1 Table of content SUSPENSION Page 1. List of symbols and abbreviation 2. Important Terminologies 3. Overview of Suspension Systems a. Introduction b. Types of Suspension systems i. Dependent Suspension Systems ii. Independent Suspension System 4. Front Suspension for BAJA vehicle 1. Selection of suspension system 2. Suspension Geometry 3. Simulation of Suspension 5. Design of front suspension components 1. Wishbone Design 1.Geometry 2.Modeling 3.Analysis 2. Front Upright Design 1.Designing Parameters 2.Material Selection 3.Modeling 4.Loading Scenarios 5.Analysis 3. Front Wheel Hub design
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Design,Analysis & Fabrication of suspension of all terrain vehicle
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1
Table of content
SUSPENSION
Page
1. List of symbols and abbreviation
2. Important Terminologies
3. Overview of Suspension Systems
a. Introduction
b. Types of Suspension systems
i. Dependent Suspension Systems
ii. Independent Suspension System
4. Front Suspension for BAJA vehicle
1. Selection of suspension system
2. Suspension Geometry
3. Simulation of Suspension
5. Design of front suspension components
1. Wishbone Design
1.Geometry
2.Modeling
3.Analysis
2. Front Upright Design
1.Designing Parameters
2.Material Selection
3.Modeling
4.Loading Scenarios
5.Analysis
3. Front Wheel Hub design
2
1.Introduction
2.Wheel bearing Selection
3.Designing Parameters & Considerations
4.Modeling
5.Loading Conditions
6.Analysis
4. Stub Axle design
1.Designing Parameters & Considerations
2.Modeling
3.Analysis & Material Selection
6. Rear Suspension for BAJA vehicle
1. Selection of suspension system
2. Suspension Geometry
3. Simulation
7. Design of rear suspension components
1. Trailing Arm & Rear upright design
1.Geometry
2.Designing parameters & Modeling
3.Analysis
2. Rear Wheel Hub design
1.Introduction
2.Wheel bearing Selection
3.Designing Parameters & Considerations
4.Modeling
5.Loading Scenarios
6.Analysis
8. Fabrication of Front Suspension Components
3
1. Wishbones
2. Front Uprights
3. Front Wheel Hubs
4. Stub axles
5. Trailing arms
6. Rear uprights
7. Rear wheel hubs
STEERING
9. Introduction to Steering System
10. Types of Steering Gearboxes
11. Design of Steering System
1. Steering geometry
2. Collapsible Steering Assembly
12. Fabrication of Steering System
13. Summary
14. Bibliography
4
5
1. Units Used
Steering Effort Fst N
Braking torque Tb N-m
Radius of brake disk rb M
Cornering Force Fcorner N
2. Important terminologies
1. Upright: the component that holds together the suspension control arms to the tyre
of the vehicle, the upright accommodates the upper and lower ball joints. It allows
for the pivotion of the tyre to perform steering action, the upright only pivots
about its axis, and moves in a vertical path tracing the tyre.
2. Hubs: The hubs are the rotating components that allow for the rotation of the tyre,
they are placed on the stub axle or live axle, usually supported by wheel bearings,
these hubs carry the rim along with the tyre and in most cases the brake rotors.
3. Stub axle: it is an integral part of the upright, it is an extension from the knuckle
that will eventually be used to carry the hub, the stub axle is a replacement for a
live-axle.
4. Mounting tabs: These are plates that are welded on to the frame, with the required
provision for bolting the suspension components such as wishbones and the shock
absorber. They are positioned according to the required suspension geometry.
5. Shock absorbers: The primary damping components used in a vehicle to allow the
dissipation of energy absorbed by the tires during motion, which may be bump
force or any other form of shock loading. Shocks control spring motion, that is,
they slow down and reduce the magnitude of the spring’s oscillation. The process
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is known as damping. In technical terms, a shock controls the frequency and
amplitude of the suspension's oscillation. In layman’s terms, a shock controls
how fast and how much the suspension compresses and rebounds.
6. Wheel rate: is the actual rate of a spring acting at the tire contact patch. This value
is measured in lbs/inch or N/mm.
7. Suspension frequency: refers to the number of oscillations or "cycles" of the
suspension over a fixed time period when a load is applied to the vehicle.
8. Ductility: In materials science, ductility is a solid material's ability to deform
under tensile stress; this is often characterized by the material's ability to be
stretched into a wire.
9. Finite Element Methods: The finite element method (FEM) is a numerical
technique for finding approximate solutions to boundary value problems for
differential equations. It uses vibrational methods to minimize an error function
and produce a stable solution. Analogous to the idea that connecting many tiny
straight lines can approximate a larger circle, FEM encompasses all the methods
for connecting many simple element equations over many small sub domains,
named finite elements, to approximate a more complex equation over a
larger domain.
10. Hardness: Hardness is a measure of how resistant solid matter is to various kinds
of permanent shape change when a force is applied. Hardness is dependent on
ductility, elastic stiffness, plasticity, strain, strength, toughness, and viscosity.
11. Hardness number: A number representing the relative hardness of a mineral,
metal, or other material as determined by any of more than 30 different hardness
tests such as Brinell hardness number, Rockwell hardness number.
12. Endurance Limit: The maximum stress that a material can withstand for an
infinitely large number of fatigue cycles; maximum cyclic stress level a metal can
withstand without fatigue failure. See also fatigue strength.
13. Bore: In terms of machinery bore is a process of enlarging a hole to a precise
diameter with a cutting tool within the hole by rotating either the tool or the work
piece.
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3. Overview of Suspension Systems
3. 1 Introduction
Vehicle dynamics is concerned with two aspects of the behavior of the machine. The first
is isolation and the second is control. It is a study of the behavior of the vehicle in a
dynamic state.
Chart reference: Multi-body systems approach to Vehicle dynamics by Mike Blundell
and Damian harty.
The most important constituents of vehicle dynamics are given below with a brief
explanation.
