KATHMANDU UNIVERSITY

SCHOOL OF ENGINEERING
DEPARTMENT OF MECHANICAL ENGINEERING

PROJECT REPORT ON

DESIGN AND ANALYSIS OF DOUBLE WISHBONE FRONT SUSPENSION SYSTEM

Manish Aryal [42062]

Rabin Pradhan [42095]

Sagar Chand Thakuri [42449]

September 2016
 AUTHORIZATION

I hereby declare that we are the author of the project.

I authorize the Kathmandu University to lend this thesis to other institutions or individuals for
the purpose of scholarly research. I further authorize the Kathmandu University to reproduce the
thesis by photocopying or by other means, in total or in part, at the request of other institutions or
individuals for the purpose of scholarly research.

___________________________________________
Rabin Pradhan [42095]

September 2016
 PROJECT EVALUATION

DESIGN AND ANALYSIS OF DOUBLE WISHBONE FRONT SUSPENSION SYSTEM

By

Manish Aryal [41062]

Rabin Pradhan [41095]
Sagar Chand Thakuri [41449]

This is to certify that I have examined the above Project report and have found that it is complete
and satisfactory in all respects, and that any and all revisions required by the report examination
committee have been made.

_________________________________________
Dr. Daniel Tuladhar
Project Supervisor
Department of Mechanical Engineering

_________________________________________
Mr. Binaya Baidar
Project Co-supervisior
Turbine Testing Lab

_________________________________________
Dr. Daniel Tuladhar
Project Co-ordiantor
Department of Mechanical Engineering

September 2016
 ACKNOWLEDGEMENTS
We wish to acknowledge Kathmandu University, School of Engineering, Department of
Mechanical Engineering for allowing us to do our project and providing necessary machineries
and equipment to complete our project. We are grateful to our Supervisor Dr. Daniel Tuladhar
for guiding us in this project. We are also thankful to Mr. Krishna Prasad Shrestha, Mr. Gokarna
Poudel, Mr. Suman Karki for helping us to complete our project. We are thankful to our friends
Mr. Kamal Sapkota, Mr. Sabin Bhattarai, Mr. Laxman Bhatta, Mr. Saroj Neupane, Mr. Sweekar
Dhakal, Mr. Biraj Dhakal for helping and sharing knowledge.
 ABSTRACT

This project deals with design and analysis of the double wishbone suspension parameters. The

suspension design is critical to control the vehicles in motion. The vehicle while in motion

encounters pitching, rolling, bouncing, yawing. These motions are initiated from the suspension

system. In this project double wishbone type of suspension system is considered for the analysis.

We hierarchically structured design items from design variables that represent suspension

geometry to evaluation criterion related to practical operation situations. Some crucial

parameters affecting both static and dynamic stability of vehicle are considered for the analysis

and results are tabulated. Finally we showed that the optimal design solutions can be obtained.
 Table of Contents
ABSTRACT .................................................................................................................................... v

LIST OF FIGURES ..................................................................................................................... viii

LIST OF ABBREVIATION .......................................................................................................... ix

CHAPTER 1 INTRODUCTION .................................................................................................... 1

1.1 Background ...................................................................................................................... 1

1.2 Objectives ............................................................................................................................. 2

1.3 Scope of the Project.......................................................................................................... 2

CHAPTER 2 METHODOLOGY ................................................................................................... 3

2.1 Conceptual Framework ......................................................................................................... 3

2.2 Study Design ......................................................................................................................... 4

2.2.1 Literature Review........................................................................................................... 4

2.2.2 Design and Calculation ................................................................................................ 11

2.2.3 Analysis........................................................................................................................ 11

2.2.4 Report Submission ....................................................................................................... 12

CHAPTER 3 DISCUSSION ......................................................................................................... 12

3.1 Work Accomplished: .......................................................................................................... 12

3.2 Calculations......................................................................................................................... 12

3.4 Final Design ........................................................................................................................ 20

3.5 Stress and Deflection Analysis in ANSYS ......................................................................... 20

3.5.1 Analysis of spring………………………………………………………………… 20

3.5.2 Analysis of Control Arm:............................................................................................. 21

3.6 Experimental Verification ................................................................................................... 22

3.7 Dynamic Simulation in ADAMS ........................................................................................ 23

3.7.1 Analysis of Change in Camber Angle with Wheel Travel .......................................... 26

 3.7.2 Analysis of Change in Toe Angle with Wheel Travel ................................................. 27

3.7.3 Analysis of Change in Spring Stiffness with Wheel Travel ........................................ 27

CHAPTER 4 GHANTT CHART ................................................................................................. 29

CHAPTER 5 BUDGET ................................................................................................................ 30

CHAPTER 6 LIMITATION ......................................................................................................... 31

CHAPTER 7 RECOMMENDATION .......................................................................................... 31

CHAPTER 8 CONCLUSION....................................................................................................... 32

REFRENCES ................................................................................................................................ 33

