tobias doyle final thesis

JAMES COOK UNIVERSITY

SCHOOL OF ENGINEERING &

PHYSICAL SCIENCES

EG4011

Mechanical Engineering

Design & Experimentation of Un-sprung

Suspension Components for Formula S.A.E.

Vehicles

Tobias Doyle

Thesis submitted to the School of Engineering & Physical Sciences in

partial fulfilment of the requirements of the degree of

Bachelor of Engineering with Honours (Mechanical)

5th October 2012

Doyle, T. Design & Experimentation of Un-sprung Suspension Components

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Statement of Access

I, the undersigned, the author of this thesis, understand that James Cook University will

make it available for use within the University Library and, by microfilm or other means,

allow access to users in other approved libraries. All users consulting with this thesis will

have to sign the following statement:

In consulting this thesis I agree not to copy or closely paraphrase it in whole or

in part without written consent of the author; and to make proper public written

acknowledgement for any assistance which I have obtained from it.

Beyond this, I do not wish to place any restriction on access to this thesis.

Tobias William Doyle Date


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Sources Declaration

I declare that this thesis is my own work and has not been submitted in any form for another

degree or diploma at any university or other institution of tertiary education. Information

derived from the published or unpublished work of others has been acknowledged in the text

and a list of references is given.

Tobias William Doyle Date


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Abstract

Formula S.A.E. is an international undergraduate competition that involves students

designing and building an open wheel style race car. The design and fabrication of the car is

proven in various events aimed at testing performance and safety. Currently the James Cook

University Motorsports team is designing and building a first generation car for competition

in the Australasian Formula S.A.E competition.

This project is aimed at developing and implementing a suspension verification process for

the James Cook University Motorsport vehicle. In order to be competitive in the Formula

S.A.E competition vehicles must be safe and reliable, which is only achievable through

testing and continuous development. However, as James Cook University Motorsports team

is building a first generation car there is no current method of testing components or

determining with confidence forces for design.

As a part of this project a suspension test apparatus has been designed and manufactured to

allow laboratory testing to be undertaken to test suspension components without having to

drive the car. This also allows verification of modelling techniques against experimental data

to increase confidence in design methods. This iterative process will allow for development

and optimization of design and ultimately result in safer and more reliable designs.

Using the designed apparatus strain results were collected showing that the most stressed

component of the suspension system is the bottom A-Arm. Using the strain results as a

comparison, FEA models of the suspension varied in accuracy. Two of the models

developed had an accuracy of approximately 30%. Further investigation in to modelling

techniques for the suspension can be developed based on these models, and the assumptions

made in these models.

The Upright and Hub shaft have been designed and verified and are ready for manufacture.

FEA modelling on these components was performed with multiple load cases. Both

components are expected to have an infinite, as both components have adequate fatigue

safety factors.


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Acknowledgements

I would like to thank Dr David Holmes, my thesis supervisor and lecturer at James Cook

University for his help and guidance throughout this thesis and for his input in to this final

thesis report.

I would also like to thank Curtis Arrowsmith for his guidance and knowledge in

manufacturing techniques and for his involvement in the fabrication of thesis based

components.


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Table of Contents Statement of Access ............................................................................................................ 2

Sources Declaration ............................................................................................................ 3

Abstract .............................................................................................................................. 4

Acknowledgements ............................................................................................................. 5

List of Figures ..................................................................................................................... 9

List of Tables .................................................................................................................... 11

List of Symbols ................................................................................................................. 12

1 Introduction ................................................................................................................... 13

1.1 Problem Definition .................................................................................................. 13

1.2 Project Aims ............................................................................................................ 13

1.3 Motivation ............................................................................................................... 14

1.4 Suspension Components .......................................................................................... 14

2 Literature Review ........................................................................................................... 16

2.1 Analytical and Computational Methods for Suspension Design ................................ 17

2.1.1 Past JCU Motorsport Analysis............................................................................ 18

2.1.2 Other FSAE Designs ........................................................................................... 22

2.2 Vehicle Behaviour Inputs ........................................................................................ 23

2.3 Suspension Test Methods ........................................................................................ 25

2.3.1 Formula 1 Honda Racing Team Suspension Testing ........................................... 26

2.4 Data Acquisition ...................................................................................................... 29

2.5 Fatigue Life Calculation .......................................................................................... 30

2.6 FSAE Specific Load Cases ...................................................................................... 30

2.7 Conclusion .............................................................................................................. 31

3 Design Methodology ...................................................................................................... 32

3.1 Test Apparatus Design ............................................................................................. 32

3.2 Detailed Apparatus Design ...................................................................................... 37

3.3 Analysis and Verification of Apparatus .................................................................... 40

4 Experimental Analysis ................................................................................................... 46

4.1 Experimental Apparatus Setup Procedure ................................................................ 46


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4.2 Apparatus Experimental Procedure .......................................................................... 50

4.2.1 Risk Assessment................................................................................................ 50

4.3 Apparatus Experimental Results .............................................................................. 51

4.3.1 Apparatus Experimental Results........................................................................ 51

4.3.2 Experimental Discussion ................................................................................... 56

5 Computational Simulation .............................................................................................. 58

5.1 Computational Simulation Results ........................................................................... 58

5.1.1 Further Model Investigation.............................................................................. 64

5.1.2 Further Push Rod Simulation ............................................................................. 67

5.2 Computational Simulation Discussion...................................................................... 69

6 Upright Verification ....................................................................................................... 71

6.1 Detailed Upright Analysis........................................................................................ 73

6.3 Upright Recommendations ...................................................................................... 78

7 Hub Shaft Redesign ........................................................................................................ 79

7.1 Hub Shaft Testing Method Investigation .................................................................. 81

7.2 Hub Shaft Detailed Design ...................................................................................... 85

7.3 Keyway Detailed Analysis ....................................................................................... 91

7.3.1 Weld Analysis ................................................................................................... 94

7.4 Hub Shaft Recommendations ................................................................................... 95

8 Conclusion ..................................................................................................................... 98

9 References ................................................................................................................... 100

Appendix ........................................................................................................................ 102

A.1 Apparatus Drawings ............................................................................................. 102

A.3 Hub Shaft Drawings ............................................................................................. 107

A.4 Relevant FSAE Suspension Rules ......................................................................... 117

A.5 Project Timeline ................................................................................................... 118

A.6 Hand Calculations ................................................................................................ 119

A.6.1 Weight Arm Calculations ................................................................................ 119

A.6.2 Push Rod Strain Calculations ........................................................................... 119

A.6.3 Hub Shaft Weld Calculations ........................................................................... 120


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A.7 Bolt Calculations .................................................................................................. 121

A.8 Risk Assessment Documents ................................................................................ 127

A.10 HBM Strain Gauge Calibration Data ................................................................... 129

A.11 Material Properties and Data Sheets .................................................................... 130

A.11.1 Alumec Material Properties .......................................................................... 130

A.11.2 7075 Aluminium Alloy Fatigue Data .............................................................. 131

A.11.3 4140 Steel Properties ................................................................................... 132

A.11.4 Welding Properties ....................................................................................... 133

Word Count: 19, 110


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List of Figures

Figure 1: Solidworks Drawing of Omega with Basic Suspension Components [3] ............. 14

Figure 2: Front Suspension Geometry Diagram without Wheel* ........................................ 15

Figure 3: Design Process ................................................................................................... 16

Figure 4: Transient Structural Analysis using Spring, for Un-sprung Mass [3] ................... 18

Figure 5: Point Mass Study [1] .......................................................................................... 18

Figure 6: Remote Force Application to Wheel, Hub Shaft and Upright [1] ......................... 19

Figure 7: Upright 2011 Design [1] ..................................................................................... 20

Figure 8: Exploded View of Recommended 2011 Hub Shaft Design [2] ............................ 20

Figure 9: Final Upright FEA Model (McMune et al [11]) .................................................. 23

Figure 10: Vehicle Pitch, Roll and Yaw [13] ..................................................................... 24

Figure 11: Wheel Camber [15] .......................................................................................... 24

Figure 12: MOOG 4-poster suspension tester [17] ............................................................. 25

Figure 13: MOOG 16-DOF suspension test rig [18] ........................................................... 26

Figure 14: Uniaxial suspension test [7] .............................................................................. 27

Figure 15: Diagram of Forces Acting on a Wheel [7] ......................................................... 27

Figure 16: 12-Axis Suspension Testing [5] ........................................................................ 28

Figure 17: Strain gauge diagram [20] ................................................................................ 29

Figure 18: FSAE skip pan event ........................................................................................ 31

Figure 19: Test Apparatus Prototype 1............................................................................... 35



Figure 22: Final Apparatus Design with Suspension Geometry .......................................... 38

Figure 23: Suspension Test Apparatus Diagram................................................................. 38

Figure 24: ANSYS Apparatus Meshing ............................................................................. 40

Figure 25: Force Loading on Apparatus ............................................................................. 41

Figure 26: Apparatus Safety Factor ................................................................................... 41

Figure 27: Apparatus High Stress Areas ............................................................................ 42

Figure 28: Deflection of Apparatus.................................................................................... 42

Figure 29: Weight Arm Stress Distribution ........................................................................ 43

Figure 30: Weight Arm Total Deformation ........................................................................ 43

Figure 31: Weight Arm Safety Factor ................................................................................ 44

Figure 32: Axis Positioning ............................................................................................... 44

Figure 33: Hub Shaft and Upright Geometry Exploded View [1] ....................................... 48

Figure 34: Exploded View of Suspension Assembly .......................................................... 49

Figure 35: Metal Finish before Strain Gauges Attached ..................................................... 51


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Figure 36: Completed Strain Gauge ................................................................................... 51

Figure 37: Strain Gauge Positioning .................................................................................. 52

Figure 38: Positioning of Strain Gauge Locations on Push Rod ......................................... 52

Figure 39: Strain Gauge Positions on the Push Rod ........................................................... 52

Figure 40: Completed and Assembled Test Apparatus ....................................................... 53

Figure 41: Strain Result Graph .......................................................................................... 54

Figure 42: Spring Length vs. Load .................................................................................... 55

Figure 43: Spring Force on Push Rod ................................................................................ 56

Figure 44: Model A Setup ................................................................................................. 58

Figure 45: Model A Mesh ................................................................................................. 59

Figure 46: Model A Strain Results .................................................................................... 59

Figure 47: Model B Setup and Mesh ................................................................................. 60

Figure 48: Model B Strain Results ..................................................................................... 61

Figure 49: Model B Exaggerated Deformation .................................................................. 62

Figure 50: Model C Setup ................................................................................................. 62

Figure 51: Model C Strain Results ..................................................................................... 63

Figure 52: Model D Setup and Strain Results .................................................................... 63

Figure 53: Strain Gauge 1, Model B and D Strain vs. Experimental Strain ......................... 64



Figure 56: Model B Push Rod Result................................................................................. 67

Figure 57: Simple Push Rod Analysis using Realistic Constraints ...................................... 68

Figure 58: 2011 Upright Design ........................................................................................ 71

Figure 59: Upright Machining Configuration ..................................................................... 72

Figure 60: 2012 Upright Design ........................................................................................ 73

Figure 61: Load Case 1 Geometry ..................................................................................... 74

Figure 62: Load Case 1 Mesh ............................................................................................ 75

Figure 63: Load Case 1 Stress ........................................................................................... 76

Figure 64: Load Case 1 Fatigue Safety Factor.................................................................... 76

Figure 65: Forces Analysed for Load Case 2 ..................................................................... 76

Figure 66: Load Case 2 Mesh ............................................................................................ 77

Figure 67: Load Case 2 Stress ........................................................................................... 78

Figure 68: Load Case 2 Fatigue Safety Factor.................................................................... 78

Figure 69: 2012 Current Progress of Upright ..................................................................... 78

Figure 70: Hub Shaft General Assembly ............................................................................ 79

Figure 71: Exploded Conceptual View of Hub Shaft 3.0 .................................................... 80

Figure 72: Basic Hub Shaft Solid ...................................................................................... 81


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Figure 73: Force Application Verification Mesh ................................................................ 82

Figure 74: Forces on Wheel using a General Body to Ground ............................................ 82

Figure 75: Force Application for Wheel Contact Patch Model ........................................... 83

Figure 76: Fatigue Safety Factor for, ................................................................................. 83

Figure 77: Remote Force Location .................................................................................... 84

Figure 78: Fatigue Safety Factor for Remote Force Loading .............................................. 84

Figure 79: Load Case A Mesh and Model .......................................................................... 86

Figure 80: Load Case A Setup ........................................................................................... 86

Figure 81: Load Case A Stress Plot (Cross-Sectional View) .............................................. 87

Figure 82: Load Case A Fatigue Safety Factor ................................................................... 87

Figure 83: Load Case B Moment Setup ............................................................................. 88

Figure 84: Load Case B Torque Only Stress (Cross-Sectional View) ................................. 89

Figure 85: Load Case B Moment Fatigue Safety Factor ..................................................... 89

Figure 86: Load Case B Remote Force Setup ..................................................................... 90

Figure 87: Load Case B Remote Force Stress (Cross-Sectional View) ............................... 90

Figure 88: Load Case B Remote Force Stress .................................................................... 91

Figure 89: CV Verification Setup ...................................................................................... 92

Figure 90: CV Verification Stress ...................................................................................... 92