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Isolation: The process of separating the driver from disturbances occurring as a result of
vehicle operation. It is evident that there are two major types of disturbances, external
which are due to factors outside the vehicle system and mostly beyond the control of the
driver such as aerodynamics, which is weight gain, lift etc, road factors speak about the
road nature, roughness etc, internal factors constitute the vibrations due to engine and
transmission components etc. The behavior of the vehicle in response to road undulations
is referred to as ‘ride’ and could conceivably be grouped with refinement, as the ride of
the vehicle can be tuned to achieve a desired effect.
Control: response of the vehicle to driver demands. The driver continuously varies both
path curvature and speed, subject to the limits of the vehicle capabilities, in order to
follow an arbitrary course. Speed variation is governed by vehicle mass and tractive
power availability at all speeds on different types of terrain.
The suspension of an automobile is a system that consists of mechanical components that
are designed to fit in a preferred geometry, being able to handle the effects of road
irregularities or dynamic characteristics. The primal objectives of the system are
● To be able to locate all the four wheels of the vehicle
● To be able to maintain the required ride height (ground clearance)
● To provide a stable and comfortable ride and handling
● To provide road contact, even in inhospitable terrain.
This thesis is about the designing process and fabrication methods involved for an all-
terrain vehicle that has been designed and manufactured to compete in the BAJA
SAEINDIA 2014 competition.
Designing suspension system for any car requires technical knowledge in several
disciplines such as design of machine members, kinematics and geometry. The geometry
essentially means the board subject of how the unsprung mass of the vehicle is connected
to the sprung mass.
These connections or links dictate the path of motion and also control the forces that are
transmitted between them.
A suspension geometry must be designed to meet the requirements or ideals of the
vehicle to be built, a lot of factors must be taken into consideration such as the ride
height, the travel, the spring rates etc.
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Usually the chassis or the frame of the vehicle needs to be modified or designed
according to the suspension system, hence one setup cannot be used elsewhere. There is
no single best geometry.
Apart from the basic objectives a few vehicle specific suspension objectives have been
made to design the suspension for the all-terrain BAJA vehicle, they are:
● Large suspension travel
● Adjustability, to tune the suspension.
● Good ground clearance.
● Simple and lightweight construction.
Every suspension setup is an assemblage of control arms, shock absorbers, uprights, axles
and tyres which are laid out in a way to do their part, quietly in the background as the
vehicle is put to all types of loads.
The design of this system gets complex in way because, while being restricted (
controlled ) in their motion path by the control arms, the wheel will have camber, caster
and toe change. Therefore sometimes just links or control arms are not sufficient to
provide a good suspension characteristic. In such a scenario various components such as
toe link, camber link, shock absorber mounting become extremely important.
The suspension design has been done in a phase-wise manner to ease up the task and to
provide better flexibility of the results and also to allow for modification of the design if
there may be a need.
Phase 1: Determination of geometry to be used and the type of setup to be used
Considerations made during Phase 1 were:
● Independent nature.
● Comply with rule book track width of 64”.
● Smaller packaging.
● Fabrication limitations.
● Weight reduction.
Phase 2: Determination of spring and damper system.
Considerations made during Phase 2 were:
● Motion Ratio
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● Installation Ratio
● Spring rates
● Damping characteristics
● Weight of the Spring-Damper system
● Initial compression of spring
● Ride frequency
In this regard, various other parameters which are explained in later chapters have also
been taken care off, some of them being, position of roll centre, minimization of scrub
radius, anti-squat, anti-dive. To avoid rollover the vehicles centre of gravity has been put
as low as possible, by doing this we have restricted the movement of the centre of gravity
to an extent.
After establishing the design parameters the team has done different types of market
surveys locally and on the internet to find components that are well suited for the
purpose, the emphasis was on manufacturing most of the components to avoid
outsourcing, although expensive, it would serve all the requirements as well as have clean
engineering ethics rather than modifying an existing setup to suit ours.
Objective of steering system:
● Allow for sharp steering angles.
● Allow for extended suspension travel.
● Provide straight-line stability.
● Minimize bump steer.
● Have good steering return ability
● Be precise and compact.
3. 2 Types of Suspensions
3. 2.1 Dependent suspension system
This type of system normally has an axle which holds both the wheels parallel to each
other and perpendicular to the axle. When certain amount of changes occur in one wheel
due to some external causes such as bump then the same amount of changes occur in the
other wheel in the same manner.
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Dependent systems may be differentiated by the system of linkages used to locate them,
both longitudinally and transversely. Often both functions are combined in a set of
linkages
Types of dependent suspension system are:
1. Solid axle
1. Solid axle over leaf spring
2. Solid axle over coil spring
2. De-Dion
3. Torsion beam suspension
1. SOLID AXLE:
A solid axle also known as beam axle is a dependent type of suspension system in which
a single beam or a solid axle or a shaft is connected to a pair of wheels in a lateral
manner. With a beam axle the camber angle between the wheels is the same no matter
where it is in the travel of the suspension
1. Solid over leaf spring:
A Leaf Spring is a simple form of spring commonly used in heavy commercial vehicles
and four wheeled drive vehicles. The leaf spring is arc-shaped and works by suspending
the chassis of the vehicle to avoid contact with wheels. In this type of suspension system
the drive axle is clamped to the leaf spring and the shock absorber normally bolted
directly to the axle. The end of the leaf springs is attached directly to the chassis, as are
the tops of the shock absorbers. A leaf spring takes the form of a slender arc-shaped
length of spring steel of rectangular cross-section. 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
12
and results in restriction in the motion of the suspension. It 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.