ANNEX......................................................................................................................................... 34
 LIST OF FIGURES
Fig 1.1 Double wishbone suspension system…………………………………………..1
Fig 2.1.1 Description of roll center……………………………………………………..3
Fig 2.2.1 Flow Chart of Methodology………………………………………………….4
Fig 2.2.1.1 Leaf Spring…………………………………………………………………6
Fig 2.2.1.2 Macpherson Strut Suspension System……………………………………..7
Fig 2.2.1.3 Double Wishbone Suspension System……………………………………..8
Fig 2.2.1.4 Multilink Suspension System………………………………………………9
Fig 2.2.1.5 Semi Trailing Arm………………………………………………………...10
Fig 3.2.1 Static Axle load on vehicle…………………………………………………..12
Fig 3.2.2 Forces acting on a vehicle during braking…………………………………...12
Fig 3.2.3 Forces acting on a vehicle during cornering…………………………………13
Fig 3.2.4 Forces acting on a vehicle on a downhill grade……………………………...14
Fig 3.2.5 Kinetic analysis of double wishbone suspension system…………………….15
Fig 3.2.6 Notation used in wishbone design……………………………………………16
Fig 3.2.7 Determination of roll center…………………………………………………..16
Fig 3.2.8 Quarter Car Model…………………………………………………………….18
Fig 3.4.1 Orthographic View of Double Wishbone Suspension System………………..20
Fig 3.5.1 Analysis of Spring……………………………………………………………..21
Fig 3.5.2 Analysis of Control Arm……………………………………………………….22
Fig 3.6.1 Orthographic View of Test Rig for Double Wishbone Suspension System…...23
Fig 3.7.1: Setup of Double Wishbone Suspension For Dynamic Analysis in ADAMS….25
Fig 3.7.2 Graph Showing Change in Camber Angle with Wheel Travel…………………27
Fig 3.7.3 Graph Showing Change in Toe Angle with Wheel Travel……………………..27
Fig 3.7.4 Graph Showing Change in Spring Stiffness with Wheel Travel………………..28
 LIST OF ABBREVIATION
ADAMS Automatic Dynamic Analysis of Mechanical System
CG Center of Gravity
ICR Instantaneous Center Rotation
RC Roll Center
d Offset of upper arm
e Offset of lower arm
H Distance of upper ball joint from ground
h Distance between upper ball joint and lower ball joint
Z Lift of the wheel center
Z1 Lift of upper arm
Z2 Lift of lower arm
Y1 Lateral displacement of upper ball joint
Y2 Lateral displacement of lower ball joint
R1 Length of upper arm
R2 Length of lower arm
a Vertical distance of upper ball joint with respect to inboard point
b Vertical distance of lower ball joint with respect to inboard point
 CHAPTER 1 INTRODUCTION

1.1 Background
The double wishbone suspension can also be referred to as double 'A' arms, and short long arm
(SLA) suspension if the upper and lower arms are of unequal length. A single wishbone or A-
arm can also be used in various other suspension types, such as MacPhersonstrut and Chapman
strut. The upper arm is usually shorter to induce negative camber as the suspension jounces
(rises). When the vehicle is in a turn, body roll results in positive camber gain on the inside
wheel. The outside wheel also jounces and gains negative camber due to the shorter upper arm.
The suspension designer attempts to balance these two effects to cancel out and keep the tire
perpendicular to the ground. This is especially important for the outer tire because of the weight
transfer to this tire during a turn.
In automobiles, a double wishbone (or upper and lower A-arm) suspension is an independent
suspension design using two (occasionally parallel) wishbone-shaped arms to locate the wheel.
Each wishbone or arm has two mounting points to the chassis and one joint at the knuckle. The
shock absorber and coil spring mount to the wishbones to control vertical movement. Double
wishbone designs allow the engineer to carefully control the motion of the wheel throughout
suspension travel, controlling such parameters as camber angle, caster angle, toe pattern, and roll
center height, scrub radius, scuff and more.[3]

Fig 1.1 Double wishbone suspension system [1]

1
1.2 Objectives
 To design double wishbone suspension system.
 To calculate required parameter affecting the stability of vehicle.
 Performance analysis of double wishbone suspension system.

1.3 Scope of the Project

The suspension systems mostly used in cars of Nepal are Macpherson suspension system which
has stability issues during cornering and has small displacement. To overcome this problem we
are studying about wishbone suspension system.

2
 CHAPTER 2 METHODOLOGY

2.1 Conceptual Framework

Determination of roll centre plays a very important role in deciding the wishbone lengths, tie rod
length and the geometry of wishbones. Roll Centre in the vehicle is the point about which the
vehicle rolls while cornering. The location of the geometric roll centre is solely dictated by the
suspension geometry, and can be found using principles of the instant centre of rotation. Roll
centre and ICR is determined because it is expected that all the three elements- upper wishbone,
lower wishbone and tie rod should follow the same arc of rotation during suspension travel. This
also means that all the three elements should be displaced about the same centre point called the
ICR.

Fig 2.1.1 Description of roll center [1]

The vehicle centre line is drawn. The end points of wishbones are joined together to visualize the
actual position of the wishbones in steady condition. When the lines of upper and lower
wishbones are extended, they intersect at a certain point known as Instantaneous Centre (ICR). A
line is extended from ICR to a point at which tire is in contact with the ground. The point at
which this line intersects the vehicle center line is called the Roll Centre. Now, extend a line
from ICR point to the steering arm. This gives exact tie rod length in order to avoid pulling and
pushing of the wheels when in suspension
When the RC is far away from CG (lower RC), when the car corners the CG has more leverage
on the RC, so the car will roll more.
When the RC is closer to CG (higher RC), when the car corners the CG has less leverage on the
RC, so the car will roll less.
If the RC was right on top of the CG, when the car corners the CG has no leverage on the RC, so
the car would not roll at all.