Figure 91: Keyway Verification Setup ............................................................................... 93

Figure 92: Keyway Verification Stress .............................................................................. 93

Figure 93: Keyway Verification Fatigue Safety Factor ...................................................... 94

Figure 94: Hub Shaft Weld Locations ................................................................................ 94

Figure 95: Final Hub Shaft Design .................................................................................... 95

Figure 96: Final Front Hub Shaft ....................................................................................... 96

Figure 97: Final Hub Shaft Exploded View ....................................................................... 96

Figure 98: Final Thesis Assembly ..................................................................................... 98

Figure 99: Project timeline .............................................................................................. 118

List of Tables

Table 1: Deceleration (1.5G), Cornering (1.4) and Bump Force (1G) Load Case [1] .......... 19

Table 2: Additional Rear Hub Shaft Forces [2] .................................................................. 21

Table 3: Acceleration (1G), Cornering (1.4) Load Case [1] ................................................ 21

Table 4: Connection Reaction Forces (Hunter et al. [1]) Load Case 2 ................................ 21

Table 5: Upright Forces (McMune et al. (2009) [11]) Based on Figure 9 ........................... 22

Table 6: Apparatus Design Criteria.................................................................................... 34


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Table 7: Suspension Apparatus Design Matrix .................................................................. 37

Table 8: Bolt Joint Probe Results ....................................................................................... 45

Table 9: Hub Shaft and Upright Geometry BOM [1] ......................................................... 47

Table 10: Suspension Assembly BOM .............................................................................. 49

Table 11: Experimental Results ......................................................................................... 54

Table 12: Model A, Computational and Experimental Strain Comparison .......................... 60

Table 13: Model B, Computational and Experimental Strain Comparison .......................... 61

Table 14: Model C, Computational and Experimental Comparison .................................... 63

Table 15: Model D, Computational and Experimental Comparison .................................... 64

Table 16: Strain Gauge 1, Model B and D vs. Experimental Strain .................................... 65



Table 19: Push Rod Modelling Comparison ...................................................................... 68

Table 20: ANSYS Fatigue Data for Alumec ...................................................................... 74

Table 21: 4140 Steel Fatigue Properties ............................................................................. 85

List of Symbols

Stress [

Modulus of Elasticity [

Strain

kilograms

metre

millimetre

Newton

pound

foot


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1 Introduction

1.1 Problem Definition

The Formula Society of Automobile Engineers (FSAE) competition is an international

competition requiring students to design, build and market a custom racer. The design and

reliability of the car is tested through a number of different events aimed to highlight

weaknesses in the car’s design and construction. To be competitive, the car must be fast,

light and handle well. The James Cook University (JCU) motorsport team is currently

building a first generation FSAE car that is referred to as Omega.

1.2 Project Aims

This project will incorporate the re-design of Omega’s un-sprung suspension components

(i.e. the hub shaft and uprights) that are critical, complimentary items in the suspension

geometry. The un-sprung suspension components have had previous design work done [1-4],

however the size of the hub shaft was deemed inadequate for the competition stresses and

required further development [2]. For this aspect, computational methods will be further

employed in analysis using past design work, with new considerations.

An experimental test apparatus will be designed for testing the JCU Motorsport vehicle

suspension to represent FSAE competition loads. The apparatus will allow experimental

testing of the suspension geometry and spring rate without driving the vehicle. This allows

for verification of suspension components theoretical and numerical analyses and

optimisation for suspension performance. Having the ability to physically test and optimise

the suspension before driving will reduce the time required to get the car to a competitive

state and increase the safety of the driver.

The apparatus design must allow for experimentation to gather data related to the stresses

exerted on the suspension components under realistic load cases and measuring of key

performance indicators for optimisation of geometry such as spring stiffness.


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1.3 Motivation

At JCU there are presently no methods of experimentally testing car components before the

car is finished and can be driven. This poses many safety concerns, as at high speed a

catastrophic suspension component failure may result in serious consequences for the driver.

Additionally, the final product will be designed based on theoretical analysis alone, which is

never optimal. Laboratory testing has many advantages as it can confirm theoretical

measurements, leads to accelerated vehicle development, and reduces testing and tuning

required. This will result in a more competitive vehicle that is safer with increased

reliability, both important characteristics in any competition.

1.4 Suspension Components

The suspension geometry is a critical component on any vehicle as it determines the way the

vehicle will handle and also behave from external inputs such as bumps and inputs from the

driver. The current suspension geometry is attached to the car as shown in Figure 1 using

brackets. The suspension geometry shown in Figure 2 represents a double wishbone push

rod suspension, which is currently designed for Omega.

Figure 1: Solidworks Drawing of Omega with Basic Suspension Components [3]


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Figure 2: Front Suspension Geometry Diagram without Wheel*

*Wheel is attached to hub shaft

The suspension geometry is made up of several components that are typically referred to as

the un-sprung mass. The un-sprung mass incorporates the wheel rim, braking components,

tyre, hub shafts, upright and bearings. It is ideal to design the un-sprung mass as light as

possible to improve handling and limit forces exerted through the suspension system to the

car chassis (sprung mass). There are several major load cases that are important when

analysing the suspension components; acceleration, cornering and bump loads. Design load

cases and testing methods of the suspension will be discussed further in the literature review

and methodology that is to follow.


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2 Literature Review

There are many examples in Motorsport and in industry of how theoretical analysis and

experimental testing are used to develop vehicle component designs. These techniques and

other information critical to the design and testing of FSAE un-sprung components will be

reviewed in what follows.

Formula 1 represents a competition with state-of-the-art motorsport design, with continuous

development and improvement of cars during the racing season; so much so that the car at

the end of a season can be as much as 75% different to the car at the start [5]. This rapid

development of cars is only made possible by numerous methods of laboratory testing and

data acquisition, from on and off the track. Approximately 80% of the testing on Formula 1

cars takes place in the laboratory either on test rigs or virtually via computer simulation [5].

Laboratory testing allows multiple components to be tested individually and together as a

system to provide information on safety, strength, performance and life before track testing.

A component may complete thousands of kilometres of testing before it even leaves the

factory [6]. Many of the design approaches used by Formula 1 teams can be simplified to the

FSAE competition.

During the design phase many assumptions must be made and these assumptions need to be

verified using laboratory testing and then track testing. Through each of the verification

phase’s, data collected such as forces and stresses, can be used to reduce assumptions during

the design phase or to use in the laboratory testing for load sequences [5-7]. Creating an

iterative design and verification loop is vital to optimise performance and allow

development, this iterative process is shown in Figure 3, a process that is used by the

Formula 1 Honda Racing Team [5].

Figure 3: Design Process

Design & Computational

Analysis Techniques Laboratory Testing Track Testing


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For a first generation car such as Omega that is yet to be track drivable, no initial data on

forces is available for the design. Forces for design therefore must be determined from hand

calculations and computational experimentation. In many cases, it is difficult to fully

envelope the car’s behaviour in different suspension specific load cases, as the suspension

behaviour is dynamic and subject to inertial effects. Suspension specific load cases will be

further discussed in Section 2.1 Analytical and Computational Methods for Suspension

Design.

2.1 Analytical and Computational Methods for Suspension Design

There are different methods of design that can be incorporated to design suspension

components. Through the use of analytical formulas, forces at each tyre can be determined

based purely on weight transfer from acceleration forces at the centre of gravity. However,

the suspension geometry is made up of multiple components that all interact with each other

and capturing this behaviour is difficult using purely analytical methods. Finite Element

Analysis (FEA) is a computational method that can be used to model complex geometries.

FEA modelling is a useful technique to simulate different load cases in order to analyse

stresses to be used in the development and verification of designs.

One FEA modelling package is ANSYS, which offers many different methods that can be

used for different applications. Methods that will be relevant to this project are,

Static Structural

Transient Structural

Explicit Dynamics

Rigid Dynamics

FEA modelling results are only as accurate as the input forces being simulated and the way

these forces are applied. Therefore, the results must be verified analytically or

experimentally for complex geometries. Errors in modelling, typically occur because of

incorrect contacts or joints, poor meshing or poor load sequencing.

Fatigue testing is another useful tool in ANSYS that allows the life of a component to be

determined based on a defined a load sequence. It is the aim of this project to gain a better

understanding of the effect a load will have on the different components when taking into

account the tyre and shock absorber. Using this knowledge, an ANSYS model can be used to

predict the fatigue life of a component, as currently the expected worst case scenario is being

loaded cyclically to determine fatigue life.


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2.1.1 Past JCU Motorsport Analysis

Work done by Crouch, [3] in Design and Simulation of Formula S.A.E Car Suspension in

2011 analysed the dynamics of the vehicle under different load cases. It was determined that

the forces at the contact patch on the tyre could be determined by applying the acceleration

force at the centre of gravity. Crouch simulated the full car under acceleration forces taking

into account the shock absorber effect on the system by using a spring between the bottom

A-Arm and the body (Figure 4).

Figure 4: Transient Structural Analysis using Spring, for Un-sprung Mass [3]

Crouch’s [3] method was also utilised in the upright analysis by Hunter et al.[1] to determine

the forces on the tyre contact patch using a point mass study shown in Figure 5 (without

shock absorber). This allowed unknown forces to be determined, that could then be

reapplied to the wheel, upright and hub shaft. Analysis showed that the most critical load

experienced by a FSAE vehicle’s un-sprung mass is the combination loading of braking,

cornering and bump [1]. The force components from this load case are shown in Table 1.

Figure 5: Point Mass Study [1]


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Table 1: Deceleration (1.5G), Cornering (1.4) and Bump Force (1G) Load Case [1]

Component Location [m] Force [N]

X-component 0 2695

Y- component -0.254 3653

Z-component 0.11 2531

The location components are based on the outer surface of the hub shaft (Hub Shaft and Wheel

Connection) being the origin.

Forces shown in Table 1 are at the contact patch between the tyre and the road surface with

0° camber. The force components shown in Table 1 were then applied in a separate analysis

on just the wheel, hub shaft and upright shown in Figure 6. Joint probes were used to

determine force on bolt holes and bearing housing, which could be used to simply refine the

upright design. The final upright design shown in Figure 7 was further verified in a transient

analysis with the wheel and hub shaft.

Figure 6: Remote Force Application to Wheel, Hub Shaft and Upright [1]


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Figure 7: Upright 2011 Design [1]

In the work of Hunter et al.[1], the forces determined in the computational analysis do not

take into consideration suspension effects, effectively simulating when the vehicle has

bottomed out (when the suspension has no more travel). It is not fully understood if the car

will be able to achieve this load case before sliding.

The 2011 Hub Shaft design analysis by Ikealumba et al.[2],showed the initial design did not

satisfy infinite life under its determined load cases, however, a future design was

recommended from the report shown in Figure 8.

Figure 8: Exploded View of Recommended 2011 Hub Shaft Design [2]

The most critically loaded hub shaft on the car is the rear hub shaft assembly as it is loaded

via torque from the engine and torque from braking in addition to the acceleration and

cornering [1, 2]. The forces a shown in Table 2 and Table 3 and will be used for the

modified design of the hub shaft assembly.


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Table 2: Additional Rear Hub Shaft Forces [2]

Load Load

Maximum Engine Torque

Applied to One Output Shaft

of the Differential.

485.55 Nm [8] (Clockwise

Direction) (75% of 647.4 Nm)

242.78 Nm at a single wheel

maximum

Table 3: Acceleration (1G), Cornering (1.4) Load Case [1]

Component Location [m] Force [N]

X-component 0 -2439.9

Y- component -.254 1823.6

Z-component 0.11 1806.7

The location components are based on the outer surface of the hub shaft (Hub Shaft and Wheel

Connection) being the origin.

Another upright load case (load case 2) that was modelled by Hunter et al.[1], was the

reaction forces at the bolt connection points. These forces (Table 4) were determined using

joint probes at the connection points under the loads shown in Table 1 [1]. The modelling of

these two separate techniques, analyse the forces on the bearing housing via a remote force

at the contact patch, and at the upright connection points.

Table 4: Connection Reaction Forces (Hunter et al. [1]) Load Case 2

Force Components [N]

x y z

Lower A-Arm Bolt 7818 508.22 3378.4

Upper A-Arm Bolt 3543 0 3586.9

Steering Bolt 0 0 3643.5


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Upper Brake Bolt 2094 454.34 36.185

Lower Brake Bolt 262.51 1555.1 790.1

2.1.2 Other FSAE Designs

Other FSAE designs were reviewed to investigate the forces and methods used to design

suspension components such as the uprights and hub shafts. From this investigation it was

found that the majority of the forces used in the design of suspension components had been

developed from track data.

Jawad et al. [9] in the Design of Formula SAE Suspension, used FEA to verify upright

designs to a 1.3G cornering force. These forces were 895N at the upper A-Arm joint, -

2304.7N at the lower A-Arm joint.1121.1N force was also applied at the brake caliper

mounting holes. For the hub shaft, spindle designs were analysed under 1121N and 1445N

applied at the contact patch of the tyre.

Wong [10] in the Design and Optimisation of Upright Assemblies for Formula SAE Race

car also used FEA modelling to analyse stress on the upright design. The model forces were

applied to simulate a cornering scenario, however the forces were applied through the tyre

contact patch. A lateral force (y-axis) of 400lbf (1779.3N) along with a combined bump and

lateral 800lbf (3558.58N) (z-axis). The model was constrained at the upper and lower

upright bolt holes as they would by the A-Arms. The forces for this model were based on

data collected from previous vehicle testing.