There are several types of leaf springs available based on number of leafs stacked on each
other and the geometry of the spring few of which are mono-leaf spring, elliptical leaf
spring, semi-elliptical leaf spring and more. In the modern implementation of parabolic
leaf spring. The design is characterized by fewer leaves whose thickness varies from
center to ends following a parabolic curve. In this design, inter-leaf friction is unwanted,
and therefore there is only contact between the springs at the ends and at the center where
the axle is connected. The main advantage of parabolic springs is their greater flexibility,
which translates into vehicle ride quality. The basic advantage of this type of setup is that
the leaf spring acts as a linkage for holding the axle in position and thus separate linkage
are not necessary. It makes the construction of the suspension simple and strong. But as
the positioning of the axle is carried out by the leaf springs so it makes it disadvantageous
to use soft springs i.e. a spring with low spring constant. This type of suspension does not
provide good riding comfort. The inter-leaf friction between the leaf springs affects the
riding comfort. Acceleration and braking torque cause wind-up and vibration. Also wind-
up causes rear-end squat and nose-diving. The main drawback with this arrangement is
the lack of lateral location for the axle, meaning it has a lot of side-to-side slop in it.
2. Solid over coil spring:
This is a variation and update on the system described above. The basic idea is the same,
but the leaf springs have been removed in favor of either ‘coil-over-oil’ spring or shock
combos. A coil spring, also known as a helical spring, is a mechanical device, which is
typically used to store energy due to resilience and subsequently release it, to absorb
shock, or to maintain a force between contacting surfaces. They are made of an elastic
material formed into the shape of a helix which returns to its natural length when
unloaded. Because the leaf springs have been removed, the axle now needs to have lateral
support from a pair control arms. The front ends of these are attached to the chassis, the
rear ends to the axle. The variation shown here is more compact than the coil-over-oil
type, and it means we can have smaller or shorter springs. This in turn allows the system
to fit in a smaller area under the car. In addition to cylindrical springs, with which the line
of force moves along the damper axis, lateral forces compensating side load springs are
produced that have a line running diagonally to the spring center line. Use of side load
springs leads to increase in driving comfort, driving safety and optimum use of space due
to extensive compensation of lateral forces. One of the main advantage of coil spring is
that it gives a better road handling and better braking.
13
The principal advantage of the beam axle is its simplicity. This simplicity makes it very
space-efficient and relatively cheap to manufacture. Beam axles are also ideal for
carrying heavy or varying loads because they do not ever exhibit any camber change as
the suspension travels. They are nearly universally used in heavy-duty trucks and most
light and medium duty pickup trucks, SUV’s, and vans also use a beam axle, at least in
the rear. The drawbacks are that it does not allow each wheel to move independently in
response to bumps, and the mass of the beam is part of the unsprung weight of the
vehicle, which can further reduce ride quality. Also the cornering ability is typically
worse than other suspension designs because the wheels have zero camber angle gain
during body roll.
2. DE-DION:
A de Dion tube is an automobile suspension technology. It is a sophisticated form of non-
independent suspension and is a considerable improvement. A de-Dion suspension
uses universal joints at both the wheel hubs and differential, and uses a solid tubular
beam to hold the opposite wheels in parallel. Unlike an anti-roll bar, a de Dion tube is not
directly connected to the chassis nor is it intended to flex. In suspension geometry it is
close to the trailing beam suspension seen on many front wheel drive cars, but without
the torsional flexibility of that suspension. With this system, the wheels are
interconnected by a de Dion Tube, which is essentially a laterally-telescoping part of the
suspension designed to allow the wheel track to vary during suspension movement. This
is necessary because the wheels are always kept parallel to each other, and thus
perpendicular to the road surface regardless of what the car body is doing. This setup
means that when the wheels rebound, there is also no camber change which is great for
traction, and that's the first advantage of a de Dion Tube. The second advantage is that it
contributes to reduced unsprung weight in the vehicle because the transfer case /
differential is attached to the chassis of the car rather than the suspension itself.
Naturally, the advantages are equaled by disadvantages, and in the case of de Dion
systems, the disadvantages would seem to win out. First off, it needs two CV joints per
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axle instead of only one. That adds complexity and weight. Well one of the advantages of
not having the differential as part of the suspension is a reduction in weight, so adding
more weight back into the system to compensate for the design is a definite disadvantage.
Second, the brakes are mounted inboard with the calipers attached to the transfer case,
which means to change a brake disc, you need to dismantle the entire suspension system
to get the driveshaft out. (Working on the brake calipers is no walk in the park either.)
Finally, de Dion units can be used with a leaf-spring or coil-spring arrangement. With
coil spring it needs extra lateral location links, such as a Panhard rod, wishbones or
trailing links. Again - more weight and complexity.
de Dion suspension was mostly used from the mid 60's to the late 70's and could be found
on some Rovers. More recently deDion suspension has had somewhat of a renaissance in
the specialist sports car and these all uniformly now use outboard brake setups for ease-
of-use, and a non-telescoping tube, usually with trailing links and an A-bar for lateral
location (rather than a Watts linkage or Panhard rod). Unlike most fully independent
suspension there are no camber changes on axle loading and unloading (or rebound).
Fixing the camber of both wheels at 0° assists in obtaining good traction from wide tires
and also tends to reduce wheel hop under high power operations compared to an
independent suspension. And with this setup the designer has free will to opt for various
shock absorbers. The most disadvantageous point is the manufacturing cost of this setup
which is high when compared to other types of setup.
3. TORSION BEAM SUSPENSION:
A torsion bar suspension, also known as a torsion spring suspension or torsion beam
suspension, is a general term for any vehicle suspension that uses a torsion bar as its main
weight bearing spring. One end of a long metal bar is attached firmly to the vehicle
chassis; the opposite end terminates in a lever, the torsion key, and mounted
perpendicular to the bar that is attached to a suspension arm, a spindle, or the axle.
Vertical motion of the wheel causes the bar to twist around its axis and is resisted by the
bar's torsion resistance. The effective spring rate of the bar is determined by its length,
cross section, shape, material, and manufacturing process. The ride height may be
adjusted by turning the adjuster bolts on the stock torsion key, rotating the stock key too
far can bend the adjusting bolt and more importantly place the shock piston outside its
standard travel. Over-rotating the torsion bars can also cause the suspension to hit the
module”, this software allows us to accurately model the suspension system and run tests
for scenarios such as 2d bump, 2d steer,3d bump, 3d steer, 3d roll
The figure below shows the front suspension model in LOTUS SHARK.