3
2.2 Study Design

Literature Review Survey Design and

Calculation

Analysis
Report
Submission

Fig 2.2.1 Flow Chart of Methodology

2.2.1 Literature Review

The need for suspension
The study of the forces at work on a moving car is called vehicle dynamics. Some of the
concepts are needed to be understood in order to appreciate why a suspension is necessary in the
first place. Most automobile engineers consider the dynamics of a moving car from two
perspectives:

Ride- A car's ability to smooth out a bumpy road

Handling- A car's ability to safely accelerate, brake and corner.

These two characteristics can be further described in three important principles - road isolation,
road holding and cornering. These principles are explained below and also how engineers
attempt to solve the challenges unique to each. [7]

Road Isolation- The vehicle's ability to absorb or isolate road shock from the passenger
compartment. Allow the vehicle body to ride undisturbed while traveling over rough roads.
Absorb energy from road bumps and dissipate it without causing undue oscillation in vehicle. [7]

Road Holding- The degree to which a car maintains contact with the road surface in various
types of directional changes and in a straight line (Example: The weight of a car will shift from
the rear tires to the front tires during braking. Because the nose of the car dips toward the road,

4
this type of motion is known as "dive". The opposite effect "squat" occurs during acceleration,
which shifts the weight of the car from the front tires to the back. Minimize the transfer of
vehicle weight from side to side and front to back, as this transfer of weight reduces the tire's
grip on the road.[7]

Cornering- It is the ability of a vehicle to travel a curved path. Minimize body roll, which occurs
as centrifugal force pushes outward on a car's center of gravity while cornering, raising one side
of the vehicle and lowering the opposite side. Transfer the weight of the car during cornering
from the high side of the vehicle to the low side. [7]

DEPENDENT FRONT SUSPENSIONS

Dependent front suspensions have a rigid front axle that connects the front wheels. Basically, this
looks like a solid bar under the front of the car, kept in place by leaf springs and shock absorbers.
Common on trucks, dependent front suspensions haven't been used in mainstream cars for years.

LEAF SPRING
A leaf spring is a simple form of spring commonly used for the suspension in wheeled vehicles.
Originally called a laminated or carriage spring, and sometimes referred to as a semi-elliptical
spring or cart spring, it is one of the oldest forms of springing, dating back to medieval times. A
leaf spring takes the form of a slender arc-shaped length of spring steel of rectangular cross-
section. In the most common configuration, the center of the arc provides location for the axle,
while tie holes are provided at either end for attaching to the vehicle body. For very heavy
vehicles, a leaf spring can be made from several leaves stacked on top of each other in several
layers, often with progressively shorter leaves. Leaf springs can serve locating and to some
extent damping as well as springing functions. While the interleaf friction provides a damping
action, it is not well controlled and results in stiction in the motion of the suspension. For this
reason some manufacturers have used mono-leaf springs. A leaf spring can either be attached
directly to the frame at both ends or attached directly at one end, usually the front, with the other
end attached through a shackle, a short swinging arm. The shackle takes up the tendency of the
leaf spring to elongate when compressed and thus makes for softer springiness. Some springs

5
terminated in a concave end, called a spoon end (seldom used now), to carry a swivelling
member.

Fig 2.2.1.1 Leaf Spring [7]

INDEPENDENT FRONT SUSPENSIONS

In this setup, the front wheels are allowed to move independently. While there are several
different possible configurations, this design typically uses two wishbone-shaped arms to locate
the wheel. Each wishbone, which has two mounting positions to the frame and one at the wheel,
bears a shock absorber and a coil spring to absorb vibrations. Double-wishbone suspensions
allow for more control over the camber angle of the wheel, which describes the degree to which
the wheels tilt in and out. They also help minimize roll or sway and provide for a more consistent
steering feel. Because of these characteristics, the double-wishbone suspension is common on the
front wheels of larger cars.

MACPHERSON STRUT
The most widely used front suspension system in cars comprises of a strut-type spring and shock
absorber combo, which pivots on a ball joint on the single, lower arm. The steering gear is either
connected directly to the lower shock absorber housing, or to an arm from the front or back of
the spindle. In this case, when you steer, it physically twists the strut and shock absorber housing
and consequently the spring to turn the wheel.

6
 Fig 2.2.1.2 Macpherson Strut Suspension System [7]

DOUBLE WISHBONE SUSPENSION

The double wishbone suspension can also be referred to as double 'A' arms, and short long arm
(SLA) suspension if the upper and lower arms are of unequal length. A single wishbone or A-
arm can also be used in various other suspension types, such as MacPherson strut and Chapman
strut. The upper arm is usually shorter to induce negative camber as the suspension jounces
(rises). When the vehicle is in a turn, body roll results in positive camber gain on the inside
wheel. The outside wheel also jounces and gains negative camber due to the shorter upper arm.
The suspension designer attempts to balance these two effects to cancel out and keep the tire
perpendicular to the ground. This is especially important for the outer tire because of the weight
transfer to this tire during a turn.
Between the outboard end of the arms is a knuckle with a spindle (the kingpin), hub, or upright
which carries the wheel bearing and wheel. Knuckles with an integral spindle usually do not
allow the wheel to be driven. A bolt on hub design is commonly used if the wheel is to be driven.
In order to resist fore-aft loads such as acceleration and braking, the arms need two bushings or
ball joints at the body.