McCune et al.[11] in the Formula SAE Interchangeable Independent Rear Suspension

Design report under a similar load case to Hunter et al.[1] in terms of acceleration loads.

McMune et al. simulated the vehicle under 1G of longitudinal, 1.5G lateral and 3G vertical

acceleration. From this simulation forces at the connection point on the upright were

determined using Matlab software from track data. These forces are shown in Table 5.

Table 5: Upright Forces (McMune et al. (2009) [11]) Based on Figure 9

Point X Component Y Component Z Component

A 220lbf (978.6N) 302lbf (1343.3N) -41lbf (182.4N)

B 162lbf (720.6N) -290lbf (-1289.92N) 13.3lbf (59.2N)

C 361lbf (1605.7N) -408lbf (-1814.8N) 700lbf (3113.6N)


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Figure 9: Final Upright FEA Model (McMune et al [11])

Overall the forces in Table 1 and Table 3 by Hunter et al. [1] are conservative in comparison

to the forces shown in Table 5 developed by McMune et al. [11], Wong [10] and Jawad et

al. (2002) [9]. This may be due to differences in geometry or the way these forces were

developed, ie. from computational experimentation or vehicle data.

The FEA modelling techniques used to verify the upright design varied between applying

the forces at the connection points of the upright or applying the loads through the tyre

contact patch. Hunter’s et al. methodology [1] involved the use of both of these techniques

in the design of the 2011 upright to a critical load case of deceleration (1.5G), cornering

(1.4) and bump force (1G). It is proposed that this is a more accurate technique as loads at

the bearing housing and the bolt connections are analysed.

2.2 Vehicle Behaviour Inputs

Vehicle behaviour is an important consideration for testing and computational analysis, as

the test apparatus should reflect realistic loads that are experienced by the vehicle. However,

there are many variables that depict the way a vehicle will behave.

There are three basic types of inertia that have an effect on the dynamic behaviour of a

vehicle. The pitch refers to the inertia effects of a vehicle under acceleration or deceleration,

forces occur about the horizontal axis. Roll refers to cornering inertia and act about the

lateral axis, and yaw refers to spinning effects that occur around the vertical axis (Figure 10).

There is also a polar moment of inertia, which is used to predict how a vehicle will corner

and refers to how the mass of the car is concentrated [12]. For example, a vehicle that has


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light mass concentration close to the centre of gravity will handle better than a vehicle that

has heavy mass concentration distanced from the centre of gravity [12].

Figure 10: Vehicle Pitch, Roll and Yaw [13]

Friction also plays a large roll in the vehicles behaviour during cornering, as without grip the

tyre will slide. If the forces created by the car in any combination exceed the friction limit,

the vehicle will lose traction and the driver will lose control of the vehicle [12, 14]. A rolling

tyre will provide more traction than a spinning or locked tyre.

Camber refers to the angle the tyre makes with the road surface (Figure 11). Negative

camber is when the top of the tyre is closer to the body of the vehicle than the bottom.

Negative camber is preferred for racing as when the vehicle corners, theoretically more of

the tyre will come in contact with the road. Zero camber provides the best amount of traction

and most efficient slip angles [14].

Figure 11: Wheel Camber [15]


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2.3 Suspension Test Methods

Through the advent of computational analysis techniques, the ability to design complex

structures has become more efficient. However, this form of design requires simplifications

and assumptions or estimations of loading forces. Therefore the design must be verified

using different testing techniques.

Handling is an important characteristic in relation to a car’s behaviour on the road. A

passenger or ‘road’ car’s suspension is mainly designed in terms of comfort, whereas for

motorsport vehicles such as Formula 1 or FSAE race cars, the suspension is tuned in terms

of performance (although performance ‘road’ cars are increasingly common place). In either

case, it is important that the suspension be reliable and correctly designed for its particular

field of operation. To verify life and strength of suspension parts, there is currently several

different forms of laboratory testing that are used.

There are many different forms of vehicle testing developed to test and measure different

criteria such as driver comfort through to suspension durability [16]. One form of testing is

4-poster test system from MOOG [16]( Figure 12) used to evaluate ride quality, Noise,

Vibration and Harshness (NVH) and Buzz, Squeak and Rattle (BSR), as well as durability of

chassis and suspension performance. Each tyre of the fully assembled vehicle is placed on a

hydraulic actuator and the actuators are moved up and down out of sync through different

amplitudes and frequencies. This method of testing allows natural frequencies of car

components to be evaluated and suspension system tuned to keep the tyre in contact with the

road surface for as long as possible. However, this form of testing requires an entire car set

up and does not have the capacity to test multi-axis loads or cornering load cases.

Figure 12: MOOG 4-poster suspension tester [17]

Another form of suspension testing is multi-DOF test rigs [16]. These systems are used to

simulate the dynamics of the system, durability and fatigue life. With the additional axes of

movement it is capable of testing vehicle behaviour during more advanced load cases such

as drive torque, braking and steering [18]. Shown in Figure 13 is a 16-DOF suspension test


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rig, which is capable of testing the entire chassis and suspension of a car. This would require

large amounts of computing power to calculate forces and direction, and to control the entire

system. One major advantage of this suspension testing method is that it is capable of

simulating body roll and elevation changes as the body of the car body is not fixed, only the

tyres are.

Figure 13: MOOG 16-DOF suspension test rig [18]

A fatigue based suspension testing method is a Belgium Road experimental test [19]. Used

to test the durability of a component, the testing apparatus is used to exaggerate different

load cases, greatly reducing the time require to achieve failure. Using this method, exact

load cases must be known and applied realistically to achieve an accurate result. The load

must also be applied dynamically if testing durability. This is a good method for verifying

computational fatigue analysis.

2.3.1 Formula 1 Honda Racing Team Suspension Testing

Proof testing and uniaxial testing may be carried out using a universal test frame shown in

Figure 14 to load the component simplistically [6]. Proof testing is the application of a load

that is lower than the design load but higher than its required operational load, of which it is

expected to withstand during racing [5]. Components are proof tested in tension and

compression with a safety factor of 1.3, critical components such as the push rods may also

be tested to a higher safety factor [5].

Proof testing is done to ensure that the component’s design and fabrication is adequate for

race conditions. This is done by applying the maximum load cases that usually occur during

braking at the car’s natural resonance frequency [6]. This method of testing is conservative,

as infrequent loads are being loaded frequently, however the loads applied are from data

collected from racing and offer an accurate indication of a components life [6]. Before the

component will be used on the track it undergoes 100,000 load cycles in a uniaxial


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suspension test, followed by multi-axis test to ensure durability. This allows confidence in

the parts during track testing and more time for performance tuning.

Figure 14: Uniaxial suspension test [7]

Service load testing or durability testing requires much more expensive and comprehensive

testing using multi-axis test rigs [6]. Multi-Axis test systems [5-7] are another form of

testing used and partly developed by the Honda Formula 1 Race team. The forces taken into

account are shown in Figure 15, in their corresponding axes. These forces are determined

from data collected from track testing and race data [6].

Figure 15: Diagram of Forces Acting on a Wheel [7]

The loads are applied to a “dummy” wheel from the hydraulic actuators using light weight,

low friction bell cranks and push rods. The dummy wheel is made to be as stiff and rigid as

possible, whilst still maintaining the same overall weight [7]. This ensures realistic

behaviour of the suspension geometry and is done to simulate the tyre when rotating at high

speeds. A stiffer “dummy” tyre may be used as the inertia of the tyre when rotating at high


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speeds will cause the tyre to become stiffer in the sense that it would become less

deformable.

Data collected from track testing and races are recorded using transducers located on the car

and record different parameters such as acceleration, velocity and displacement [7]. These

loads are software manipulated and then replayed on test components on the 6 axis test

apparatus using servo hydraulic actuators [6, 7]. Maximum forces on the suspension

geometry occur during cornering and braking at high amplitude and low frequency and these

forces can be represented in a vertical and longitudinal components shown in Figure 15 [7].

These forces can be simulated using hydraulic actuators in the vertical and longitudinal

planes and are applied at the tyre contact patch.

During cornering, many forces and reactions of a car during cornering depend on the other

side of the car. Using the test 12 axis suspension test apparatus (6 axis’s per side) shown in

Figure 16, forces can be applied to the suspension geometry realistic manner.

Figure 16: 12-Axis Suspension Testing [5]

To simulate forces in a 12 axis suspension test rig, the front or rear of the car are restrained

to a stiffer and fixed wall [5-7]. This can only be done on the assumption that the inertia or

body roll forces of the body have negligible effect on the suspension geometry behaviour.

However the forces recorded from transducers on the car (recorded during race conditions

and used as inputs) will have these forces taken into account. It must also be noted that the

forces cannot be directly measured at the tyre patch; the forces at the tyre patch for the

simulation are a relation to the forces recorded on other components.

During laboratory testing using the 12 axis suspension test rig data, is also collected from 64

transducers on the suspension geometry [7]. By data logging during laboratory testing, the


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testing technique can be continuously verified with track data. Allowing changes to be made

to the suspension setup and evaluated without running the car on the track.

A 12 axis suspension test rig is not an idealistic method of simulating car behaviour during

cornering because the test rig is only capable of applying forces to the tyre contact patch.

The current rig is not capable of simulating vertical and rotational movements of the

vehicles body caused from inertia or track elevation change. However these simplifications

in formula 1 racing is valid as the car’s suspension is much stiffer and body roll effects are

restrained from large aerodynamic down force effects acting on the car during a corner.

The aim of this research is to provide a simplified version of this type of stand-alone

suspension test apparatus for the use in the design of JCU FSAE car suspension.

2.4 Data Acquisition

Data acquisition is important to allow for efficient and accurate testing to find the most

optimum design and validate components. In formula 1, data is continuously collected and

monitored during the race, during track testing and during laboratory testing [6, 7]. There are

many different techniques and sensors that can be utilised to gather the data required for

experimental testing or fatigue prediction. The strain gauge is simple yet useful method of

determining minimum and maximum forces on a component and load frequency [20].

The strain gauge shown (Figure 17) as the name suggests is used to determine experimental

strain that can be used to calculate stress and force. A strain gauge is a resistive transducer

that is excited by a voltage, when there is a change of the overall length of the strain wire,

there is a change in resistance and therefore a change in voltage [20]. The strain determined

can be converted to stress using Hooke’s Law shown in Equation 1.

Figure 17: Strain gauge diagram [20]


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(1)

Strain results from experimentation can be compared to strain results in FEA models to

verify the computational model. This is done through a direct comparison between the

model and experimental strain results [21-23]. Strains at specific locations can be used to

develop a strain history that can be used to improve the accuracy of computational fatigue

life [24].

2.5 Fatigue Life Calculation

Fatigue is a major failure mode found in repetitive loading situations. Fatigue life of a

component can only be accurate if there is accurate data supporting it. Data acquisition

allows for load cases to be developed and an accurate fatigue life can be determined.

Without the input data, load forces can be inferred from theory & calculated using

computational analysis, however these forces must be verified using experimental testing to

allow confidence in the result.

There are several different methods of fatigue testing and prediction, such as durability

experimentation or through the use of FEA based software. One method that has been used

in failure analysis is to carry out experimental loading until fatigue cracks occur, thus

illustrating the areas of highest stress and using FEA analysis to verify the results [25, 26].

In this method, no direct data was used from the experiments to verify the FEA simulation.

Fatigue life was calculated using an S-N curve from stresses calculated in the FEA analysis

[25, 26]. This method does not provide a full verification of the FEA model, as forces from

the experimental analysis are not collected and compared to the model. However, fatigue life

cycles in the experiment and computational model can be compared.

A more accurate method of verification is to use experimental data to verify the FEA model

[21, 22], as computational modelling is typically a more cost effective method of testing

than many experiments and reduces time to develop a product. The experimental analysis is

conducted under realistic loading conditions and data on forces experienced by different

components is collected. This is then compared to FEA model results. This method verifies

the accuracy of the computational model and allows confidence in the computational results.

This is the approach missing in the JCU FSAE team design and is the main purpose for the

test apparatus.

2.6 FSAE Specific Load Cases

The tilt test and the skid pan are two events in the FSAE competition that are directly related

to suspension testing. The tilt test is used to test roll over stability as the car is tilted to 60°


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which represents 1.7 G’s lateral acceleration. The skid pad is used to test suspension

performance by completing the track shown in Figure 18. The suspension must also be

designed to have 50.8 mm suspension travel, which leads to a soft suspension setup with

small amounts of body roll. For specific FSAE rules relating to the suspension, refer to

Appendix A4[27]. The timed laps and endurance races of the event will also test the

durability and reliability of the suspension.

Figure 18: FSAE skip pan event

2.7 Conclusion

The 12-axis suspension test apparatus is an accurate method for simulation of forces on a

Formula 1 car. It enables forces acquired from track and race to be virtually replayed on the

front or rear portion of the car to verify components or tune performance parameters. The

main assumption of this method is that horizontal inertia effects such as roll, are

insignificant compared to normal forces. This may be an accurate assumption in Formula 1

due to the large amounts of grip that can be generated from larger tyre, stiffer suspension,

heavier vehicle and down force. However, in FSAE, this may not be the case as the

suspension set up is required to be softer, and normal forces down on the car are not as large.