26
Fid. 3 dimensional bump, note there is no excessive camber change or toe change.
Fig: 3 dimensional steer.
Fig: 3 dimensional roll.
Graph 1:Camber (deg) vs Roll angle ( deg)
27
the vehicle gains 9 degrees of
negative camber at full roll
Graph above shows the camber change (deg) with respect to steer travel.
The following simulations have been done to ensure the suspension behaves in a planned
manner to determine the ride and handling, a lot of testing is required on various terrains,
tuning the suspension becomes of paramount importance to achieve the best required
control and handling of the vehicle.
28
5. Design of Front Suspension Components
5. 1 Design of Control arms
5. 1.2 Modeling
The following dimensions for the modeling of upper and lower control arms were
provided by suspension analyzer. The length of them were based on front nose
dimensions, track width and various other performance significant parameters. The model
ought to reflect the lengths mentioned above.
29
Various shapes for control arms were considered initially. Since the control arm is the
link between tire and body of the vehicle, it need to be stiff and strong to support also
control the tire motion.
There were a lot of variant in design of control arms few are:
30
After studying various design, the finalized design was make it out of a single tube by
bending into a parabola.
31
Reasons for selection of the design
1. Easy to fabricate
2. Consumes less time in production
3. Sophisticated jig not required
4. Less number of welds hence low heat effected areas.
5. Easy correction in design by opening and closing of bend.
The upper wishbone according to geometry was prepared as following
32
The upper stock connected the upright. It will be inserted by stock steering rod ends of a
commercial vehicle.
The model of lower wishbone is different from the mentioned as it needs to mount the
shock absorber. It is done by providing a lateral tube with the mount. Since the design is
for front suspension lower one mounts the shock absorber where as in case of it being
used in rear suspension mounts are given on upper wishbone due to obvious reasons
33
The tubing for control arm was chosen to be AISI 4130 steel with outer diameter 25mm
and 3.5 mm wall thickness. The tabs for shock mount were from 4mm thick mild steel
sheet.
5. 1.3 Analysis
5. 2 Design of Upright
5. 2.1 Design parameters
Before designing of any components there are various parameter that are to be included
in it. Irrespective of other details the main design parameters determine mostly the
performance, adaptability with the environment, mates with the sub-component in an
assembly, space occupancy etc. They are Special consideration and often are the
constrains which are to be met.
The parameters that molded the design of the upright were:
1. Include castor angle of 6o along the vertical axes of upright.
2. Project the brake caliper mounts at one side of upright.
3. Provide sufficient thickness to brake caliper mounts to endure sudden torque from
the disk rotors.
4. Check alignment of the brake caliper mount on both of the uprights i.e. Left and
Right uprights. Since the brake caliper doesn’t have plane of symmetry along its
center, brake mounts will have different spacial arrangements along the side of
both uprights.
5. Steering arm mount be on the opposite side that of the brake caliper mount of the
respective uprights.
6. Twin steering arm mount be provided to facilitate double shear for steering arm
bolts.
7. Dual bolt holes will be provided to counteract the moment in the steering arm.
8. Have sufficient fillet radius throughout the design to minimize notch sensitivity.
9. Length of upright will be taken as per the suspension design that gave the
optimum results.
10. Upper and lower wishbone mounting will be dependent on the kingpin
inclination obtained from the suspension geometry
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11. Bore be provided to accommodate the stub axle.
12. Press fitting tolerance to be provided in the central bore diameter to press fit the
stub axle.
13. Sufficient wall thickness to make the component rigid, unsusceptible to external
moment.
14. Design optimization will be done after the component is analyzed for various
loading cases to relief weight.
5. 2 .2 Material Selection
As it was mentioned above in design parameters, weight consideration was the main
objective that dictated the choice of material for the suspension components. Also as
mentioned earlier the preference of low unsprung mass in the vehicular system, it was
necessary to opt the material which could bear the forces induced during the motion as
well be light.
Market survey revealed the local availability of the following suitable material candidate
for the component.
1. Cast Ductile Iron
2. Titanium Alloy Ti 6Al-4V
3. Aluminum Alloy AA 6351 T-6
4. Alloy Steel AISI 4140
Specific Strength comparison
It is the strength to weight ratio of the material. Strength here can be tensile or yield
strength. A higher ratio dictates the material has appreciable strength compared to its
weight.
35
The above graphs shows the usage of Titanium alloy to be most apt choice for the
component. Nevertheless other factors also have be considered before material selection
is to be made.
Cost per unit yield strength
Comparison also has to be made with the cost of the material. The cost comparison (C) is
made by following equation;
𝐶 = 𝑐𝑚 ∗𝜌
𝜎
Whereas,
Cm ꞊ Cost of the material per unit mass (Rs/Kg)
𝜌 ꞊ Density of the material (Kg / m3)
σ ꞊ Yield strength of the material (N/m2)
77674.6
186816.4
101851
58598.7
0
20000
40000
60000
80000
100000
120000
140000
160000
180000
200000
Ductile Iron Ti6Al4V AA 6351 T6 AISI 4140
SPEC
IFIC
STR
ENG
HT
N M
/ K
G
36
Ductility
Defined as the percentage elongation under stress before the material ruptures. This
property is important as it gives a clear warning before the metal breaks down.
The above graph shows the cost value for unit strength of the material. It can be seen that
Titanium alloy have the highest value for its strength whereas ductile alloy have the least.
Interpretations from above three detrimental property graphs;
1. Titanium and aluminum seems to be viable candidates where the strength to
weigh ratios are to be considered.