7
 Fig 2.2.1.3 Double Wishbone Suspension System
At the knuckle end, single ball joints are typically used, in which case the steering loads have to
be taken via a steering arm, and the wishbones look A or L-shaped. An L-shaped arm is
generally preferred on passenger vehicles because it allows a better compromise of handling and
comfort to be tuned in. The bushing in line with the wheel can be kept relatively stiff to
effectively handle cornering loads while the off-line joint can be softer to allow the wheel to
recess under fore aft impact loads. For a rear suspension, a pair of joints can be used at both ends
of the arm, making them more H-shaped in plan view.
Alternatively, a fixed-length driveshaft can perform the function of a wishbone as long as the
shape of the other wishbone provides control of the upright. [3]

MUTLI-LINK SUSPENSION SYSTEM

A multi-link suspension is a type of vehicle suspension design typically used in independent
suspensions, using three or more lateral arms, and one or more longitudinal arms. These arms do
not have to be of equal length, and may be angled away from their 'obvious' direction.
Typically, each arm has a spherical joint (ball joint) or rubber bushing at each end.
Consequently, they react on loads along their own length, in tension and compression, but not in
bending. Some multi-links do use a trailing arm or wishbone, which has two bushings at one end.
On a front suspension one of the lateral arms is replaced by the tie-rod, which connects the rack
or steering box to the wheel hub.

8
 Fig 2.2.1.4 Multilink Suspension System
Multi-link suspension allows the auto designer the ability to incorporate both good ride quality
and good car handling in the same vehicle. In its simplest form the multi-link suspension is
orthogonal - that is, it is possible to alter one parameter in the suspension at a time, without
affecting anything else. This is in direct contrast to a double wishbone suspension where moving
a hard-point or changing a bushing compliance will affect two or more parameters. [7]

TRAILING AND SEMI-TRAILING

A trailing-arm suspension is an automobile suspension design in which one or more arms (or
"links") are connected between (and perpendicular to and forward of) the axle and the chassis. It
is usually used on rear axles. A 'leading arm' as used on a Citroën 2CV, has an arm connected
between (and perpendicular to, and to the rear of) the axle and the chassis. It is used on the front
axle.
Trailing-arm designs in live axle setups often use just two or three links and a Panhard rod to
locate the wheel laterally. A trailing arm design can also be used in an independent suspension
arrangement. Each wheel hub is located only by a large, roughly triangular arm that pivots at one
point, ahead of the wheel. Seen from the side, this arm is roughly parallel to the ground, with the
angle changing based on road irregularities.
A semi-trailing arm suspension is a supple independent rear suspension system for automobiles
where each wheel hub is located only by a large, roughly triangular arm that pivots at two points.
Viewed from the top, the line formed by the two pivots is somewhere between parallel and
perpendicular to the car's longitudinal axis; it is generally parallel to the ground. Trailing-arm
and multilink suspension designs are much more commonly used for the rear wheels of a vehicle
where they can allow for a flatter floor and more cargo room.

9
 Fig 2.2.1.5 Semi Trailing Arm [7]

SUSPENSION TERMINOLOGY
Camber: This is the angle of the rim/tire from vertical as viewed from the front or the rear of the
car. Be sure the wheels are pointed straight ahead when measuring this angle. [7]
Caster: This is the angle of the steering axis as viewed from the side of the car. The axis may
pass through upper and lower ball joints or the upper strut bushing and a lower ball joint. Be sure
the wheels are pointed straight ahead when measuring this angle. [7]
Center Of Gravity: This is the imaginary point in a car where it would be exactly balanced if
lifted by a hoist. [7]
Ride Height: This is the height above the road that the car sits. [7]
Roll Center: This is an imaginary point about which the car rotates while in a turn. Each axle
has it's own roll center. The higher the roll center, the tipsier the car will feel. [7]
Sprung Weight: This is the weight of a car that is supported by the suspension. The engine,
body, interior, passengers, cargo, etc. are all sprung weight. [7]
Toe-In/Toe-Out: Toe is the dimensional difference of the distances between the front and rear
edges of the wheels on an axle. If the front edges are closer than the rear edges, there is toe-in.
Toe-out is when the rear edges are closer together. [7]
Unsprung Weight: This is the weight of a car that is not supported by suspension. Wheels, tires,
brakes, hubs, etc. are unsprung weight. Suspension components such as control arms, anti-roll
bars, shocks, and struts are a percentage sprung weight and a percentage unsprung weight. The
actual percentage depends on the application. [7]
Weight Distribution: This is the amount of weight on the front and rear axles expressed as
percentages. [7]

10
Bump Steer: This happens when the suspension compresses, causing the control arms and tie-
rods to move vertically. Because they differ in length and location, the result is the rim/tire being
steered without any movement of the steering wheel. Cars having control arms and tie-rods
parallel to the road will exhibit minimal bump steer. [7]
Counter Steer: If a car is torque steering to the left, turning the steering wheel to the right will
maintain a straight line of travel. A car that is over steering to the right can be brought back into
line by turning the steering wheel to the left. In both cases the driver is counter steering to correct
the car's direction of travel. [7]
Neutral Steer: This is the theoretic ideal steer characteristic when the front and rear tires lose
traction at the same time.[7]
Over steer: When the rear tires lose traction before the front tires, a car is over steering.
Recovery from an over steer situation must be quick since directional control can be lost. [7]

2.2.1.1 Survey
We went to Hyundai Service Center at Kupondole to study about the suspension used in different
model of Hyundai Company. In recent cars Hyundai uses Macpherson Strut suspension system
only. We then went to Sipradi Trading Pvt. Ltd at Soalteemode and learned about double
wishbone suspension system. Tata Automobiles uses double wishbone suspension system in
model like Strome, 207 DI, Xenon, Telecoline etc.

