Result
Nosecone

The Nosecone model was simulated using ANSYS Fluent. The nose cone was placed in an
enclosure for simulation. This mimics the nose cone being placed in a wind tunnel. Further
two sub-enclosures were defined as bodies of influence(boi) were defined near the nose cone
to increase number of elements close to the nose cone. The inner most enclosure is boi-
nearfield with the finest mesh of 8mm for accurate simulation in the critical regions, with a
courser mesh in the boi-farfield of 16mm, followed by the coarsest mesh in the enclosure.

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 Nosecone geometry with enclosure

Boundary conditions: Air was simulated to enter the wind tunnel inlet at a speed of 22.22 m/s
or 80km/h. A pressure boundary condition was given at the outlet. The pressure was set to
zero-gauge pressure or atmospheric pressure. Turbulence model: Since the flow around the
car is mostly turbulent a suitable turbulence model has to be chosen. The k-omega model was
chosen as it gives the most realistic results and is usually used in most literature relating to
FSAE.
Results: The Drag coefficient is 0.30696 as shown on
Characteristics
the table below, while the lift coefficient is at -0.30680
which results in an aerodynamic efficiency of 0.99.
Downforce 13.139N
Although the drag coefficient is slightly higher for this
type of nosecone, the model with the combination of Drag 13.146N
peaked lines and cockpit deflector on the top surface,
produced a high efficiency. Although the aim of the Cl -0.30680
nosecone’s design is the minimization of drag the final
model that is selected has a high drag coefficient but Cd 0.30696
considering that it has a high lift coefficient its
aerodynamic efficiency is satisfactory. Efficiency (Cl/Cd) 0.99947

Weight 16.3kg

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 Pressure distribution and velocity streamlines around the final nosecone
The pathlines of flow around the nosecone is shown in the figures. The pathlines are coloured
according to their velocity. One can observe that there is no separation along the nosecone.
Separation can be induced if the curvature of the nose cone is in excess. This can cause an
increase in pressure drag of the nosecone.

Undertray

Similar to nosecone, the same conditions were applied and simulation was conducted for the
undertray in ANSYS Fluent.
Results: The downforce obtained is 56.36N with an efficiency of 9.321, with a drag of
6.05N. Although this is a decent result for a first attempt, the weight penalty is on the higher
side as a result of using glass fiber. Further iterations must be made to increase downforce
and reduce weight as much as possible. The pressure contours show lower pressures in the
venturi tunnel as a result of the faster airflow shown in the velocity streamline and the
velocity contours.

Pressure contour below and above the undertray

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 Velocity Streamlines and Contours

Characteristics

Number of diffusers 3
Conclusion
Side diffuser angle 17.5
 Satisfactory nosecone designed
 Undertray needs further refinement to reduce Central diffuser 12
weight and increase downforce angle
 Use of carbon fibre must be explored subject to
Downforce 56.36N
budget available
Drag 6.05N

Cl -8.100

Cd 0.869

Aerodynamic 9.321
Efficiency

Weight 9.7kg

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Ergonomics - RULA and REBA Analysis
 To check the comfort of driver RULA AND REBA Analysis was performed using CATIA V5 and
ERGO PLUS to get the proper score so that we can assure that our chassis can accommodate
95th percentile driver. The chassis was designed in such a way that driver has maximum
comfort with proper view angle, lap angle, knee angle and seat angle. REBA SCORE 2, RULA
SCORE 3

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ELECTRICAL SYSTEM
Electrical Architecture

Electrical System Architecture

The above diagram sums up the electrical architecture in the car. From the power distribution
board, we have a regulated 5V and 12V supply. The starter motor is energised using the starter
solenoid. Sensors are provided with the 5V supply. The ECU (Haltech Elite 1500) is powered
from the 12V supply. ECU takes in data from the sensors and commands the actuators in order
to make sure the sensor values are in alignment with the engine maps provided. The data from
the ECU is then displayed on the dashboard (driver interface). The shutdown circuit is a safety
circuit that kills the car when the driving conditions become unsafe. Various sensors are used
such as brake pressure sensors, oil temperature sensor, Camera (GoPro) (for FPV video),
Accelerometer and Gyroscope sensor and wheel speed sensors. Raspberry-Pi is used as the
master device for data logging and transmission of data to the pit station.

Haltech Elite 1500 ECU

ECU data paths and control

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Dashboard
It is a modular system component that forms the vital visual interface between the driver and
the state of the car.

A highly integrated indicator system. Its main purpose is to pinpoint faults in the safety system
and effectively communicate various other information to the driver. These system information
include
● Speed.
● Gear position.
● Critical system warning lights.
● Oil temperature sensor.
● Tachometer.

It'll also include various switches such as

● Start switch
● Kill switch

Dashboard interface overview

At a preliminary level these are all the sensors and interfaces that are planned to be used.
The dimensions of the dashboard are obtained from car chassis design. As a part of testing a
cardboard board can be used to check the placement of components and finalize the design.
Doing this would give the largest surface area to work with and optimize the placement (of
components) with utmost flexibility and also help with future improvements and add-ons.

Placement of Components
At a high level overview placement is done considering factors such as
● Easy access to elements such that the components such as shutdown buttons, other
critical system information are not obstructed by and the driver is easily able to access
the needed components. (especially important in case of cockpit shutdown button for
safety)
● Logical placement of components.

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Shutdown Circuit Schematic
The main purpose of the shutdown circuit is to shut down the car whenever unsafe driving
conditions occur. The circuit consists of a series of switches in which an open circuit would
break the power supply to the ignition/injection and the fuel pump, ultimately killing the
engine. The shutdown circuit consists of Battery, Low Voltage Master Switch, BSPD, Inertia
Switch, Cockpit Shutdown Button, Left Shutdown Button, Right Shutdown Button and BOTS.
Mechanical relays are used to actuate the switches, the fuel pump and injection. The below
schematic illustrates the shutdown circuit with the BPSD faults actively participating in the
circuit.

Shutdown Circuit Schematic

Brake System Plausibility Device (BSPD) Schematic

The BSPD plays a crucial part in the shutdown circuit. It is a standalone and non-programmable
circuit. The BSPD actuates the shutdown circuit by analysing the data from the brake pressure
sensors in order to detect hard braking. If hard braking is detected the BSPD immediately opens
the shutdown circuit thus killing the engine. With the usage of two brake pressure sensors, if
the implausibility in the values is persistent for more than 500ms, the BSPD taes action. The
threshold for the brake system pressure sensor to detect hard braking is set to <=30 bar. The
below schematics of the BSPD consists of two input ports to read the data from two brake
pressure sensors. The data is then sent to a comparator block in order to check for threshold
and implausibilities in values. A timer of duration 500 ms is then activated if an implausibility
is detected and if the error remains for more than 500ms, the relay opens the shutdown circuit,
thus killing the engine.

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 BSPD Schematic

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