0
2000
4000
6000
8000
10000
12000
14000
16000
18000
Ductile Iron Ti6Al4V AA 6351 T6 AISI 4140
SPEC
IFIC
CO
ST R
S K
G /
M N
0
5
10
15
20
25
30
Ductile Iron Ti6Al4V AA 6351 T6 AISI 4140
ELO
NG
AT
ION
%
37
2. Ductile cast iron is the cheapest material which is to be considered, also fact that
most of the commercial automobile are fit with ductile cast iron uprights.
3. Titanium alloys have the highest specific cost among all
Considering all the above, it has been decided AA 6351 T-6 seems to be the most viable
material for the component. The strength on weight ratio is sufficient to meet our
standards. Choosing aluminum alloy also seems to be a most economical selection.
Since Aluminum alloy have been chosen, fatigue characteristics of haven to be taken into
account.
Components subjected to fluctuating loading fail at much lower loads than their service
loads due to fatigue. When not considered in the design stage it can lead to catastrophic
failures in their service life. Factors causing a fatigue loading may be many which are:
1. Large fluctuation of Loads i.e. Stress amplitude of high magnitude
2. Sufficient large number of cycles of stress applied
Additional factors which may exaggerate fatigue failure are:
1. Corrosion
2. Residual stress
3. Stress concentration
4. Temperature
5. Surface finish
6. Stress range
7. Use of welding
The mechanism of fatigue failure can be simple put into two stages i.e. crack initiation
and crack propagation to the point of static failure. This crack once formed begins to
grow in each cycle of loading. Growth is also accelerated by higher amplitude of loading
often characterized by multiple cracks initiated. Final catastrophic failure occur when a
crack has grown to a significant length such that the next application of load results in
static failure due to reduced area in that region. Fatigue cracks may start at various places
such as at concentration of plastic strains, extrusion or intrusions of the surface, grain
boundaries, internal voids and surface scratches.
Design methodology for any aluminum component used in car was to:
1) Load be highly approximated to actual condition which is being acted upon the
component
38
2) Optimize the component so that the life of the component be under 105 cycles
3) Check for maximum stress under the stated endurance range instead of checking
for yield criteria.
5. 2.3 Modeling
The uprights geometry determines the dynamic characteristics of the vehicle. Irrespective
of the model, geometry is kept same in all the models. The models prepared were first
based on manufacture feasibility, economic consideration and analytical sustainability.
Various models were initially prepared for the same.
The following are the various models prepared and the reasons for disapproval of the
designs.
39
Design-1 : Saruman
The first of the designs was initially desired to be made up of Steel plates. The plates were given
the thickness of 10mm. The plates were countered cut with the required measurements and
profile. Joining process was chiefly welding i.e. TIG (Tungsten Inert Gas) welding.
Adaptability of the Design :
o Cheap material cost o Easy Manufacturing
o Less event completion time
o Provision of steering stopper
Disapproval of Design
o Weight of the component- 2.5 Kgs.
Since high thickness of plates were being used the weight of the component was considerably
40
high. Also use of 10mm thick sheet was indispensable as any reduction of thickness was
detrimental.
o Low fatigue strength
Since the plate were joined by welding process the fatigue strength of the component was
considerably effected. This is because uncontrolled cooling and heating rates produced
around the weld zone, also called as heat effected zones. This microstructural variation
produces areas of high hardness whereby reducing strength.
o Steering upright geometry not accurate
Since manual fabrication processes were being employed, the dimensional accuracies were
uncertain.
Design-2 : Gollum
Mild Steel pipe of outer diameter 65mm and thickness of 2.5 mm was found to be adequate to
press fit wheel bearing. The profiled Mild steel rods following the geometry was welded to the
mother rim.
Adaptability of design
41
o Simple design
o Easy manufacturing
o Easy procurement of the materials required
Disapproval of design
o Intricate couping requirement.
o Welding reduces the fatigue strength and improper hardness in the sample.
o Strength not uniform due to alternate heating and cooling rates.
o Brake mounting arms are weak.
Brake mounting arms due to relative length from the parent pipe, became long and prone to
failure
o Weight of the component
As mild steel tubes were used the weight of the component was measured up to 1.2 Kgs.
o Poor Strength
Finite Element analysis found it to be weak at weld zones. Hence prone to failure
42
Design-3 : Gandalf
Stock Aluminum alloy 6351 T-6 would have been used. Manufacturing process being CNC
milling.
Adaptability of design
o Geometry of the upright was captured better o CNC milling process was fast and accurate
o High torsional rigidity.
Disapproval of design
o Mounting points for upper and lower wishbone was too small whereby increasing stress in those areas.
o Increasing the upper and lower wishbone mount areas increased the weight of the
component considerably.
43
o Weight of the component
It weighs 2.5 Kgs comparable to its steel counterpart.
Design-4 : Bilbo
It was then decided to provide stub-axle on the upright itself. The aluminum Upright was then
modeled to be CNC milled.
Adaptability of design
o Fast and accurate manufacturing
o Holes and slots provided to relief weight. It weighs only 760 gm. The least weight until now.
o Geometry captured accurately and effectively
o Bearing size is considerably reduced, whereby reducing weight.
44
o Bearing lock could be provided on built in stub axle
Disapproval of design
o Many relief holes increased notch sensitivity of the component o Built in stub axle cannot have wheel bolt as threading on aluminum shears off easily.
o Required hub offset cannot be changed.
o In case of failure of stub axle whole upright have to be changed
o Longer stub axle created immense bending strength on upright o Fillets and notches are requirement on the stub axle to be able to hold other components
well. These fillets and notch sensitivity added up stress on bending stress, creating more
stress it can hold o Built in stub axle was failing in large loading conditions.
o Longer steering arm also made it more susceptible to shear from the upright
o Also built in steering arm dictates changing the upright when steering arm fails, the
failure which is very common in the testing.
Rejection of all the other designs was valid as even though they all met with the
requirement of designing parameters but failed at various other important suppositions.
New design was required to have the modeling advantages of all other designs at the
same time to rectify all the drawbacks of the previous designs.
The new and improved design was modeled in the following steps.