2.2.2 Design and Calculation

We made our final design in solid works and made calculation of the parameters required for
designing the double wishbone suspension system.

2.2.3 Analysis
We did structural analysis of suspension system and its component in ANSYS. We did stress
analysis and deflection analysis. We also performed experimental verification for deflection and
compared it with ANSYS data. We performed dynamic analysis in ADAMS and interpret the
result for change in camber angle, toe angle with wheel rate.

11
2.2.4 Report Submission
After completing design, calculation and analysis we prepared final report and submit it within
the deadline.

CHAPTER 3 DISCUSSION
3.1 Work Accomplished:
1. Literature survey was done referring various sites, books and journals which have been
mentioned in the reference section
2. From the result of literature survey, we have done the calculation for the different
parameters.
3. After calculation of parameters we made the final design in solidworks.
4. We performed stress and deflection analysis of different component of suspension system
in ANSYS
5. Experimental setup of double wishbone suspension system.
6. Dynamic analysis of suspension system.

3.2 Calculations
The weight of vehicle is: G  m  g Eqn 1 [8]
The load on front and rear axles are found by using equilibrium equation

Fig 3.2.1 Static Axle load on vehicle

12
 G L R
G FA
L
Eqn 2 [8]

G L F
G RA
L
Eqn 3 [8]

Case 1: Vehicle braking on level ground

Fig 3.2.2 Forces acting on a vehicle during braking

By using equilibrium of moments, the normal loads on front and rear axle are
G  LR  m  a  H
G FAdyn
 Eqn 4 [8]
L
G  LF  m  a  H
G RAdyn  L
Eqn 5 [8]

The transferred load on front axle is found to be

GT=GFAdyn-GFA
Eqn 6 [8]

13
Case 2: Vehicle at instant of cornering

Fig 3.2.3 Forces acting on vehicle during cornering

The centrifugal force due to velocity V and radius of bend R is given by

m V
2

F C
 Eqn 7 [8]
R
The cornering force produced by the tires, SL+SR results in lateral acceleration.

S  SL  SR  f s
G  f G
s LSdyn

 G RSdyn Eqn 8 [8]

Where fs is the friction coefficient between road and tire

The dynamic axle loads are found using moment equilibrium

G V B 
2
 B
  
GLSdyn B  g  R  2  sin   H  cos     cos   H  sin   Eqn 9 [8]
  2 

G V  
2
B  B
 
GLSdyn B  g  R  H  cos    sin     cos   H  sin   Eqn 10 [8]

2  2 

Transferred load from the left side to the right side of the vehicle while cornering;
G
G Gc RSdyn

2
Eqn 11 [8]

14
Case 3 Vehicle on a downhill grade

Fig 3.2.4 Forces acting on a vehicle on a downhill grade

The dynamic axle loads are found using moment equilibrium

G FAdyn

G
H  sin   LR  cos  Eqn 12 [8]
L

G 
G
  cos  H  sin  
L LF
RAdyn
Eqn 13 [8]

Transferred load from rear axle to front axle is

GT=GFAdyn-GFA Eqn 14 [8]

Kinetic Analysis

Fig 3.2.5 Kinetic analysis of double wishbone suspension system

The maximum force Gdyn and lateral force Sdyn at the centre of the front axle tyre contact for the
vehicle braking and cornering on a downhill grade are defined as shown in fig 3.2.5. The forces
Bx and By are on the joint B of the lower suspension and Ax and Ay are on the joint A of the upper
suspension control arm.
The loads on the A and B joints are found by summing moments about points“A” and “B”.

15
The moment equilibrium; ΣMB = 0;

A x
c A  a  b  G
y dyn
 b  S dyn .d Eqn 15 [8]

where AX=FA.cosδ and Ay = FA.sinδ , in which FA is the force acting on the link AE. The force
equilibriums in the direction x and y;

B X
 S dyn  A x
Eqn 16 [8]

B Gy dyn
 A y
Eqn 17 [8

Calculation of Control Arm Length

D=150mm
e =130 mm
H =350 mm
h = 100 mm
Z = 60 mm
Camber angle = 3 degree
Z 2  Z  e    53.2mm
Z 1  Z  d    52.14 mm
Y2  tan   Z 2  2.78mm
Y1  tan   Z 1  2.73mm
Z a
2
Z1
Y1   1
2  R1 R1
R1  173mm
Z b
2
Z2
Y2   2
2  R2 R21
R2  240 mm

Fig 3.2.6 Notation used in wishbone design [2]

16
Calculation of roll center

Fig 3.2.7 Determination of roll center

Assumptions:
1. Camber angle= -2 degree
2. Toe angle= 0 degree
3. Caster angle= 0 degree
4. Kingpin angle= 0 degree
5. Upper arm length= 173 mm
6. Lower arm length= 240 mm

Given:
1. Sprung weight= 240 kg
2. Unsprung weight= 50 kg
3. Weight bias= 40:60
4. Wheel travel= 10 cm
5. Wheel track= 92 cm
6. Wheel base= 153 cm

The angle of upper and lower control arms were determined by determining the roll center as
shown in figure 1.3. According to the literature the difference between c.g and roll center must
be in the range of 2.5 cm to 5 cm. The optimum difference was found to be 3.28 cm when upper