There are other requirements of having a 12 axis testing method such as 12 hydraulic

actuators to be controlled via sophisticated computer systems. This is expensive and not

realistic for the FSAE competition. However, the verification method applied in Formula 1

is an accurate way of ensuring that the designs are safe and efficient. A simplified version

of this testing method will be investigated in the rest of this report.


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3 Design Methodology

The objective of this project was to develop a testing and verification method for future JCU

Motorsport design in relation to the suspension geometry. Currently, as the JCU Motorsports

team is building a first generation race car, there is no method for physically validating

designs. Components are being designed based on theoretical techniques or computational

simulation, or are being fabricated and assembled on the car without prior testing. This

method of design is not ideal as many assumptions are made (For example FEA modelling

constraints) that may lead to over conservative or unsafe designs.

The implementation of a laboratory testing technique will allow for experimentation to

verify assumptions made during the design process in computational models or loads. This

can be achieved by comparing computational results, against data collected from laboratory

experiments. By having data to compare computational results to, the accuracy of modelling

techniques is increased or validated, avoids designs from being overly conservative.

Computational models or simulations can typically be compared to theoretical calculations

to ensure the results that are obtained are realistic. However, due to the complexity of the

suspension system, it is hard to calculate values with theory alone. Using strain data from

experiments is another method that may be used to verify computational models. In this

project, computational strain was compared against experimental strain data to validate

contact, joint and load application assumptions.

With this new information on loads, other components in the suspension geometry such as

the uprights and hub shafts will be redesigned. The aim was to achieve a more complete and

accurate solution for these components.

3.1 Test Apparatus Design

The suspension apparatus is ultimately being designed to allow strain data on critical

suspension components to be recorded under different load cases for the verification of

suspension design methods and assumptions. From the designs investigated in Section 2 the

main limitation of the suspension test apparatus will be the application of force. Due to cost

limitations it is not viable to use hydraulics to apply forces. Therefore, a more simplistic

method was investigated to allow forces to be applied to the suspension geometry.

There are many forces that act on the suspension geometry and affect the dynamics of the

system. As discussed in Section 2, the Formula 1 Honda Racing Team have recognised six

main forces (Figure 15) that affect the suspension geometry. These forces are:


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Longitudinal forces created from friction force on the tyre,

Vertical forces created from bumps or elevation changes in the track,

Steering forces due to the rotation of the wheel,

Braking forces from de-acceleration creating inertial weight change about the pitch

axis,

Camber forces created from the varying contact patch when cornering. This changes

the magnitude of the forces and the direction they are applied,

Cornering forces due to the inertial weight change about the roll axis.

While a test apparatus that investigates all of these forces would be ideal, such a setup would

require a complex dynamic system similar to the 12 axis suspension test rig [5]. Instead, the

scope of the suspension test apparatus designed was limited to at testing cornering forces

directly, with other forces such as braking, camber and steering forces tested indirectly.

Design criteria were developed to ensure that the final design was adequate to achieve the

project goals. The final design of the suspension apparatus had to satisfy design criteria

shown in Table 6.


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Table 6: Apparatus Design Criteria

Criteria Reasoning

Allow for different suspension pick up

points

This allows future generation suspension

design to be tested on this apparatus. This

also for small changes in design to be tested

to tune different parameters such as camber.

Forces applied by weights or pulleys

Achievable in this project with time and

budget restraints. Reduces complexity of the

system without hydraulics and control

systems.

Allow full suspension system to be

attached

By incorporating as much of the suspension

system as possible the more realistic the

testing will be. The full suspension geometry

is shown in Figure 2.

Capable of testing forces on contact patch

of approximately 4000N

Maximum vertical force applied in previous

computational design 3653 N. The apparatus

must be able to achieve this force.

Encourages correct suspension behaviour

Moves in a way that will that simulates as

best as possible the way the entire car system

will.

Cost and Manufacturability The final design of the apparatus must be

cost efficient and capable of being

manufactured in the JCU Mechanical

Workshop.

Conceptual designs were developed with the main philosophy of the design being able to use

rotation effects to simulate body roll effects during cornering. Prototype 1 design (Figure


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19), utilised a plate hinged at the bottom to connect the suspension geometry to. However,

it’s design did not allow easy application of weights or connection of the shock absorber.

Figure 19: Test Apparatus Prototype 1

Test apparatus prototype 2 (Figure 20) changed the point of rotation to a more realistic

location, as on Omega. This design allowed for flexibility in suspension geometry

connection however, due to the box frame being made of Rectangular Hollow Sections

(RHS) the time required for fabrication was higher than prototype 1. Test apparatus

prototype 3 (Figure 21) shows a similar design as prototype 2, with the addition of rocker

and shock absorber mounts and weight application method. Prototype 3 was designed to be

made from metal plate which reduced manufacturing time as only 4 sheets were required to

be cut rather than the 12 RHS sections in prototype 2.


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The prototype designs were compared in a design matrix (Table 7) against the required

design criteria (Table 6) to determine the best apparatus design.


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Table 7: Suspension Apparatus Design Matrix

Criteria Weighting Prototype

1

Prototype

2

Prototype

3

Capable of different suspension

pick-up points Yes/No Yes Yes Yes

Capable of applying forces by

weights or pulleys. Yes/No Yes Yes Yes

Capable of full suspension system

to be attached 0.4 5 5 9

Capable of testing forces on contact

patch of approximately 4000N 0.2 5 7 8

Encourages correct suspension

behaviour 0.2 8 8 8

Cost and manufacturability 0.2 9 5 6

Total Score 6.4 6 8

From the design matrix, (Table 7) the best design based on criteria weighting was prototype

3. From this, prototype 3 will be further analysed and verified.

3.2 Detailed Apparatus Design

The final design was refined by prioritising ease of manufacture, by reducing the complexity

of the components and creating the apparatus geometry similar to the actual car, Omega. The

final design of the apparatus is shown below in Figure 22 with suspension components

attached. Figure 23 shows a diagram labelling main components.


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Figure 22: Final Apparatus Design with Suspension Geometry

Figure 23: Suspension Test Apparatus Diagram

Force application on the weight arm is designed to rotate the apparatus about the hinge

point, effectively loading the suspension system like would result from body roll on Omega.

This is due to outward movement of the top A-Arm and inward movement of the bottom A-

Arm, which will apply pressure to the tyre contact patch. By having the suspension

geometry attached as in Figure 22, the outside tyre during a corner is simulated. The inside


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tyre can be simulated by placing suspension components on the opposite side of the

apparatus.

With the use of the weight arm, mechanical advantage can be utilised to reduce the amount

of weight needed to be applied whilst still achieving the desired load at the tyre contact

patch. The weight application point is located one metre from the hinge point and the tyre

contact patch is 0.658m. Therefore, if the required force at the contact patch is 3653 N

(Table 1, Y force component) only 2404 N (245kg) of force is to be applied on the weight

arm. For detailed calculations on the force to be applied, refer to Appendix A6.1.

The width of the apparatus was chosen to replicate the distance between the front suspension

mounts (485mm). The hinge point height from the ground was chosen to be approximately

the same height as the roll centre of the car when completed (200mm). Both of these

measurements encourage the suspension to behave as it would on Omega, to allow for

accurate testing.

The apparatus design allows for flexibility in the attachment points of the suspension

assembly being tested, so future suspension assembly designs can be tested using this

apparatus. The suspension apparatus is capable of testing front or rear suspension, inside or

outside tyres.

The main limitation of the apparatus is that it is not loaded dynamically and the apparatus

cannot test fatigue. Another limitation of the apparatus is that 245kg of weight is required to

apply the maximum load. However, further modifications beyond the scope of this project

may allow pulleys or hydraulics to be added to the apparatus for easier application of force.

Force for initial verification experiments will be approximately 50kg as this is a small

weight that can easily be worked with.

The weight of the apparatus is not matched to the weight of the car that it is representing.

This is not expected to have an effect on the results as the weight being applied (roll force) is

much larger than this weight. Bearings were not used at the hinge point as this would add

cost and complexity to the design, which was not deemed warranted.

MS 250 steel was chosen for the fabrication of the apparatus due to its low cost and

workability, while still providing adequate strength characteristics for the application. MS

250 steel has similar properties to structural steel in ANSYS, therefore the structural steel

properties will be used for computational analysis.


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3.3 Analysis and Verification of Apparatus

The apparatus design was verified in ANSYS to ensure that the design was sufficient to

allow the correct forces to be loaded. The loads specified simulate a worst case situation that

will occur when the suspension has “bottomed out”. For this case, the bottom of the base and

the A-Arm brackets were connected to ground using a body to ground joint. All other

contacts were set up as bonded contacts. The generic mesh was used for the majority of the

apparatus, with smaller mesh sizing of 0.01m used for areas where high stress was likely to

occur, such as the weight lever arm.

Figure 24: ANSYS Apparatus Meshing

The forces applied on the apparatus are based on the Y-component force shown in Table 1.

The forces were loaded on the apparatus as shown in Figure 25, with force B being the

weight force required. Force A is the total Y-component force that may be loaded on to the

rocker connection point.


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Figure 25: Force Loading on Apparatus

The static safety factor for this load case is 2, shown in Figure 26. The high stress areas

between the apparatus and weight arm shown in Figure 27, are due to the method of which

the components are bonded. The assumption was made that the bolts will be tightened to 90

% of proof load, effectively bonding the apparatus and weight arm together.

Figure 26: Apparatus Safety Factor


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Figure 27: Apparatus High Stress Areas

It is important that the weight arm does not deflect a large amount as this may affect the

loading and accuracy of the results. The maximum deflection of the apparatus under this

load case was 0.002 m which is acceptable for this application (Figure 28). Fatigue life of

the apparatus was not tested as it will not be dynamically loaded and is not expected to fail

under fatigue.

Figure 28: Deflection of Apparatus

Further analysis was carried out on the weight arm, as it is a critical component of the test

apparatus and the bolt holes will create stress concentration. Due to the positioning of the

bolts and the loading of the weight arm, the bolt holes experience large variations in stress

demonstrated in Figure 29. The weight arm was loaded in the same way as shown in Figure

25 with 2404 N at the load point and the bolt holes fixed using a body to ground joint. This

simulates a worst case scenario where the suspension has “bottomed out”. A body mesh of

0.01 m elements was used.


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Figure 29: Weight Arm Stress Distribution

The deformation shown in Figure 30 was compared to Figure 28 to verify the weight arm

model results against the results of the apparatus analysis. A maximum deformation

comparison between Figure 28 and Figure 30 results in a 6.3% difference, which is

generated due to the difference in support methods. In both models the minimum static

safety factor is approximately 2 shown in Figure 26 and Figure 31, which is a suitable safety

factor for this application.

Figure 30: Weight Arm Total Deformation


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Figure 31: Weight Arm Safety Factor

In the weight arm model joint probes were used at the bolt holes to determine the resultant

forces on the bolts, these forces are shown in Table 8 The axis positioning for the force

components is shown in Figure 32 on the left hand side. The coordinate system on the right

hand side of the figure is the model coordinate system.

Figure 32: Axis Positioning


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Table 8: Bolt Joint Probe Results

(Bolt positioning as shown in Figure 29, Figure 30 and Figure 31)

From the forces shown in Table 8, bolt B has the largest resultant force and is therefore the

most critical. To ensure that an M12x1.75 grade 8.8 bolt is adequate for the application,

calculations were performed in Excel. The bolts in this application are loaded in shear and to

decrease shear loading the bolts will be tightened to 90 % load. This increases the area the

surfaces are in contact and frictional forces that must be overcome, effectively bonding them

together. This being the case, the static axial safety factor is 3.23 and the shear safety factor

is 4.53. For more details on calculations refer to Appendix A7.

From the analysis done, the apparatus design is deemed to be adequate for the application.

For full fabrication design drawings refer to Appendix A1. In order to verify that the

designed apparatus works, an experiment was carried out to record strain data at several

different weight applications. This was done to verify that the use of strain gauges would

indeed give results proportional to the forces being applied to the weight arm.

Force

Component

Bolt Force [N]

Bolt A (Top

Left)

Bolt B (Top

Right)

Bolt C (Bottom

Left)

Bolt D (Bottom

Right)

X – Component 282.95 9044.8 -988.3 -8339.5

Y-Component 38.664 -103.41 612.23 -2951.5

Z-Component 0 0 0 0

Total 285.58 9045.4 1162.6 8846.3


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4 Experimental Analysis

Apparatus experimental testing with the manufactured suspension test apparatus was

undertaken to record data on critical suspension components. Strain gauge measurements

were used to determine stress at critical locations on the components, with values used for

verification of simulation approaches.

4.1 Experimental Apparatus Setup Procedure

Below is the procedure for setup of the suspension test apparatus including the attachment of

the strain gauges to suspension components.

