Formula SAE Suspension Design

THE UNIVERSITY OF QUEENSLAND Bachelor of Engineering Thesis

Formula SAE Suspension Design

Student Name: DANIEL RAYMOND BURT Course Code: MECH4500 Supervisor: Dr. Ross McAree Submission date: 7th November 2003 A thesis submitted in partial fulfillment of the requirements of the Bachelor of Engineering degree program in the Division of Mechanical Engineering

School of Engineering

Faculty of Engineering, Physical Sciences and Architecture

Daniel Raymond Burt 62 Ellen St Woody Point QLD, 4019

7 November 2003

Prof. J. M. Simmons Head of School School of Engineering University of Queensland Brisbane Queensland 4072

Dear Sir, I hereby submit my Thesis titled “Formula SAE Suspension Design” for consideration as partial fulfilment of the Bachelor of Engineering degree. All the work contained within this Thesis is my original work except where otherwise acknowledged. I understand that this thesis may be made publicly available and reproduced by the University of Queensland unless a limited term embargo on publication has been negotiated with a sponsor. Yours sincerely,

Daniel Raymond Burt 33628055

ABSTRACT

Formula SAE is a student project undertaken by the Mechanical Engineering department of the University of Queensland and various other universities in Australasia, America, Europe and England. It is a competition to engineer and build a racing car to compete in design and track events.

The objective of my thesis is to analyse the performance of the 2001 and 2002 formula SAE racing car of the University of Queensland, identify it’s short comings in terms of suspension/steering geometry, set up and structural integrity and improve the design for the 2003 formula SAE.

The thesis follows the format of a design analysis. The investigation and analysis of the performance of the previous 2 years formula SAE race cars of the University of Queensland is used as a platform as to the complete redesign of the suspension system of the 2003 University of Queensland formula SAE race car.

It will discuss the design of the suspension and steering for the 2003 University of Queensland formula SAE racecar in order to optimise its performance and ability to be tuned to a particular racing course.

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ACKNOWLEDGEMENTS

I would like to express my appreciation to the following people for their valuable contribution and assistance in the completion of this thesis:

All fellow UQ Racing team mates from both the 2002 and 2003 teams, for their dedication, passion and extreme time commitment needed to be a part of this formula SAE team. In particular, George Commins and Francis Evans for their countless hours of support and technical advice/ input on the suspension design.

Mr George Dick, for his patience, technical tuition, guidance and dedication to the formula SAE project throughout the year.

The workshop staff; John, Ross, Dave and Neil, for your technical assistance and attention to the formula SAE project throughout the year.

Graham for time spent using instron to test rod ends strength and spring rates.

Professor Ross McAree for being my thesis supervisor, and being inspirational in his systematical approach to all engineering problems.

Professor David Mee for being the academic supervisor of the Formula SAE project in the University of Queensland.

Professor Hal Gurgenci for making the Formula SAE project available to students at the University of Queensland.

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CONTENTS ABSTRACT........................................................................................................................ I ACKNOWLEDGEMENTS ............................................................................................ II 1. INTRODUCTION......................................................................................................... 1 2. COMPETITION & DESIGN OBJECTIVES ............................................................ 2 2.1 Formula SAE Competition............................................................................... 2 2.2 Formula SAE Rules - Suspension/Steering ..................................................... 3 2.3 Formula SAE Rules - Dynamic Events............................................................ 5 2.4 Design Objectives............................................................................................. 11 3. BRIEF LITERATURE REVIEW............................................................................. 12 3.1

Wheelbase & Track .................................................................................. 12

3.2

Roll Centres ............................................................................................... 12

3.3

Pitch Centre ............................................................................................... 16

3.4

Anti-dive, Anti-lift & Anti-Squat ............................................................. 17

3.5

Camber....................................................................................................... 19

3.6

Toe .............................................................................................................. 21

3.7

Steering Geometry .................................................................................... 22 3.7.1

Kingpin Inclination & Castor .................................................................22

3.7.2

Ackerman Steering .................................................................................23

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4. CRITICAL ANALYSIS OF 2001/2002 SUSPENSION DESIGN/SETUP............ 24 5. 2003 SUSPENSION GEOMETRY DESIGN ........................................................... 27 5.1

Wheelbase .................................................................................................. 27

5.2

Track .......................................................................................................... 28

5.3

Roll Centres ............................................................................................... 29

5.4

Anti-dive, Anti-squat, Anti-lift & Pitch Centre...................................... 31

5.5

Anti-roll Rates ........................................................................................... 31

5.6

Steering Geometry .................................................................................... 33

5.7

Kingpin/Castor.......................................................................................... 34

5.8

Camber Gain............................................................................................. 34