17
arm and lower arm were kept at an angle of 22 degree and 37 degree with respect to wheel base
respectively as shown in fig 3.2.7.
Calculation of spring
Material Selection: Carbon Steel
Allowable Shear Stress (Ԏ) = 420 MPA = 420 N/mm2
Modulus of Rigidity (G) = 80 KN/m2 = 80 * 103 N/mm2
Load (W) = 250 kg =2452.5 N
Spring Index (C) = 6
Deflection (δ) = 100 mm
Now, Wahl’s Stress Factor

4C  1 0.615
K   1.2525 Eqn 18 [6]
4C  4 C
8  K W  C
Since, 
 d2
Therefore, wire diameter (d) =10.5 mm
Mean diameter (D) = C*d = 63.5 mm
Outer Diameter (Do) = 74 mm
Now,
 G d
Number of turns (n) = =20 turns Eqn 19 [6]
8 W  C 3
Therefore, Total number of turns
For square and ground end,
n’=n+2=22 turns
Free length of spring = n’*d+ δ+0.15* δ = 347.5 mm
Freelength
Pitch of coil = =16.5 mm Eqn 20 [6]
n'1
W
Stiffness of spring (k) =  24.52 N/mm Eqn 21 [6]

Calculation of Damper
Motion Ratio = Wheel Displacement = 0.714 Eqn 22 [9]
Spring Displacement
Stiffness of tyre = 200 N/mm

18
Now, Kc = M.R2 * Ks =17.51 N/mm Eqn 23 [9]
Kc  Kt
Wheel Spring rate (Kw) = - 16.1 N/mm Eqn 24 [9]
Kc  Kt

1  Kw 
Natural Frequency =    1.27 Hz
2  Ms  Eqn 25 [9]
Selection of damper:
Damping ratio (ζ) = 0.5 to 0.7 for performance car
Therefore, taking average of the above value we select ζ = 0.6
Mdamper = Msprung/ M.R = 350 kg Eqn 26 [9]
Rebound damping factor (Cr) = 4    Mdamper*f =5585.75 kg/s Eqn 27 [9]
Damping factor (C) = ζ*Cr = 3351.45 kg/s Eqn 28 [9]

Quarter Car Model Analysis

F1  K s   x 2  x1 
. .
F2  C s  ( x 2  x1 )
F3  K t  ( x1  xu )
..

M u  x1  K s  x 2  x1   C s  ( x 2  x1 )  K w  x1  xu 
. .

..

M s  x 2   K s  x 2  x1   C s  ( x 2  x1 )
. .

K  K K  K  Ks K s  4 K s2 
 21, 2  0.5   t s
 s   t   
 M u Ms  Mu M s  M u  M s 
Eqn 29

Therefore, Frequency of sprung mass (f1) = 1.85 Hz Fig 3.2.8 Quarter Car Model

Frequency of unsprung mass (f2) = -13.31 Hz

19
3.4 Final Design

Fig 3.4.1 Orthographic View of Double Wishbone Suspension System

The final design of the suspension system was designed in solidworks as shown in fig 3.4.1. The
assembly of components like control arms, springs, damper and wheels were done to create a
final design.

3.5 Stress and Deflection Analysis in ANSYS

3.5.1 Analysis of spring
Spring is analyzed in Ansys analysis software so as to determine the actual maximum deflection
of spring corresponding to the maximum spring force as shown in 3.5.1. Also, the maximum
stress value corresponding to the maximum spring force is determined

20
 Fig 3.5.1 Analysis of Spring
.In spring analysis, one end of spring is fixed and vertical load has been applied on the other side.
Spring Analysis Results
Parameters Value
Maximum Force 2450 N
Maximum Deflection 60 mm
Maximum Stress 995.14 MPa

3.5.2 Analysis of Control Arm:

After modelling the wishbones in solidworks software, these models were imported into Ansys
Analysis Software.
Various boundary conditions and load cases were applied for determining the maximum stress
and maximum deflection for wishbone.
Input parameters are as follows,

21
Material: AISI 1040
Vertical Load: 2450 N

Fig 3.5.2 Analysis of Control Arm

Results of Analysis of lower wishbone in Ansys is as shown in fig 3.5.2
Maximum Stress 515.82 MPa
Maximum Deflection 60 mm
Allowable Stress 790 MPa
Since, Maximum stress induced in wishbone is less compared to allowable stress, hence the
wishbone is safe.

3.6 Experimental Verification

After the completion of Double wishbone suspension system in ANSYS we performed
experimental verification. The experiment was a roller test experiment where suspension was
fitted with a wheel and another wheel from bottom is coupled with motor as shown in fig 3.6.1.
The bottom wheel rotates the top wheel and the bump is introduced in the middle of two wheels.
The suspension system shows certain deflection for every bump respective of their size. Those
deflections were noted and compared with the data obtained from ANSYS.

22
 Fig 3.6.1 Orthographic View of Test Rig for Double Wishbone Suspension System
The data obtained from both analyses is shown below in the table.
S.N Weight (Kg) Data from Experiment (mm) Data from ANSYS (mm) Error (%)
1 20 9 9.53 5.88
2 30 13 14.298 9.98
3 50 21 23.83 13.47
4 100 - 47.659
5 200 - 132.27
6 240 - 107.98

The error or deviation in the data is due to not considering tire stiffness and damper damping
factor while doing analysis in Ansys Software.