Step 1:
The basic mold was created around the upright geometry. This captures almost all
the features which were important to suspension design.
45
Step 2 :
The next step was to differential the upper and lower wishbone mounting
positions. An arbitrary thickness was given which will be verified once the model
is added to testing module. Care was also taken care to be able to provide
sufficient room for accommodation of castle nuts.
Step 3 :
The brake mount position was carefully positioned so as when brake caliper in
engagement would reach the end of the brake disk, for it to hold it firmly when
under braking effect. The longer brake mount was given a parabolic profile to
have a uniform strength also to reduce material.
After calculated width of material was left for the center hole on the upright which
would hold the stub axle.
The extended length of the steering arm was reduced to decrease the added
moment at the end of the upright.
Step 4:
After bail calculation considering the material, diameter and step dimensions for
the stub axle. the central bore for the same was provided on upright
46
Step 5:
It was seen as an added strength to steering arm by placing its bolt under double
shear. Also to counteract the effect of couple number of holes for its bolts were
increased from 1 to 2.
Step 6:
The castor was given 5 degrees, according to which the position of upper and
lower mounting points were determined. The hole diameter was kept equal to the
rod-end diameter.
47
The model was thus prepared with rectifying all the drawbacks obtained from the
previous failures of various models, also every model was an improvement over the other
until the final design (Hobbit) was considered to be selected.
The various dimension were kept variable in view of further analysis. Finite element
analysis will dictates all the dimension of the component.
5. 2.4 Loading scenarios
Contemplation of actual force distribution on a single suspension component during
motion of car can be quite complicated. To ease this complication and various
48
complicated vector forces experienced by the component, it is usually split down into
various individual scenarios. These scenarios take up the maximum force derived from
calculation and constrains the model appropriately to represent the actual event.
The following scenarios of loading will be considered for the analysis
1. Steering effort
2. Brake mounting force
3. Remote loading
4. Cornering Force
It is to be noted that weight of the component was considered during analysis.
1. Steering Effort Calculation,
As steering effort was calculated in section-12, Steering effort (Fst) is taken as 1.5g
𝐹𝑠𝑡 = 4400
The tie rod being designed to take purely axial load, whereas such is not the case with
steering arm. As the wheel travels the direction of forces changes. To accommodate this
effect the direction of steering force wasn’t taken parallel to ground but at an angle 45o
with the horizontal.
Then components along X and Y comes out to be equal i.e.
𝐹𝑠𝑡 𝑥 = 𝐹𝑠𝑡 𝑦 = 3100
2. Braking Mount force calculation,
49
The following free body diagram shows the position of brake caliper mounting pints with
respect to center of the disk. This is because the brake mounting position is not
symmetrical with the center of the caliper. And hence the force experienced by the each
mount will be different.
Braking Mount Force (Fbm)
𝐹𝑏𝑚 =𝑇𝑏
𝑟𝑏
By sum of forces and applying moment at a point we get,
Force on larger brake mount ꞊ 3200 N
Force on smaller brake mount ꞊ 2300 N
3. Remote loading
Due to rim and hub offset, the bumps force is not a direct loading but an eccentric
loading. The values of the mentioned dimension have already been calculated.
Fshock have already been calculated (6. c.vi) ꞊ 22,000 N
4. Cornering force
As per calculation the cornering force was taken to be 1.5g
𝐹𝑐𝑜𝑟𝑛𝑒𝑟 = 5000
50
5. 2.5 Analysis
An upright is the crucial component as every force experienced by the car is passed on
through it. Also analyzing every effect of this complex ever changing moment and forces
becomes complicated. A worst case scenario’s is therefore recommended where forces
are scaled and restrains are applied in view of real time. The team performed various
finite element computational analysis to match approximately with the actual conditions
that will be experienced by the component so as to avoid failure in real-time.
There are many different types of analysis that must be completed to ensure that the part
in question is able to withstand the applied loads. In addition, there are other factors that
must be included in each analysis to ensure that the analysis itself is correct. In order to
analyze the part correctly, the restraints must be an accurate representation of the real
world scenario and the loads must be calculated for different loading scenarios.
Finally, the mesh must be as homogenous as possible. This would include minimizing the
difference in aspect ratio between elements, as well as maximize element mapping
quality. We must ensure that all of the meshes we use for the different components in our
assembly are set up to be compatible with one another.
The main objective behind analysis was to check the maximum stress induced, predict the
life of component and establish a suitable factor of safety in design.
Since most of suspension component used were aluminum it must be noted forth that it
doesn’t exhibit a fixed fatigue limit unlike major steel categories.
All of the analysis was done in student package of Solid works 2013 Simulation module.
The meshing package utilizes the tetrahedral mesh on the component. To give better
accuracy the mesh size was made finer until the results became stagnant. Any more
decrease in mesh size would just waste computer resource without marginally increasing
any solution accuracy.
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Modal analysis was also done on the component to determine how the system behaves in
its displacement dynamic response. It was done to check how different frequencies
naturally excite the system to a degree where resonance fluctuates through the component
resulting in dropping of life expectancy of the system. For this study no external load is
applied to the component, while it is known that external loads do not affect the natural
frequency. The component was however restrained.
In all the analysis self-weight of the component was considered.