1. Attach strain gauges to critical components using to information below,

a. Clean surface thoroughly of contaminates using degreaser.

b. Using wet and dry sand paper of varying grade, sand the area where the

strain gauge is to be placed smooth. Ensuring that there is no paint left on

the surface and only smooth metal remains. The use of water during this

process will keep the sand paper from clogging up with debris.

c. Clean surface and dry thoroughly.

d. Mark on surface using a fine permanent marker the centre positioning of the

strain gauge.

e. Clean a Perspex surface thoroughly and lay a strain gauge down on the

clean surface (The side where the wires are soldered on facing up). Ensure

the strain gauge does not come into contact with skin.

f. Place a piece of clear sticky tape that is longer than the strain gauge over the

strain gauge so that the strain gauge sticks to it.

g. Position the strain gauge on the component using sticky tape to hold it in

place.

h. Slowly peel the sticky tape back so that the strain gauge is still attached to

the sticky tape, leaving one end of the tape steel attached to the component.

i. Apply super glue to the area where the strain gauge is to be attached.

j. Pull the sticky tape back down utilising the sticky tape to hold the strain

gauge in place.

k. Once the super glue is dry, peel the sticky tape off the strain gauge.

l. Repeat steps a. to k. for strain gauges on other components.

m. Mix a batch of araldite epoxy as per packet instructions.


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n. Place araldite epoxy over strain gauges so that the strain gauge is covered

for protection and allow to dry.

2. Attach suspension sub assembly using correct fasteners and tools as per the Bill of

Materials (BOM) shown Table 9, Figure 33 and Table 10, Figure 34.

Table 9: Hub Shaft and Upright Geometry BOM [1]

Item Number Component

1 Upright

2 Bearing

3 Brake Caliper*

4 M8x30 Bolt

5 M10x40 Bolt

6 M10-W-N Nut

7 Oil Seal

8 Hub shaft

9 Hub Washer

10 Stub Washer

11 Rotor Pin Cir clip*

12 Stub Shaft

13 Brake Rotor Button*

14 Brake Rotor*


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*The brake rotor and brake calliper do not have to be attached for experiments.

Figure 33: Hub Shaft and Upright Geometry Exploded View [1]


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Table 10: Suspension Assembly BOM

Figure 34: Exploded View of Suspension Assembly


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4.2 Apparatus Experimental Procedure

Before starting experimentation, the suspension settings were set as desired on the shock

absorber. The settings that were used for initial verification testing were:

Spring Length 113 mm on the 350 x 2.45 spring

Volume Adjust Knob set at 2

Compression Dampening

o Red Knob set to least

o Black Knob set to least

Experiment Procedure

1. Turn on Data Acquisition System.

2. Attach strain gauge wires to input channels on the Data Acquisition System.

3. Program data acquisition system with gauge factor coefficients.

4. Apply specified load case to the end of the apparatus weight arm. Apply weights in

stages until goal weight is reached ensuring body parts are away from moving

components and correct lifting procedures.

5. Record strain values.

6. Repeat Steps 4 and 5 for other loads.

Important Note: The weights must be loaded slowly in small increments to represent a

quasi-static nature of loading. This reduces the risk of catastrophic failure of a test

component that may cause injury.

4.2.1 Risk Assessment

A completed and signed risk assessment for this thesis has been under taken and is attached

(Appendix A8). The main risks involved are with the experimentation, however, using the

appropriate techniques and strategies, the risks can be greatly reduced. Appropriate PPE, and

safe working techniques must be employed to reduce the risks involved. A safety induction

for the test area has already been undertaken.

The major risk with this experiment is the failure of a component under the large loads

applied. The design of components has been done with safety in mind to minimise this risk,

however, caution must be used when loading the apparatus. Using correct loading technique

and visually checking components for over load should be employed to reduce the chance of

catastrophic failure.


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4.3 Apparatus Experimental Results

Once the apparatus was fabricated, initial experimentation was undertaken to verify the

design and record strain data at different weight applications. This was done to prove that

the strain gauges would indeed give results proportional to the forces being applied to the

weight arm. This strain data is also used later in comparison to computational model results.

4.3.1 Apparatus Experimental Results

The surface of the metal was prepared for strain gauges as shown in Figure 35, to a smooth

consistent surface using various grades of wet and dry sandpaper. This was done to ensure a

smooth finish for the strain gauge and the adhesive. The strain gauge was attached using the

experimental procedure defined (Figure 36). For the A-Arms, one strain gauge was placed

half way along each member (Figure 37). This was done as the A-Arm members are

expected to be in bending and that the highest strain or deformation would be at these points

on the top or bottom surfaces.

Figure 35: Metal Finish before Strain Gauges Attached

Figure 36: Completed Strain Gauge


52 | P a g e

Figure 37: Strain Gauge Positioning

For the push rod, four strain gauges were attached in pairs of two on opposite sides of the

push rod. Of the strain gauge pairs, one strain gauge was placed vertically along the shaft

and the other horizontally (Figure 37; Figure 38; Figure 39). This was done at this point half

way along the push rod, as bending, torsion and buckling are possible load cases. Using this

configuration, any form of bending, torsion or buckling, in any direction can be measured.

Figure 39: Strain Gauge Positions on the Push Rod

5

6

7

8

Figure 38: Positioning of Strain Gauge

Locations on Push Rod


53 | P a g e

The completed and assembled apparatus is shown in Figure 40 with all necessary

components in place. However, in place of the upright, hub shaft and wheel is a fabricated

support that has been done to replicate a stiff wheel. This was done as the upright and hub

shaft had not yet been manufactured, so the support was made to replicate a similar

geometry to the upright and wheel assembly. This substitution is a valid as the pressure

inside the tyre and inertia of the wheel spinning would make the tyre rigid. This assumption

is conservative as it assumes that none of force is absorbed by the tyre.

Figure 40: Completed and Assembled Test Apparatus

The apparatus was tested over a range of weights from 20kg to 70kg. A Tokyo Sokki

Kenkyujo Co. Computing Data Logger (TDS-602) data acquisition system situated in the

JCU structural lab was used to read the strain from the gauges. The system needed initial

programing to set gage factor and resistance (Appendix A10). All measurements are

initialised meaning that the initial measurement at zero load is subtracted from all readings

such that values recorded are only the change in strain.

The results from the experimentation are shown in Table 11 over the range of weights

applied. For an unknown reason, strain gauge 8 did not give any reading, however, all other

gauges gave logical readings. Also shown in Table 11 is the change in spring length.


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Table 11: Experimental Results

Weight

[kg]

Force

[N]

Force

at

Contact

Patch

[N]

Strain Gauge με Spring

Length

Bottom A-

Arm

Top A-

Arm

Push Rod 350 x

2.45

1 2 3 4 5 6 7 8

0 0 0 0 0 0 0 0 0 0 0 113

20 196.2 300.5 -79 -89 -14 -3 5 -15 -17 0 109

40 392.4 600.9 -175 -188 -15 -6 12 -34 -27 0 104.5

60 588.6 901.4 -256 -279 -15 -9 19 -52 -40 0 99

70 686.7 1051.6 -327 -349 -15 -8 22 -65 -50 0 97

Note: Values shown in Table 11 were recorded after every increase in weight applied. This was done

as the apparatus was being load statically and there was minimum fluctuation in the values.

The results in Table 11 are shown graphically in Figure 41. Linear line of best fit is shown

for all of the strain gauge locations. For strain gauges 1,2 and 6, the equation for this line is

shown as it can be used to estimate strain at higher load weights.

Figure 41: Strain Result Graph

y = -4.46x

y = -4.8067x

y = -0.8886x

-400

-350

-300

-250

-200

-150

-100

-50

0

50

0 20 40 60 80

Stra

in μ

ε

Weight Applied [kg]

Strain Results

Strain Gauge 1

Strain Guage 2

Strain Guage 3

Strain Gauge 4

Strain Gauge 5

Strain Gauge 6

Strain Gauge 7

Linear (Strain Gauge 1)

Linear (Strain Guage 2)

Linear (Strain Guage 3)



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From Figure 41 the following observations have been made:

As the weight applied increases, the strain on strain gauges 1, 2, 5, 6, and 7

increases.

As expected strain gauges 1 and 2 on the bottom A-Arm are negative as the bottom

A-Arm is loaded under bending, and the top surface of the A-Arm is in

compression.

The top A-Arm does not have a significant change in strain as the weights are

applied.

The experiment was carried out using a rocker ratio between the push rod and shock

absorber of 1.8, and a spring stiffness of 350 lb/ft. The rocker ratio refers to the 1mm

movement in the push rod is 1.8mm movement in the spring. In Figure 42, the spring length

is plotted against the load weight applied. From the trend line, it is determined that the

spring has a linear relationship and that the spring behaves as according to Hookes Law.

Figure 42: Spring Length vs. Load

Based on Hookes Law and the rocker ratio, the force on the push rod can be calculated and

is shown graphically in Figure 43. These forces take into account the 1.8 ratio of spring to

push rod.

y = -4.2x + 117.1

85

90

95

100

105

110

115

0 20 40 60 70

Len

gth

[mm

]

Weight Applied [kg]

Spring Length vs. Load

Spring Length

Linear (Spring Length)


56 | P a g e

Figure 43: Spring Force on Push Rod

4.3.2 Experimental Discussion

The suspension test apparatus is capable of providing experimental data in relation to strain

at critical components, and force exerted by the spring. This data can be utilised in many

ways. The most beneficial result is the forces on components in relation to the force at the

tyre contact patch. This data can be further developed into loads or forces on different

components, which can be used in future designs and in verifying the simulation models in

relation to the strain data.

Temperature is not expected to have affected the values from the strain gauge as the

experiment was conducted over a period of 20 minutes and the change in temperature during

this time is expected to be minimal. The strain gauges are assumed to be connected correctly

and give accurate readings; however, this may not be the case and the results may be

affected if the strain gauges slip on the metal when loaded.

Further experimentation at higher load weights is required to further develop a strain vs. load

equations for the various strain gauge locations. From the strain results the data shows a

linear relationship, however, this is not expected to be the case for higher load weights. Also

for strain gauges 1, 2 and 6, the strain value recorded at a weight of 70kg shows a variation

to previous values. Further testing and results will allow more accurate assumptions from the

results to be made.

In the future, as further components are designed, more strain gauges can be added to the

system to gather more data to be used in further analysis. Once there is a drivable car, strain

gauges on suspension components can be used to validate the behaviour and loading method

0

20

40

60

80

100

120

140

160

0 20 40 60 70

Forc

e [

N]

Weight Applied [kg]

Force on Push Rod

Spring Force


57 | P a g e

of the suspension test apparatus. In the section to follow, computational analysis will be

compared to results from the experiments to validate computational techniques and

assumptions.


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5 Computational Simulation

Computational experimentation has been carried out in ANSYS 13.0 to develop a working

model of the suspension geometry. Results from the computational model will be compared

to the experiment results to validate the model. By doing this comparison, confidence in the

model methods can be verified against physical data. The points for comparison were the

positions where strain gauges were placed on the experimental apparatus.

5.1 Computational Simulation Results

All models in this section were modelled using ANSYS Transient Structural Analyses over a

period of one second, with all forces ramped. All models were solved over one time step,

using a minimum number of sub steps of five and a maximum number of sub steps of ten.

A basic load case (Model A) was modelled to gain an understanding of the behaviour of the

bottom A-Arm and build a foundation for more complex models to follow. The model

simulated the bottom A-Arm and bracket, with the bracket attached to the bottom A-Arm

using a bonded contact. The bracket was fixed to ground using a general body to ground

joint at the bolt holes, fixed in all directions. The other system constraints are shown in

Figure 45 along with the mesh in Figure 45. The sizing of the mesh is constant for all

subsequent models to allow for more accurate comparison.

Figure 44: Model A Setup


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Figure 45: Model A Mesh

The force applied to the end of the A-Arm equals the reaction force that would have

occurred during the experimental testing at a weight of 70kg (1051N at contact patch). This

allows all results to be compared to the 70kg strain results. The strain results from the

computational simulation are in agreement with those from the experimental testing, with

the top of the A-Arm being in negative bending. The strain results from this simulation are

shown in Figure 46, with the strain shown at nodes close to the location of the strain gauges

used in the experiment (blue probes). The green highlighted probes show the location where

the strain is approximately the same as the experimental data. In Table 12 the strain results

from the computational model are compared against the experimental strain data.

Figure 46: Model A Strain Results


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Table 12: Model A, Computational and Experimental Strain Comparison

Strain Gauge Strain με

Error %

Experimental Computational

1 327 272 16.82

2 349 303 13.18

From the error percentages shown in Table 12, the lowest error is 13.18% in comparison to

the 70kg experimental load case. This error is expected to be due to the bracket constraint,

having the bracket fixed at the bolt holes hinders the bending of the A-Arm.

The same model was simulated again including the push rod to see what effect the push rod

had on the strain results (Figure 47) (Model B). The push rod was constrained to the bracket

using a general body to body joint.

Figure 47: Model B Setup and Mesh

The strain results are shown in Figure 48 and the results are compared in Table 13. The error

percentage for this model is much less than Model A. However, the strain for the push rod is

over conservative with a 56.92% error. This may be due to the spring not being included in

the model or the push rod constraints.


61 | P a g e

Figure 48: Model B Strain Results

Table 13: Model B, Computational and Experimental Strain Comparison


Error % Experimental Computational

1 327 304 7.03

2 349 343 1.72

6 65 102 -56.92

Below in Figure 49(a), the exaggerated deformation of the bottom A-Arm and push rod are

shown. The deformation of the bottom A-Arm is as expected with the largest bending

occurring around the stiffness change. However, the deformation for the push rod is not

correct; Figure 49(b). On the apparatus the push rod is connected via tie rods which only

restrain movement in one direction. Using general body to body joints in the model restrains

the push rod in three directions. This caused the push rod to bend in an S shape.