3.7 Dynamic Simulation in ADAMS

After completion of experiment we performed dynamic simulation on software named ADAMS.
We performed a parallel wheel travel analysis on the suspension and steering assembly, and the
results were plotted using the following steps,
1) Importing the suspension and steering subsystems and assembly

23
 2) Defining Vehicle Parameters
3) Performing the Analysis
4) Animating the Results
5) Plotting the Results

Step 1: Importing the suspension and steering subsystems and assembly:

You create the front suspension subsystem based on a double-wishbone design stored in the
standard template named double_wishbone.tpl, and then save it. After you create the subsystem,
you save it in an ADAMS/Car database. The suspension can be further customized like making
the wishbone equal and parallel, unequal and parallel and unequal and not parallel using the hard
points.

Fig 3.7.1: Setup of Double Wishbone Suspension for Dynamic Analysis in ADAMS
Hardpoint Modification Table
Location Code word Loc X Loc Y Loc Z Remarks
Drive shaft hpl_drive_shaft_inr 0.0 -200.0 225.0 none
Lower control arm front hpl_lca_front -200.0 -400.0 150.0 none
Lower control arm outer hpl_lca_outer 0.0 -750.0 100.0 none

24
Lower control arm rear hpl_lca_rear 200.0 -450.0 155.0 none
Lower strut mount hpl_-lwr_strut_mount 0.0 -600.0 150.0 none
Front subframe hpl_subframe_front -400.0 -450.0 150.0 none
Rear subframe hpl_subframe_rear 400.0 -450.0 150.0 none
Inner tierod hpl_tierod_inner 200.0 -400.0 300.0 none
Outer tierod hpl_tierod_outer 150.0 -750.0 300.0 none
Top mount hpl_top_mount 40.0 -500.0 650.0 none
Upper control arm front hpl_uca_front -150.0 -400.0 530.0 none
Upper control arm outer hpl_uca_outer 40.0 -675.0 525.0 none
Upper control arm rear hpl_uca_rear 200.0 -400.0 530.0 None
Wheel center hpl_wheel_center 0.0 -800.0 300.0 None

Step II Defining Vehicle Parameters: Before performing a suspension analysis, we must

specify several parameters about the vehicle in which you intend to use the suspension and
steering subsystems. These parameters include the vehicle’s wheel base and sprung mass,
whether or not the suspension is front- or rear-wheel drive, and the braking ratio. For this
analysis, you enter the parameters to indicate front-wheel drive and a brake ratio of 64% front
and 36% rear. For the analysis we put the following value

S.N Parameters Value

1 Tire Stiffness 200 N/mm
2 Wheel Mass 1 kg
3 Sprung Mass 250 kg
4 CG height 300 mm
5 Wheel Base 1530 mm
6 Camber angle -1.0 degree
7 Drive shaft offset 75 mm
8 Toe angle 0.0 Degree

Step III Performing the Analysis: Now that we have defined the vehicle parameters, we run the
parallel wheel travel analysis. During the analysis, the test rig applies forces or displacements, or
both, to the assembly, as defined in a loadcase file. For this analysis, ADAMS/Car generates a
temporary loadcase file based on the inputs you specify. This parallel wheel travel analysis
moves the wheel centers from -100 mm to +100 mm relative to their input position (wheel
center), while holding the steering fixed. During the wheel motion, ADAMS/Car calculates

25
many suspension characteristics, such as camber and toe angle, wheel rate, and roll center height.
The analysis is completed in 50 steps.
Step IV To animate the results: For animating the result first of all we select animation control
from review menu then select play tool. After selecting play tool ADAMS/Car animates the
motion of the suspension analysis. During analysis the suspension moves up (bump) to down
(rebound) about wheel center and steering does not rotate.
Step V Plotting the Results: In this section, we create several plots from the parallel wheel
travel analysis results. In a plot configuration file, we have all the information that ADAMS/Car
needs to create the plots. The plot configuration file not only specifies which plots ADAMS/Car
should create, but also how the plots should look, including their horizontal and vertical units,
and colors. Storing plotting information in a plot configuration file lets you quickly regenerate
plots after each analysis.

3.7.1 Analysis of Change in Camber Angle with Wheel Travel

Fig 3.7.2 Graph Showing Change in Camber Angle with Wheel Travel
In the above graph, the wheel travel is plotted in X-axis and camber angle is plotted in Y-axis. At
the initial position or wheel center (Wheel travel = 0.0) the camber angle is found to be -1. As
the wheel travel upward i.e. as wheel goes on a bump, the camber angle decreases and reaches
upto -3 degrees for the bump of height 100 mm. Similarly when wheel travels downward i.e as
wheel goes on rebound, the camber angle increases and reaches up to 0.5 degrees for 100 mm
rebound height. The negative camber is obtained most of the time as it gives higher stability,
wider distance and makes handling of vehicle easier. Most of the off road and racing vehicle uses
negative camber angle whereas positive camber is used in simple vehicles.

26
3.7.2 Analysis of Change in Toe Angle with Wheel Travel

Fig 3.7.3 Graph Showing Change in Toe Angle with Wheel Travel

The above graph shows the change in toe angle as the wheel travel from rebound to bump of
-100 mm to 100 mm. At the initial position or wheel center (Wheel travel = 0.0) the toe angle is
found to be 0.0. As the wheel travel upward i.e. as wheel goes on a bump, the toe angle
decreases and reaches upto -2.9 degrees for the bump of height 100 mm. Similarly when wheel
travels downward i.e as wheel goes on rebound, the toe angle increases and reaches up to 5.85
degrees for 100 mm rebound height. Normally toe angle is kept at zero degree but it is suitable to
keep slight toe in for rear wheel drive vehicles and toe out for front wheel drive vehicles.