Scenario Bump loading
Loading Stub-axle bore at an
offset of 50mm
Constrains Upper and lower
wishbone mounting
Force 22,000 N
Maximum Stress 109 MPa
Maximum Deflection 0.053 mm
Factor of safety 2.42
Scenario Brake caliper mount
loading
Loading Caliper holes
Constrains Upper and lower
wishbone mounting
Force 3200, 2300 N
respectively
Maximum Stress 100 MPa
Maximum Deflection 0.3 mm
52
Factor of safety 1.88
Scenario Cornering of vehicle
Loading Upper and lower
wishbone mounting
Constrains Stub-axle bore
Force 5000 N
Maximum Stress 53 MPa
Maximum Deflection 0.16 mm
Factor of safety 4.91
Scenario Steering of vehicle
Loading Steering arm holes
Constrains
Upper and lower
wishbone mounting
points
Force 3000 N
Maximum Stress 90 MPa
Maximum Deflection 0.13 mm
Factor of safety 2.73
The above plots of the results for various cases that might be experienced the component
during its operation life cycle
53
Being the suspension component of an automobile, it is under constant non uniform
loading condition. Also the loading condition vary drastically from terrain. In order to
simplify the loading scenarios fully reversed loading i.e. R (Stress Ratio) ꞊ -1 was
preferred over other stress ratios ranging from -∞ to ∞. The tabular values of Sa (Stress
amplitude) with the corresponding life cycles [1] was fed into Solid works Fatigue
Analyzer. The corresponding S-N diagram was obtained from the data.
Solid works doesn’t have built in S-N graphs for all the materials that are present in its
directory, hence it was needed to plug them manually into the software.
Following is the obtained fatigue data that has to be fed into the system to perform
fatigue analysis.
Table 1: Fatigue data distribution for Aluminum 6061 T6
No of Cycles Stress Amplitude
(Zero Mean Stress)
N/m2
10.000000 482000000.000000
70.000000 482000000.000000
100.000000 420000000.000000
200.000000 325000000.000000
500.000000 241000000.000000
1000.000000 198000000.000000
2000.000000 168000000.000000
5000.000000 142000000.000000
7000.000000 135000000.000000
10000.000000 120000000.000000
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20000.000000 99000000.000000
50000.000000 80000000.000000
100000.000000 71000000.000000
200000.000000 64000000.000000
500000.000000 58000000.000000
1000000.000000 55000000.000000
2000000.000000 53000000.000000
5000000.000000 51000000.000000
10000000.000000 50000000.000000
20000000.000000 49880000.000000
50000000.000000 49280000.000000
100000000.000000 48990000.000000
200000000.000000 48700000.000000
500000000.000000 48500000.000000
1000000000.000000 48400000.000000
55
Figure 1: S-N curve for zero mean stress of Aluminum 6061 T6
The fatigue analysis could be done for individual scenarios however for a combined force
situation it seems logical to optimize. For this bump force, heavy braking, steering pull
were applied to component.
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It can be seen the life of the component is well defined for more than 95% area having
1E9 cycles. Some of the areas shows a reduced life but that can be ignored potentially.
All attempts were made to restrict the maximum von misses stress under 100 MPa, after
studying the fatigue properties of aluminum. Also not compromising on weight gave the
best combination of strength, weight and fatigue characteristics for the component.
5. 3 Design of Front Hub
5. 3.1 Introduction
A wheel hub is a mounting position for wheel of the vehicle, it houses the wheel bearing
as well as supports the lugs and brake disk. It can either transmit power or be just rolling.
Its function is basically to keep the wheel spinning freely on the bearing while keeping it
attached to the vehicle. Designing a hub is very crucial as it alones is the interface
between the wheels and rest of the vehicle.
Lug bolts are usually integral to hub, hence these are also called as locking bolts.
Spacers are also used to fit between the hub and the brake disks. This is done to
accommodate different brake calipers as to avoid the scraping between them and the
calipers.
Usually in hubs in commercial vehicle are made up of alloy steels or cast iron. But for a
mini Baja vehicle it is advantageous to look at alternative material to make it light
weight.
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5. 3.2 Bearing selection
Selection of suitable bearing for a particular purpose is immensely important in view of
the load it is meant to take at given rotational speed giving a certain life.
Max gear ratio of the transmission,
7.6
Max rpm of the engine crank,
3800 𝑟𝑝𝑚
The maximum rotational speed attained by the tire is,
500 𝑟𝑝𝑚
Assuming the life of the bearing to be designed for is,
1000 ℎ𝑜𝑢𝑟𝑠
Loading ratio (From data book)
𝐶
𝑃 = 3.11
Axial Force is assumed to be during cornering, which is taken as 1.5g (Pr)
2500 𝑁
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Radial Force is the drop weight of the car (Pa)
4000 𝑁
the ratio of axial to radial,
𝑃𝑎
𝑃𝑟 = 0.45
for the corresponding ratio, the value of equivalent load
𝑃 = 𝑆(𝑋𝑃𝑟 + 𝑌𝑃𝑎)
𝑃 = 4000 𝑁
The dynamic load carrying capacity was found out to be,
12995 𝑁
For the given life and load rating the bearing number SKR 6007 was chosen for the front
wheels.
The given are dimensions of the bearing
5. 3.3 Designing Parameters & Considerations
Following are the Parameters which guided the design of Front Hubs,
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1. Pitch Circle Diameter of Lug Bolts on rim
Since stock Maruti rims were chosen to be on the vehicle on all four wheels. In order that
the hub to sit on rim, the pitch circle diameter of the rim had to match with the designed
hubs.
Pitch Circle Diameter of Lug bolts on hub ꞊ 144 mm
2. No of Lug bolts and their size
Since the rim had 4 equi-spaced lug bolts holes. With the hole size of 12.5 mm diameter.
The holes to match with rim had to be provided on hub.
4 Lug Bolt holes with diameter ꞊ 12 mm ø
3. Bearing Provision
Since as mentioned earlier the bearing selected was SKF 6007. The bore on the hub with
sufficient tolerance is to be provided for the bearing to sit inside the hub.
The bearing bore on hub ꞊ 62 −0.03 0
4. Bearing Seat
For able to lock the bearing in the hub on one end it should have bearing seat. The
thickness of seat must be enough to withstand axial force while cornering.
5. Common Holes for Rim and Brake disk mounting
Common holes for both rim and brake will minimize the number of holes from 8 to 4 on
the hub.
6. Bearing Lock provision
On other end of the hub, the bearing will be locked by an internal circlip. A groove of
2mm is to be provided to facilitate the circlip.