62 | P a g e

Figure 49: Model B Exaggerated Deformation

A more complex model (Model C) was tested including the A-Arm brackets and the stiff

member (Figure 51). In this model only the bottom A-Arm and the push rod were flexible,

and general body to body joints were used to connect the bottom A-Arm to the A-Arm

brackets. The bottom A-Arm bracket was bonded to the A-Arm using a bonded contact as in

the previous models. The A-Arm brackets were connected to the ground using a general

body to ground joint. The push rod was fixed using a fixed support.

Figure 50: Model C Setup

The strain results from this model are shown in Figure 51. The comparison to the

experimental results is shown in Table 14. This model had a consistently large error from the

experimental results, which is a counter intuitive result as the attachment of the stiff member

and brackets creates a larger bending moment. However, the constraints of the joints inhibit

bending of the components in a realistic behaviour.

a b


63 | P a g e

Figure 51: Model C Strain Results

Table 14: Model C, Computational and Experimental Comparison


Error %

Experimental Computational

1 327 114 65.14

2 349 112 67.91

6 65 7.5 88.46

A further model similar to Model C (Figure 52) was simulated instead with tie rods and

contacts used in place of joints as it is expected that the joints limit the movement of the

model (Model D). The tie rods were included in the model as rigid, and a force was applied

at the end of the bottom A-Arm to simulate loading form the stiff member. From the

comparison of the strain results in Table 15, this model had a reduced error at strain gauge

locations 1, 2 and 6 in comparison to Model C.

Figure 52: Model D Setup and Strain Results


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Table 15: Model D, Computational and Experimental Comparison


Error % Experimental Computational

1 327 245 25.08

2 349 261 25.21

6 65 108 -66.15

5.1.1 Further Model Investigation

Model B and Model D were the most accurate models simulated. Further modelling was

undertaken to investigate the accurracy of the model over all of the load weights tested. This

involved further testing at 20, 40 and 60kg load weights. These strain values were then

compared to the respective experimental values (Figure 53, Figure 54 and Figure 55).

The exact values and comparison is shown in

Table 16, Table 17 and Table 18.

Figure 53: Strain Gauge 1, Model B and D Strain vs. Experimental Strain

y = 4.46x

0

50

100

150

200

250

300

350

0 20 40 60 80

Stra

in [

με]

Weight Load [kg]

Strain Gauge 1

Model B

Strain Gauge 1

Model D

Poly. (Model B)


Poly. (Model D)


65 | P a g e

Table 16: Strain Gauge 1, Model B and D vs. Experimental Strain

Weight

Load [kg]

Experimental

Data [με]

Model B

Strain [με]

Model B

Error %

Model D

Strain [με]

Model D

Error %

20 79 57.7 26.96 68.3 13.54

40 175 118 32.57 139 20.57

60 256 178 30.47 209 18.36

70 327 304 7.03 245 25.08


y = 4.8067x

0

50

100

150

200

250

300

350

400

0 20 40 60 80

Stra

in [

με]

Weight Load [kg]

Strain Gauge 2

Model B

Strain Gauge 2

Model D

Poly. (Model B)


Poly. (Model D)


66 | P a g e


Weight

Load [kg]

Experimental

Data [με]

Model B

Strain [με]

Model B

Error %

Model D

Strain [με]

Model D

Error %

20 89 58.9 33.82 68.50 23.03

40 188 121 35.64 135.00 28.19

60 279 178 36.2 214.00 23.30

70 349 343 1.72 261.00 25.21


y = 0.8886x

0

20

40

60

80

100

120

0 20 40 60 80

Stra

in [

με]

Weight Load [kg]

Strain Gauge 6

Model B

Strain Gauge 6

Model D

Poly. (Model B)


Poly. (Model D)


67 | P a g e


Weight

Load [kg]

Experimental

Data [με]

Model B

Strain [με]

Model B

Error %

Model D

Strain [με]

Model D

Error %

20 15 21.3 -42 36.00 -140.00

40 34 39 -14.71 70.00 -105.88

60 52 62.4 -20 102.00 -96.15

70 65 102 -56.92 108.00 -66.15

5.1.2 Further Push Rod Simulation

The push rod deformation in Figure 49 was not as expected. Further investigations were

made into this result (Figure 56). It was found that strain located more on the side of the

push rod. The value was closer to that found from experimental strain gauge 6. This

discrepancy in the results may be due to the positioning of the strain gauges on the push rod

in the experiment. Their positioning may not have been in the same plane as the bending.

Figure 56: Model B Push Rod Result


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A simple push rod was analysed in a separate model using forces determined from a joint

probe between the push rod and bottom A-Arm from Model B. The tensile force from the

joint probe and the maximum spring force (147N Figure 43) were applied to the push rod

using a simple joint constraint at one end. The strain results (Figure 57) correlate with the

results from Model B.

Figure 57: Simple Push Rod Analysis using Realistic Constraints

Basic hand calculations on the push rod using the same tensile force (Appendix A6.2)

determined the strain to be 1.89E-4. All three strain results are compared in Table 19. The

most accurate technique from this comparison was the simple push rod model.

Table 19: Push Rod Modelling Comparison

Technique Strain [με] % Variation to Experimental Strain Value (65με)

Model B 118 81.5

Simple Push Rod 87 33.8

Hand Calculation 123 89


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5.2 Computational Simulation Discussion

From the graphical results for strain gauge 1, 2 and 6 in Figure 53, Figure 54 and Figure 55,

the results of the experimental strain can be approximated by a linear equation. The

computational data for load cases 20, 40 and 60 kg can also be approximated by a linear

equation. This is evident from the relatively consistent error between the computational and

experimental strain for these weight loads (

Table 16, Table 17 and Table 18). The linear properties of the material used in the analysis

are the reason for this linear strain relationship if the same nodes are chosen. Further

computational testing using non-linear material properties may be required if the

experimental strain values cannot be approximated using a linear relationship.

The computational strain results for Model B at weight load 70kg vary to the other load

weights, by either greatly increasing the accuracy or decreasing the accuracy. This may be

due to the wrong strain node being chosen or there being a large change in force between

load weight 60 kg and 70kg. Further investigation is required to determine the reasoning for

this error and a more accurate method for determining the strain at particular nodes found.

Model D is the most consistent in comparison to Model B (Table 16, Table 17 and Table

18). The error for Model D for strain gauges 1 and 2 is approximately 20%, whereas Model

B is approximately 30% with only the 70kg test having much more accurate results. Both

models fail to accurately predict the strain of the push rod. Model D, with the contacts, is

consistently a better modelling method for the bottom A-Arm.

From the comparison in Table 19 the most accurate method of modelling for the push rod is

the use of the simple model. However, this method will not take into account buckling.

Further investigation is required to fully understand the modelling of the push rod. For future

calibration of push rod modelling, tensile testing of the push rod should be undertaken to

develop a force strain relationship.

Errors in the modelling or experimental technique may have caused errors in the results. One

main error that may have caused discrepancies in the computational modelling is the method

by which the strain values were found at specific locations. A probe was used to find the

strain at the desired location; however, this location may have deviated between models or

tests. This is not ideal, as accuracy in this area is required to develop accurate relationships

between experimental and computational models.


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From the modelling in this section it was found, that in terms of connections, joints are

easier to get solutions. However, the solution may not be accurate. Care must be taken in the

constraints used in the joint, to not over constrain the model. Contacts require more effort to

solve but in general, offer a more accurate solution. This was shown in the results by

comparing Model C against Model D.


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6 Upright Verification

As a part of this report the, 2011 Upright design (Figure 58) was modified and verified for a

larger hub shaft. The original upright was designed to house a bearing for a 30mm outside

diameter hub shaft, however, this was found by the 2011 Hub Shaft design team to be

inadequate. Modifications to the design were verified in ANSYS using the load components

shown in Table 1 on the upright, hub shaft and wheel assembly.

Figure 58: 2011 Upright Design

Due to the upright material block of Alumec already being purchased, the upright was

redesigned to fit a bigger bearing, whilst still being able to be machined from the same

Alumec block of 500mm x 250mm x 72mm using the configuration in Figure 59. From the

2011 Upright report, the bearing recommended was a SKF Tapered bearing of ID 40mm and

OD 68mm (32008x). This was the maximum allowable bearing bore as a larger bearing

would require the brake caliper mounts to be moved.


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Figure 59: Upright Machining Configuration

The other main reason for the upright redesign was to make the upright easier to

manufacture. The original design assumed that the CNC milling machine in the JCU

workshop could be used, however this was not the case. Therefore the upright design and

shape was altered to make use of other techniques such as water jet cutting and conventional

milling techniques.

The 2012 upright design (Figure 60) was chosen as it is a more cost effective design and

utilises cheaper forms of manufacturing due to its less complex shape in comparison to the

2011 upright (Figure 58). The 2012 upright is also 3.3mm wider to reduce bending moment

reaction forces from the hub shaft. The new upright also satisfies the same constraints as the

2011 upright and new constraints as previous outlined. For full drawings refer to Appendix

A2. The constraints as outlined by the 2011 Upright Report (Hunter et al.[1]) that have been

satisfied by the 2012 upright design are:

Distance between the top and bottom A-Arm mounting points is 155 mm.


73 | P a g e

The brake caliper mounting points match those of the already purchased Nissan

brake caliper’s as per the Brake Caliper and Rotor Amendment report (Wallace et al.

[28]).

The upright is universal and may be used at any wheel.

For full constraints refer to 2011 Upright Report (Hunter et al. [1]).

Figure 60: 2012 Upright Design

6.1 Detailed Upright Analysis

The 2012 upright was verified in ANSYS 13.0 using the same load cases and forces

developed in the 2011 Upright Report. The material used for the analysis was Alumec, an

aluminium alloy. For relevant material properties refer to Appendix A11.1.

In both load cases, a transient structural method was used to simulate the loading of forces

over a one second time period. All models were solved over one time step, using a minimum

number of sub steps of five and a maximum number of sub steps of ten.

The upright fatigue life was determined based on ANSYS fatigue data for Aluminium. The

life of the upright was analysed to a life of 10E8 cycles as Alumec has different fatigue

properties to steel and this is assumed to be infinite life. In ANSYS fatigue tool, a fatigue

strength factor of 0.7 and zero based loading was used.

Alumec is a 7000 grade of aluminium and has similar properties to 7075 aluminium alloy.

The S-N data used for fatigue analysis is based on fatigue data for 7075 aluminium alloy.


74 | P a g e

This data, shown in Table 20, was based on the experimental data from Military Handbook -

MIL-HDBK-5H pg. 3-382. For more information refer to Appendix 11.2.

Table 20: ANSYS Fatigue Data for Alumec

Cycles Strength [MPa]

1E4 551

1E5 413

1E6 344

1E7 275

1E8 206

Load case 1 analyses the upright and hub shaft together to apply reaction forces to the

bearing housings by applying a remote fore to the outer hub shaft face (Figure 61). A mesh

body sizing of 0.005m was used (Figure 62). The remote force components are the same as

in Table 1. The remote force location is at the road and tyre contact patch. The top and

bottom A-Arm and steering arm connection points are fixed using joints.

Figure 61: Load Case 1 Geometry


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Figure 62: Load Case 1 Mesh

With the upright filleted, the highest stress areas are around the top and bottom A-Arm

connection bolt holes (Figure 63). There is a stress concentration around the bottom of the

upright as shown, however, the critical stress is only 95 MPa resulting in a fatigue safety

factor of 2.4 (Figure 64).


76 | P a g e

Figure 63: Load Case 1 Stress

Figure 64: Load Case 1 Fatigue Safety Factor

Load case 2 analyses the upright under loads applied at the bolt holes. The loads for the bolt

holes are taken from Table 4. The model setup is shown in Figure 65. The loads are applied

to the whole bolt hole using a force and the upright is supported using two cylindrical

supports in place of the bearings. A mesh of increased relevance was used at the bolt holes

to increase the accuracy of the results (Figure 66).

Figure 65: Forces Analysed for Load Case 2


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Figure 66: Load Case 2 Mesh

This load case is more critical than load case 1 as it has a higher maximum stress of 117

MPa located at the top of the upright (Figure 67). It is also evident that much more of the

upright is stressed. This results in a fatigue safety factor of 1.9 as shown in Figure 68.


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Figure 67: Load Case 2 Stress

Figure 68: Load Case 2 Fatigue Safety Factor

6.3 Upright Recommendations

From the ANSYS analysis, the 2012 Upright design is suitable for application on Omega in

FSAE competition as the critical fatigue safety factor from load case 1 and 2 is 1.9.

The uprights are currently being manufactured. The Alumec block has been water jet cut to

the 2D shape of the upright (Figure 69) with milling to follow to machine the bearing

housing, brake caliper mount and top and bottom to the correct thickness. The bolt holes for

the steering arm, A-Arm’s and brake caliper also need to drilled. For the full upright

manufacturing drawing refer to Appendix A2.

Figure 69: 2012 Current Progress of Upright


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7 Hub Shaft Redesign

The hub shaft requires further design before fabrication, as the 2011 design analysis found

that a hub shaft diameter of 30 mm was not sufficient to withstand the required load cases.