3.7.3 Analysis of Change in Spring Stiffness with Wheel Travel

Fig 3.7.4 Graph Showing Change in Spring Stiffness with Wheel Travel
At initial position when the vehicle is in static position (wheel travel=0.0) the weight of the
vehicle compresses the spring to some extent and the value found for the stiffness is nearly 14

27
N/mm. During rebound the spring elongates little more than static condition and reaches the
value of around 16.5 N/mm. The spring elongates due to the depth of 100 mm it has to travel
downwards. But during the bump of 100 mm the stiffness curve shows steady nature upto 33 mm
bump and the graph shows the nature of ‘S’ curve. This is due to the reason that small bump of
height of 30-35 mm has almost no effect on spring stiffness as tyre stiffness is enough for such
small bump but as the bump height rises the stiffness decreases to some extent and increases
exponentially and when spring reaches its maximum, it decreases somewhat and becomes steady.
As our designed spring stiffness lies in the region in this graph so it is safe design.

28
 CHAPTER 4 GHANTT CHART

Dec Jan Feb March April May June July Aug Sep

Proposal Defense

Literature Review

Survey of Suspension
System in Existing
Vehicle of Nepal

Design of Double
Wishbone Suspension

Calculation of the
Parameters

Analysis of stress and

deflection in Ansys

Dynamic analysis in
ADAMS

Fabrication of
experimental setup,
testing and collecting
data
Final Report Submission

Work Accomplished

Work Remaining

29
 CHAPTER 5 BUDGET

S.N Particulars Price (Rs)

1 Spring and damper 4700

2 Survey 2000

3 Nut and bolt 160

4 Wiper motor 4500

5 Wheels with rim 1600

Total 12960

30
 CHAPTER 6 LIMITATION
 For experimental setup, test can be performed for more weight and data can be collected
for optimum value (240 kg).
 During experimental setup only cylindrical obstacles were used. So other shapes
obstacles can also be used.

CHAPTER 7 RECOMMENDATION
 Double wishbone suspension system is passive suspension system. Study must be made
about active suspension system also along with passive suspension system.
 The suspension system can be tested by installing it in Go-kart and data can be collected.

31
 CHAPTER 8 CONCLUSION

In this project, we designed double wishbone front suspension system for Go-kart. Calculated for
length of control arms, angle of inclination, spring stiffness, number of turns and design of spring
with damping factor of the damper. Quarter car model analysis of two degree of freedom to
obtain natural frequency of the sprung and un-sprung mass. The design of control arms with
spring and damper was done in Solid work design software. Stress and deflection analysis were
performed in ANSYS. An experimental test rig was fabricated for testing and verification of
those data. Finally dynamic analysis was performed on ADAMS software to study variations in
camber angle, toe angle and spring stiffness with wheel rate.

32
 REFRENCES
[1] Suspension System. Assesed on 30 Novemeber 2015
<http://thekneeslider.com/archives/2009/01/29/wesll-4-wheel-leaning-suspension-system/ >
[2] Calculation of double wishbone suspension system. Assesed on 1 December 2015
<https://www.quora.com/How-do-I-calculate-weight-of-double-wishbone-suspension >
[3]Double wishbone suspension system. Assesed on 30 Novemeber 2015
<https://en.wikipedia.org/wiki/Double_wishbone_suspension>
[4] V.B. Bhandari, “Machine Design”, , McGraw Hill, 2012.
[5] Ahmad Keshavarzi, “Optimization of Double Wishbone System with Variable Camber angle
by Hydraulic Mechanism”, World Academy of Science, Engineering and Technology, 2010.
[6] R.K Khurmi, “Machine Design”, S. Chand Publication.
[7] Double wishbone suspension system parameters and types of suspension. Assesed on 20
November 2015 <Double wishbone suspension/Automobile Ride, Handling and Suspension
Design.html>
[8] Milliken & Milliken, “Race car vehicle dynamics”, Society of Automotive Engineers, Inc.
1995
[9] Optimizing the Kappa Platform Suspension. Assessed on 15 January 2016.
<file:///D:/project/4th%20year/4th%20year%20project/damper.html>
[10] Attia, H.A., 2002. “Dynamic Modelling of the Double Wishbone Motor Vehicle Suspension
System”, European Journal of Mechanics A/Solids, Volume 21, pp.167174.
[11] Gillespie, T.D., 1992. “Suspensions”, in Fundamentals of Vehicle Dynamics, (Society of
Automotive Engineers, USA), pp.97-117 and pp.237-247.
[12] Remling, J., 1983. “Independent Front Suspension Systems”, in Steering and Suspension,
(Wiley, NewYork), pp.189-198.

33
ANNEX

34
 Experimental Verification

Fig: Prototype of Double Wishbone Suspension System

Fig: Roller Test of Double Wishbone Suspension System

35
 Some of the designs of double wishbone suspension system available in Nepal

Fig: Double Wishbone Suspension with torsion bar but without coil spring between control arms

Found in vehicles like Tata Xenon, Tata Telecoline, Tata Strome

Fig: Double Wishbone Suspension with coil spring between control arms but without torsion bar

Found in Tata 207 D.I, Tata Sumo Gold

36
Stress Analysis in ANSYS for different Load

For 20 kg Load For 30 kg Load

For 50 kg For 100 kg

For 240 kg

37
Dynamic Simulation in ADAMS

Fig: Hardpoint Modification Table

Fig: Defining Vehicle Parameter Table

38
Fig: Suspension Analysis Table

Fig: Animation Control Table

39