5. 3.4 Modeling
A literature survey was undertaken before modelling of the component began. Initially as
a reference the stock Maruti hub was taken. The model was loaded in the Solid-works.
Initially it was decided to use it but due to its weight idea was soon dropped. Also
component was analyzed to develop a lighter and stronger equivalent in aluminum alloy.
60
The details in the model were removed to simplify the analysis. The brake mounting
holes were also removed to see the effect of analysis. Since the goal of the analysis will
be to target potent areas of weight reduction and geometry changes to suit the Baja
vehicle. Such method reduces the chances of failure of design as the adopted design is
commercially utilized. Also as mentioned in design parameter stub axle will be
eliminated in the front hubs to accommodate hub bearing which in turn will hold the axle.
An arbitrary force of any value have been loaded on the stub axle of the hub in context.
The constrains were the lug holes. The plot shows the stress distribution close to stub axle
extending towards lug holes. It was also evident that rest of the area experienced a
relatively less induced stress.
A plot of design insight reveals the observation to a scaled level.
61
The study of stock hub gave us the insights to new design of the hubs that would be
precisely follow its design guidelines.
Following is the timeline of the model of the component.
The blank is initially made and stepped to
host the wheel bearing. Provision is also
given to lock the bearing in the other
direction by a bearing seat.
The Pitch circle diameter equivalent to rim is
kept on the hub.
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A single chuck of material is removed to
observe the effect.
The pattern is then repeat around to obtain
optimized wheel hub.
The grove for the circlip is provided to lock
the bearing in opposite direction.
Hence modelling is complete.
5. 3.5 Loading scenarios
Following are the various loading that were applied to hubs for analysis. This of these
forces so obtained will help the prediction of actual results.
1. Drop test
63
This test try to replicate stress produced when the vehicle falls from a certain height. To
obtain the impact load on the hubs,
Assuming the vehicle falls from the height (h),
ℎ = 1 𝑚
The impact velocity (v) of the vehicle is,
From kinematic relation for rectilinear motion
𝑣 = √(2 ∗ 𝑔 ∗ ℎ) − 𝑢2
whereas,
g ꞊ acceleration due to gravity (9.8 m/s2 )
u ꞊ Initial velocity of the vehicle before the fall ( 0 m/s)
Substituting the above value,
𝑣 = 4.42 𝑚/𝑠
And now for calculating the impact force (F)
From Work-Energy Principle,
Change in Kinetic Energy of the object ꞊ Work done on the object
1
2 𝑚 ( 𝑣2 − 𝑢2) = 𝐹 ∗ 𝑑
whereas,
d ꞊ the compression of the shock springs ( 0.15 m)
Thus,
𝐹𝑠ℎ𝑜𝑐𝑘 ꞊ 22,800 𝑁
2. Braking Torque Test
This test analysis the effect of panic braking on wheel hub. Since brake disk and hub are
directly connected, while under braking, brake disks induces opposite torque on the hub
to halt the vehicle.
Assuming the vehicle comes to a complete halt from 55 Km/h in a distance within 6m on
an off-road track.
Initial velocity (u) ꞊ 15.2 m/s
64
Braking distance (d) ꞊ 6m
Deceleration (ad),
𝑣2 = 𝑢2 + 2 ∗ 𝑎𝑑 ∗ 𝑑
whereas,
v ꞊ final velocity
thus,
𝑎𝑑꞊ 11.7 𝑚/𝑠2
Assuming weight distribution is 50:50, Force on Front wheels (Ff),
𝐹𝑓 ꞊ 2047.5 𝑁
Braking torque on front wheels (Tf) is,
𝑇𝑓 ꞊ 𝐹𝑓 ∗ 𝑅𝑡
whereas,
Rt ꞊ Radius of Tire (0.3048 m)
Thus,
𝑇𝑓 ꞊ 624 𝑁𝑚
3. Cornering & Skidding
Slip angle changes at the turn of the vehicle, sometimes amateur driver may fail
understand its significance and it results in vehicle skidding in turns.
It has been taken as 1.5g force for skidding whereas 1g for cornering force. As speed of
the vehicle is restricted the assumed values holds good.
Cornering Force ꞊ 3500 N
Skidding Force ꞊ 5200 N
4. Rim fracture
In the event of a rim failure, i.e shear of rim across the lug bolts or failure from flange, it
produces an eccentric loading at the hub due to its rim offset.
To take this into account the self-weight of the vehicle will be taken in consideration.
65
5. 3.6 Analysis
Scenario Drop Test
Loading Central bore
Constrains 4 Lug holes
Force 11,000 N
Maximum Stress 100 MPa
Maximum Deflection 0.041 mm
Factor of safety 2.5
Scenario Braking Test
Loading Central bore
Constrains 4 lug holes
Torque 700 Nm
Maximum Stress 83 MPa
Maximum Deflection 0.022 mm
Factor of safety
Scenario Cornering & skidding
Loading Skidding-Bearing seat
66
Cornering-Central bore
Constrains 4 Lug holes
Force Skidding-5200 N
Cornering-3500 N
Maximum Stress 131 MPa
Maximum Deflection 0.049
Factor of safety 1.96
Scenario Rim fracture
Loading 3 spokes
Constrains Central bore
Force 4000 N
Maximum Stress 77 MPa
Maximum Deflection 0.11 mm
Factor of safety 3.2
67
To optimize the component even further it was then added into Solid works Optimization.
The variable that was kept in the study was the thickness of the blank of the hub.
Following results were found,
Component
name Units Current Initial Optimal Scenario1 Scenario2
thickness mm 8 8 4 4 8
Stress1 N/mm^2
(MPa) 85.493 85.493 172.74 172.74 85.493
Mass1 Kg 0.151107 0.151107 0.115572 0.115572 0.151107
It can be seen that reducing the thickness from 8 mm to 4 mm would increase the max
stress in the component to 172.74 MPa which is fairly under the limit. The weight of the
component could also be reduced by 25%. But the optimization wasn’t carried out in the