This project aimed at redesigning the hub shafts with new load cases based on the load

components shown in Table 1 and Table 2. The aim of the design was to create a hub shaft

design that is cost effective by utilising machining techniques available in the JCU

mechanical workshop. The main material considered for this application is 4140 steel as it

has high strength properties and is widely used for rotating shafts.

The hub shaft is responsible for the transmission of torque from the engine and provides

support for the rotating wheel. A general assembly with other components associated with

the hub shaft are shown in Figure 70.

Figure 70: Hub Shaft General Assembly

The hub shaft design must meet several constraints in order to be used on the Omega car.

These constraints are either outlined in the FSAE Rules or are constraints based on the

components that have already been purchased to fit to the hub shaft. These constraints are:

The hub shaft must bolt on to the wheel by 4 x M12 bolts, equally spaced at a radius

of 50mm.

The constant velocity joint tripod must bolt on to the inner side of the hub shaft

using 6 x M10 bolts equally spaced at a radius of 37mm.

The outer bore of the hub shaft where the bearings are placed may be no greater than

40mm.


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The distance between the outer bearing and the inner surface of the hub shaft wheel

flange must be minimised but no less than 52.5mm. This is due to the positioning of

the brake caliper and its size and will stipulate the positioning of the brake rotor.

The brake rotor must be attachable to the hub shaft via 6 half hole notches at an

outer radius of 45mm. The radius of the holes is 6 mm.

The outer distance between the bearings is 60mm.

The rear hub shaft must be capable of transferring torque to the wheel.

An initial hub shaft design that satisfied these constraints is shown in Figure 71. This design

(Hub Shaft 3.0) was chosen as it a suitable solution to the design constraints, whilst

achieving a reduced complexity in shape to allow for easier fabrication. The hub shaft was

designed to reduce waste of material and will be made up of several parts that will be welded

or fitted together to achieve the overall shape. The method chosen for torque transmission is

a keyway as it is easier to manufacture and is capable of being done at the JCU Mechanical

Workshop. The keyway width and height dimensions conform to AS 2037 – 1977 however

the length of the keyway does not. The key and keyway was analysed to ensure that this is

an appropriate method for this application.

Figure 71: Exploded Conceptual View of Hub Shaft 3.0

Loads on the hub shaft vary between the front and rear of the car. The most critical load case

for the hub shaft on the front of the car is under braking, bump and cornering when the car

Bending Diameter

(Spindle)

CV Section

Hub Shaft Outer

Flange

Brake Rotor

Attachment


81 | P a g e

lunges forward. For this load case the forces are the same as those developed in the 2011

Upright report (Hunter et al.[1]). The critical load case for the rear hub shaft is under

acceleration, cornering and applying maximum torque. The hub shaft will be analysed under

two load cases to ensure that it is adequately designed for application in FSAE. These load

cases are:

Load Case A: Brake lock up whilst cornering and a bump which is when the brake

rotor is fully opposing the rotation of the wheel. The other forces for this load case

are shown in Table 1. The torque at the wheel based on its diameter of 508mm is

86Nm in a clockwise direction; this is from the x component of force in Table 1 and

a coefficient of friction of 1.7.

Load Case B: Cornering, acceleration and bump applied at the wheel and these

forces are shown in Table 3. Torque is also applied in this load case to create the

acceleration and the maximum torque from the drive train is shown in Table 2.

7.1 Hub Shaft Testing Method Investigation

Initial analysis was carried out using ANSYS 13.0 to verify that the use of a remote force to

simulate the application of force to the wheel is an accurate assumption. This was done by

modelling a solid hub shaft with basic geometric features. This also tested whether the hub

shaft in a solid form (Figure 72) would be capable of withstanding the forces in Table 1.

Figure 72: Basic Hub Shaft Solid

All analyses were performed in ANSYS Transient Structural, over a time period of one

second. All bodies except the hub shaft were rigid, with the mesh used for all analyses

shown in Figure 73. Bearings were used in the model to simulate an accurate support area on

the hub shaft as the hub shaft would be supported in the upright. The bearings were fixed in

all directions of translation and rotation using a general body to ground joint. The bearings

and wheel were fixed to the hub shaft using bonded contacts. The wheel and hub shaft were


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bonded as it is assumed that the bolts will be tightened to a sufficient torque such that the

friction between the hub shaft and wheel effectively bonds them together.

Figure 73: Force Application Verification Mesh

The wheel was attached to ground using a general body to ground joint with all translation

directions and rotations set to free. This allows the wheel to apply force in all directions,

including forces due to moments at the tyre contact patch. The force components are the

same as those in Table 1 (Figure 74).

Figure 74: Forces on Wheel using a General Body to Ground

Another force application model applied the forces to the tyre by using a flat contact patch

instead of applying the force to the tyre circumference (Figure 75). The flat contact patch

was attached to ground using the same method as the previous model.


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Figure 75: Force Application for Wheel Contact Patch Model

Figure 76: Fatigue Safety Factor for,

LEFT: Wheel Circumference Force Loading RIGHT: Wheel Contact Patch Force Loading

The final model tested was applying the force to the hub shaft through a remote force

situation at the contact patch between the tyre and the ground. The same force components

were applied as in Table 1 with the same contact patch location used for the upright

verification (Figure 77).


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Figure 77: Remote Force Location

The fatigue safety factor result for the remote force (Figure 78) shows that there is a reduced

safety factor around the area where the bearing is in contact with the hub shaft. This is

because the remote force is loading the edge of the hub shaft axially, radially and by a

moment from the contact patch. This is a more realistic application of the force as there is a

moment on the tyre due to the lateral forces on it. This force is not apparent when the force

is simply applied to the wheel as in the previous two models.

Figure 78: Fatigue Safety Factor for Remote Force Loading


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For all future analyses of the hub shaft, a remote force will be used to apply loads that would

occur due to loading from the wheel. By using a remote force to replicate the loading of the

hub shaft, the wheel does not need to be included in the model. This is expected to reduce

computational time of the model.

7.2 Hub Shaft Detailed Design

Analyses were carried out on the Hub Shaft 3.0 to ensure that its design was adequate for the

application. In this section, the hub shaft will be verified and optimised against load cases A

and B. The final hub shaft design shown in the following results were developed from

iteration and continuous optimization of the hub shaft shape. This was done by reducing the

diameter between the brake rotor and the hub shaft face (bending diameter). The results

shown in this section will be of the final hub shaft design only.

The material properties used for ANSYS are based on the material properties in Appendix

A11.3. The fatigue data used is shown in Table 21; these strength values were based on a

theoretical S-N curve for ferrous materials.

Table 21: 4140 Steel Fatigue Properties

Cycles Strength Equation

10 930MPa Sut

1000 837Mpa 0.9Sut

1,000,000 465Mpa 0.5Sut

The hub shaft fatigue life was determined based on imported fatigue data for 4140 steel

(Table 21). The life of the hub shaft was analysed to a life of 10E6 cycles as this is assumed

as infinite life. In ANSYS fatigue tool, a fatigue strength factor of 0.7 and fully reversed

loading were used.

The hub shaft was modelled in Transient Structural over a period of one second; all forces

were ramped. All models were solved over one time step, using a minimum number of sub

steps of five and a maximum number of sub steps of ten.

The load case A mesh (Figure 79), was used for all of the models involving the hub shaft.

The mesh sizing used was 0.003m. The load case A model (Figure 80) constraints used in


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the model are a remote force at the outer hub shaft face, located at the wheel contact patch.

The remote force was applied to the whole surface as it is assumed that the bolt tightness is

sufficient to efficiently bond the hub shaft to the wheel. Compression only supports were

used at the connection points between the hub shaft and the brake rotor. The bearings were

set as rigid and fixed to ground using a general body to ground joint. A moment force was

also applied to the outer hub shaft face to simulate the torque of the tyre before slip.

Figure 79: Load Case A Mesh and Model

Figure 80: Load Case A Setup


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Both torque and bending are acting on the hub shaft in load case A which results in a critical

zone on the outer surface of the bending diameter. This is shown in Figure 81, where the

stress results at the surface are higher than the centre for the bending diameter. The highest

stress on the hub shaft in this load case is 1.479E8 Pa which is below the yield strength of

4140 steel (6.3E8 Pa). This high stress occurs at the stress raiser fillet of 5mm between the

bending diameter and brake rotor attachment. This load case results in a minimum fatigue

safety factor of 2.36 (Figure 82).

Figure 81: Load Case A Stress Plot (Cross-Sectional View)

Figure 82: Load Case A Fatigue Safety Factor


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Load case B is made up of a moment force from the engine applied at the CV side and a

remote force at the hub shaft face located at the wheel contact patch. Ideally these two loads

are applied in the same simulation; however, due to the constraints required to get the model

to solve, the model was producing unrealistic results. Therefore, the moment and bending

load were analysed in separate load cases.

The moment force was analysed by applying the moment (247.78Nm) to the entire surface

of the CV section (Figure 83). The outer surface of the hub shaft was fixed using a fixed

support. This model simulates the maximum torque from the differential to the hub shaft in a

worst case scenario where the tyre is fixed. The bearings were connected to ground using a

general body to ground joint fixed in all direction except for rotation in z.

Figure 83: Load Case B Moment Setup

The stress from this model is shown in Figure 84 with the results showing that the critical

area is between the two bearings. This results in a critical fatigue safety factor of 4.8 in this

location (Figure 85). The result shows that there is not stress where the bearings are mated

with the hub shaft. This is a valid result as the bearings fit, must be tight enough such that

there is no hub shaft slip in the bearing. Such slip would result in stress either side of the

bearings.


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Figure 84: Load Case B Torque Only Stress (Cross-Sectional View)

Figure 85: Load Case B Moment Fatigue Safety Factor

The second model used to simulate load case B is the application of the remote force to the

outer surface of the hub shaft located at the tyre contact patch (Figure 86). The bearings

were connected to ground using a general body to ground joint fixed in all directions.


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Figure 86: Load Case B Remote Force Setup

Like load case A the critical area of the model is the outer surface of the bending diameter

(Figure 87). The highest stress is located between the brake rotor attachment and the bending

diameter (1.23E8 Pa). This results in a critical fatigue safety factor of 2.8 in this location.

Figure 87: Load Case B Remote Force Stress (Cross-Sectional View)


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Figure 88: Load Case B Remote Force Stress

The critical safety factor for the rear hub shaft is 2.8 under the remote force load. It is

assumed that a combined load case would not significantly reduce the safety factor. This

assumption is made based on the locations of critical stress being in different locations

between the two models, evident from the locations of the critical safety factors. There

would be little stress combination in the critical areas of the two models. For example, from

basic stress addition, the maximum stress at the bending diameter would be approximately

1.95E8 Pa. Still well below the yield criteria of 4140 steel.

7.3 Keyway Detailed Analysis

The CV section was analysed under the maximum torque load to verify that its design is

capable of this application. The moment load was applied to the bolt holes as shown in

Figure 89, with a cylindrical support at the centre of the CV section. One edge of the keyway

was fixed to simulate a worst case scenario where the hub shaft is fixed, and the key is

preventing movement.


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Figure 89: CV Verification Setup

High stress around the keyway was expected as it creates a large stress concentration point,

however, under fatigue the CV section has a safety factor of 15.

Figure 90: CV Verification Stress

The hub shaft keyway was verified by applying the moment resultant force (12000N) to one

edge of the keyway where the key would be impacting (Figure 91). The hub shaft was

supported by the bearings and fixed at the hub shaft wheel bolt holes. The bearings were set


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as rigid and fixed to ground using a general body to ground joint free in z – rotation. Again

as expected, the high stress area is at the keyway with 2.04E8 Pa (Figure 92) and a resulting

fatigue safety factor of 1.7 (Figure 93).

Figure 91: Keyway Verification Setup

Figure 92: Keyway Verification Stress


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Figure 93: Keyway Verification Fatigue Safety Factor

7.3.1 Weld Analysis

As the hub shaft will be made from several sections welded together (Figure 94) it is

important to verify that the stress at these weld points does not exceed the strength of the

weld electrode. For welding of 4140 steel, a Bohler Fox CM 2 Kb welding electrode can be

used as filler material. This electrode has a quenched yield strength of 480Mpa (Section

A.11.4 Welding Properties).

Figure 94: Hub Shaft Weld Locations

Basic hand calculations were done to determine the safety factor for the two critical fillets

(welds) on the hub shaft, the 5mm fillet between the spindle and the brake rotor attachment

and the 12mm fillet between the spindle and hub shaft outer flange. The 5mm weld had a

safety factor of 3.24 and the 12mm a safety factor of 4.89. These calculations were based on

12 mm Weld

5 mm Weld


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the highest stress from the ANSYS analysis at the respective locations. For full calculations

refer to Appendix A6.3

7.4 Hub Shaft Recommendations

From the ANSYS analysis undertaken on the hub shaft shown in Figure 95. This hub shaft

design has been optimised to minimise material and reduce machining costs. It is

recommended that these hub shafts be manufactured using 4140 steel. The torque

transmission method to be used is a key way on the CV side hub shaft section. A key way

was chosen because it is less complex than a spline, reducing manufacturing costs.

Figure 95: Final Hub Shaft Design

For the front hub shafts (Figure 96) the tripod connection flange does not need to be

connected as there is no torque transfer required on the front wheels. Therefore the CV

section is not needed. However, a washer and lock nut are still required to provide a preload

on the bearings.


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Figure 96: Final Front Hub Shaft

To reduce the amount of material wasted during the manufacturing process it is

recommended the hub shaft be made out of multiple sections and welded together as

necessary. An exploded view of all the different components of the hub shaft are shown in

Figure 97. For full hub shaft drawings refer to Appendix A3.

Figure 97: Final Hub Shaft Exploded View


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From the work done throughout this report on the hub shafts it is recommended that the

unsprung mass on the front and rear of the car be designed in the future to accept inboard

brakes. This is recommended as it has multiple advantages for performance and reduce

manufacturing costs of the hub shaft. These advantages are:

Reduce bending moment on the hub shaft as the distance between the upright

bearings and the wheel can be greatly reduced. This will reduce the amount of

material, time, and specialised manufacturing techniques required to make the hub

shaft and reduce weight.

By reducing the weight of the unsprung mass suspension components this is a

benefit to the overall performance and handling of the car.

By moving to inboard brakes, torque transmission and brake mounting geometry will be

required on the front and rear of the car. These modifications can be easily designed into

future frame designs. However, it is expected that the benefits of inboard brakes out-way

those of outboard brakes.


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8 Conclusion

As a part of this project, a suspension test apparatus has been designed, built and tested to

gain experimental data on strain at critical locations of various suspension components. This

apparatus is flexible in its design to allow future suspension assemblies to be tested on this

apparatus. The implementation of this suspension test apparatus will increase safety in

design and allow components to be tested before track testing.

Several computational models have been tested and compared against experimental data to

determine the accuracy of various assumptions made within each model. In future, this will

allow for more accurate design of suspension components.

The upright and hub shaft designs have been verified using computational techniques using

new considerations and modelling simulations from previous years. Both the designs of the

upright and huh shaft (shown assembled in Figure 98) satisfy the requirements for

application in the FSAE competition. It is recommended once fabricated that both designs

are tested using the suspension test apparatus.

Figure 98: Final Thesis Assembly

As a part of this thesis, several areas of future work have been identified. Future research

beyond this thesis may address the following, but not limited to;


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Further experimental testing using the experimental test apparatus with larger load

weights to develop a more accurate strain vs. load equation. More data in this area

will also allow for greater comparison to computational simulations.

Experimental testing involving the hub shaft and upright once fabricated to validate

assumptions made during analysis and verify the designs. This can be achieved by

the use of strain gauges on the hub shaft on the spindle to determine bending

reaction forces.

Further development of computational models in comparison to experimental data to

develop a working model capable of designing A-Arms. Current models in this

thesis have only moderate accuracy to experimental data.

This future work is believed to be beneficial to the development of the Motorsports Team

and to the safety and reliability of future design work involving the suspension.


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9 References

1. Hunter, R., S. Baker, and T. Doyle, Upright Report, 2011. 2. Ikealumba, W., D. Authors, and D. Brendan, JCU Formula SAE

Motorsports Project, 2011. 3. Crouch, J., Design and Simulation of Formula S.A.E Car

Suspension. 2011. 4. Crisp, B. and J. Meehan, JCU Motorsport Suspension Report,

2011. 5. Savage, G., Dealing with crisis – Solving engineering failures in

Formula 1 motor racing. Engineering Failure Analysis, 2010. 17(4): p. 760-770.

6. Savage, G., Sub-critical crack growth in highly stressed Formula 1 race car composite suspension components. Engineering Failure Analysis, 2009. 16(2): p. 608-617.

7. Gary, S., Multi-Axis Suspension Testing. Advanced Materials & Processes, 2007.

8. Pedersen, D., C. Dove, and D. Vellanikaran, FSAE Motor Sports Vehicle JCU Differential Design. 2011.

9. A., J.B. and B. J., Design of Formula SAE Suspension. SAE Technical Series, 2002.

10. Wong, A.J., Design and Optimization of Upright Assemblies for Formula SAE Racecar, in Department of Mechanical and Industrial Engineering2007, University of Toronto: Toronto.

11. McCune, M., et al., Formula SAE Interchangable Independent Rear Suspension Design, 2009.

12. Vehicle Dynamics. 2009 [cited 2012 22/4/2012]; Available from: http://ritzel.siu.edu/courses/302s/vehicle/vehicledynamics.htm.

13. RetroRims. VW Suspension Tuning. 2012; Available from: http://www.retrorims.co.uk/vw-blog/vw-suspension-tuning.

14. Smith, C., Tune to Win, West Aviation Road, Fallbrook, CA: Aero Publiashers, Inc.

15. Centre., B.T.a.E. Camber. 2012; Available from: http://www.barrystyre.co.uk/80100/info.php?p=5.

16. MOOG, Automotive Testing Overview, 2010. 17. MOOG 4-Poster suspension tester. 2012; Available from:

http://www.moog.com/markets/automotive/automotive-


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test-simulation/automotive-structural-testing/4-poster-test-rig/.

18. MOOG 16-DOF suspension test rig. 2012; Available from: http://www.moog.com/markets/automotive/automotive-test-simulation/automotive-structural-testing/suspension-test-rig/.

19. Jun, K.J., et al., Prediction of Fatigue Life and Estimation of its Reliability on the Parts of an Air Suspension System. International Journal of Automotive Technology, 2008. 9(6): p. 741-747.

20. Lee, Y.-L., et al., Fatigue Testing and Analysis2005, Jordan Hill, Oxford: Elsevier Butterworth-Heinemann.

21. Failure analysis of bolted joints: Effect of friction coefficients in torque–preloading relationship. 2010.

22. Koh, S.K., Fatigue analysis of an automotive steering link. Engineering Failure Analysis, 2009. 16(3): p. 914-922.

23. Colombo, D., et al., Analysis of a Unusual McPherson Suspension Failure. Engineering Failure Analysis, 2009.

24. Kadhim, N.A., S. Abdullah, and A.K. Ariffin, Effective Strain Damage Model Associated with FInite Element Modelling and Experimental Validation. International Journal of Fatigue, 2012.

25. Topaç, M.M., S. Ercan, and N.S. Kuralay, Fatigue life prediction of a heavy vehicle steel wheel under radial loads by using finite element analysis. Engineering Failure Analysis, 2012. 20(0): p. 67-79.

26. Topaç, M.M., H. Günal, and N.S. Kuralay, Fatigue failure prediction of a rear axle housing prototype by using finite element analysis. Engineering Failure Analysis, 2009. 16(5): p. 1474-1482.

27. FSAE. 2012 Formula SAE Rules 2012; Available from: http://students.sae.org/competitions/formulaseries/rules/2012fsaerules.pdf.

28. Wallace, J. and D. Whittering, FSAE Brake Calliper and Master Cylinder Review, 2011.


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Appendix

A.1 Apparatus Drawings


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A.2 Upright Manufacturing Drawing


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A.3 Hub Shaft Drawings


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A.4 Relevant FSAE Suspension Rules

Below are relevant suspension related rules that impact on the design methodology of this

project. For full detailed rules refer to Formula SAE Rules 2012 which is available from,

http://students.sae.org/competitions/formulaseries/rules/2012fsaerules.pdf

Tilt Test

The car is tilted in either direction to 60°, to test the car’s roll over stability. At this angle it

represents 1.7 G’s laterally. At this angle the car must not roll over and must be performed

with the tallest driver in the car.

Skid Pad

The aim of the skid-pad event is to measure the car’s cornering ability on a flat surface while

making a constant-radius turn. The skid pad track will be a figure eight configuration; the

centres of the two loops will be 18.25 m apart.

Suspension Travel Requirements

The car must be equipped with a fully operational suspension system with shock absorbers,

front and rear, with usable wheel travel of at least 50.8 mm (2 inches), 25.4 mm (1 inch)

jounce and 25.4 mm (1 inch) rebound, with driver seated.

http://students.sae.org/competitions/formulaseries/rules/2012fsaerules.pdf


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A.5 Project Timeline

Figure 99: Project timeline


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A.6 Hand Calculations

A.6.1 Weight Arm Calculations

The purpose of these calculations is to determine the force required on the weight arm to

achieve 3653 N at the contact patch of the tyre.

Assume hinge point is point of rotation.

∑

( ) ( )( )

Rearranging for Ry

( )( )

( )

A.6.2 Push Rod Strain Calculations

Total Tensile Force = 2260.7N

E = 210E9

Outside Diameter, OD=0.02m

Inside Diameter, ID=0.017m


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A.6.3 Hub Shaft Weld Calculations

All stress values have been taken from ANSYS analysis. Yield strength is taken from

electrode properties.

( )

( )

12mm Fillet Safety Factor

5mm Fillet Safety Factor


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A.7 Bolt Calculations

Suspension Test Apparatus

Given/Sourced Information

Force Components

x 9044.8 N

y -103.41 N

z 0 N

Coefficient of

friction μ 0.45

Elasticity of Mild Steel (Weight Arm)

EA 207 Gpa

Elasticity of Mild Steel (Apparatus)

ES 207 Gpa

Length of Bolt

Lb = 35 mm 0.035 m

Thickness of Apparatus Wall

LA = 5 mm 0.005 m

Thickness of Weight Arm

Ls = 30 mm 0.03 m

(Top Area of Weight Arm in Contact with Apparatus)/2 (As

there is 2 bolts on Top)

Ab = 7005 mm2 0.700 m2


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5

Capscrew grade 8.8

Bolt Dimensions [Metric] (Sourced from Fundamentals of Machine Component Design 4th

edition) Table 10.2 and 10.5

M8x1.25 Coarse Thread

Nominal Diameter

12 mm

d=

Pitch Radius

1.75 mm

p=

Stress Area

84.3 mm2 8.43E-05 m2

At =

Proof Load

600 Mpa

Sp =

Yield Strength [Sy]

660 Mpa

Sy =

Tensile Strength [Su]

830 Mpa

Su=

Stress Factor

3.0

Kf_inner (rolled)=

This is the stress factor sourced from

Fundamentals

of Machine Component Design 4th edition table


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10.6

Analysis

Calculations for Static Lower bolt

Axial Force

P=F/At

-1.23 Mpa

Shear Force

107.29 Mpa

Equivalent Von mises’ stress

185.84 Mpa

Safety Factor

N= Sp/Axial

Force

3.23

Calculations for Fatigue

Se’ [Mpa] 415

Se’ = 0.5Su

Load Factor (Ca) 0.9 Axial Loading

Size Factor (Cb) 1 Axial Loading

Surface Factor (Cc) 1 Already factored into stress factor

Temperature Factor (Cd) 1 as T< 350°C

Reliability Factor (Ce) 0.814 99% Reliability

Stress Concentration (Cf)

Notch

Sensitivity (q)

Notch Radius 0.875

Notch Radius =

p/2

q = 0.825

Sourced using 1.0

Gpa Steel

Cf = 0.377358

Cf =


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1/(1+q*(Kf-1))

Miscellaneous Factors (Cg) 1

Se = Se=Se’CaCbCcCdCeCfCg

114.73 Mpa

Bolt Stiffness

kb =

AE/L

4.99E+08

Aluminium Stiffness

kA =

AE/L

2.9E+13

Bearing Stiffness

ks =

AE/L

4.83E+12

Stiffness kc of combined aluminium and bearing stiffness

kc =

1/(1/ks+1/kA+1/kA)

3.63E+12

kb/(kb+kc )

0.000138

Fe =

-103.41 N Assuming only Axial loading

Fbmax= Fb=Fi+(kb/(kb+kc )) Fe

45521.99 N

Fbmi

n =

45522 N


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σbm

ax= σb= Fb/At

540.00 Mpa

σbmi

n=

540.00 Mpa

σa=

σa = (σbmax-σbmin)/2

0.00 Mpa

σm =

σm = (σbmax+σbmin)/2

540.00 Mpa

Goodman Criteria σa/Se + σm/Su= 1/N

1.54

Yield Criteria σa/Sy + σm/Sy = 1/N

1.22

Maximum Tensioning Force (ki =0.9)

Fi =

Fi = kiAtSp

45522 N

Maximum Tightening Torque on Bolt

T =

T = 0.2Fid

109.2528 Nm

Friction Load Carrying Capacity

Ff =

Ff = Fiμ

20484.9 N

486 Mpa Ff = Fiμ / At


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A-Arm is in

double shear

Shear Safety

Factor

4.53


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A.8 Risk Assessment Documents


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A.10 HBM Strain Gauge Calibration Data


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A.11 Material Properties and Data Sheets

A.11.1 Alumec Material Properties

General material data is shown below, for full material details refer to Bohler Uddeholm

product details.


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A.11.2 7075 Aluminium Alloy Fatigue Data

Fatigue data found from

http://www.knovel.com/web/portal/basic_search/display?_EXT_KNOVEL_DISPLAY_boo

kid=754 (Page 3-382)

http://www.knovel.com/web/portal/basic_search/display?_EXT_KNOVEL_DISPLAY_bookid=754

http://www.knovel.com/web/portal/basic_search/display?_EXT_KNOVEL_DISPLAY_bookid=754


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A.11.3 4140 Steel Properties


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A.11.4 Welding Properties


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tobias doyle final thesis

Documents

suspension system

design methods

optimization of design

vehicles tobias doyle

university library

suspension test apparatus

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