6. OVERVIEW OF 2003 SUSPENSION COMPONENT DESIGN .......................... 36 6.1

6.2

Suspension Loading .................................................................................. 36 6.1.1

Strain Gauge Testing .............................................................................36

6.1.2

Bump and Braking Shock Factors .........................................................38

6.1.3

Fatigue Design.......................................................................................39

Wishbone Construction............................................................................ 40 6.2.1

Tubing Selection ....................................................................................41

6.2.2

Rod end Selection ..................................................................................41

6.2.2.1 Testing...............................................................................................42 6.2.3

Insert Design ...........................................................................................44

6.3

Rocker and Upright Design..................................................................... 45

6.4

Spring and Damper................................................................................... 49

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6.4.1

Selection ................................................................................................49

6.4.2

Damper Dynamometer...........................................................................50

6.4.3

Spring Testing........................................................................................51

6.4.4

X-Ray.....................................................................................................52

6.5

Steering Selection...................................................................................... 53

6.6

Accuracy & Adjustment ........................................................................... 54

6.7

Component Placement.............................................................................. 55 6.7.1

Rocker, Spring and Damper and Anti-roll bar Placement .....................55

7. VEHICLE SET UP ..................................................................................................... 58 8.

RESULTS............................................................................................................... 60

9. CONCLUSIONS & RECOMMENDATIONS ......................................................... 64 10.

BIBLIOGRAPHY............................................................................................. 66

APPENDIX A .................................................................................................................. 67 APPENDIX B .................................................................................................................. 71 Roll Centre Analysis.............................................................................................. 71

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LIST OF FIGURES Figure 2.1: The incredibly small Tokyo Denki Formula SAE racecar was extremely competitive...................................................................................................................9 Figure 2.2: The 2002 UQ Formula SAE racecar at the Mount Cotton Hillclimb..............10 Figure 3.1 Roll Centre........................................................................................................13 Figure 3.2 Track Variations ...............................................................................................15 Figure 3.3 Pitch Centre ......................................................................................................16 Figure 3.4 Anti-dive ...........................................................................................................17 Figure 3.5 Forces applied to contact patch during braking................................................18 Figure 3.6 Camber..............................................................................................................19 Figure 3.7 Camber Change ................................................................................................20 Figure 3.8 Toe definitions ..................................................................................................21 Figure 3.9 Castor & Kingpin Inclination ...........................................................................22 Figure 3.10 100% Ackerman .............................................................................................23 Figure 4.1 Poor Wear on Front Tires of 2002 Car .............................................................25 Figure 5.1 Front Wishbones are Swept Backwards ...........................................................28 Figure 5.2 Front roll Centre Movement with 1° of roll (All measurements are in inches) ....................................................................................................................................30 Figure 5.3 Rear roll Centre Movement with 1° of roll (All measurements are in inches).30 Figure 5.4 Anti-roll bar Design (only half CAD modelled) ..............................................32 Figure 5.5 Relative Anti-roll Stiffness with Rear Anti-roll Bar Adjustment.....................33 Figure 6.1 Strain Gauge Testing ........................................................................................37 Figure 6.2 Rear Lower Wishbone ......................................................................................40 Figure 6.3 Initial Setup of Rod End Bearing Testing ........................................................42 Figure 6.4 Revised Setup of Rod End Bearing Testing .....................................................43 Figure 6.5 Results of Rod end Bearing Testing .................................................................44 Figure 6.6 Wishbone Insert ................................................................................................45 Figure 6.7 Front Upright ....................................................................................................46 Figure 6.8 Rear Upright .....................................................................................................47 Figure 6.9 Front Rocker .....................................................................................................48

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Figure 6.10 Rear Rocker ....................................................................................................48 Figure 6.11 Risse Racing Jupiter 5 Shock .........................................................................49 Figure 6.12 Damper Dyno Results of 1 Shock ..................................................................50 Figure 6.13 Graph of Results .............................................................................................51 Figure 6.14 Steering Rack in 2003 Racecar.......................................................................53 Figure 6.15 Chassis on Jig for suspension Pickup Accuracy.............................................54 Figure 6.16 Front Spring and Damper Placement .............................................................56 Figure 6.17 Rear Spring/Damper and Anti-roll Bar Placement.........................................57 Figure 8.1 Rear Toe Deflection..........................................................................................61 Figure 8.2 Toe Control Solution ........................................................................................62

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LIST OF TABLES Table 2.1 Competition Points ..............................................................................................2 Table 4.1 2001/2002 Suspension Parameters ....................................................................24 Table 5.1 Camber Gain Coefficients..................................................................................35 Table 6.1 Actual Spring Rates Results ..............................................................................52

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‘Formula SAE Suspension Design’

1. INTRODUCTION

The University of Queensland was represented for the second time at the third Australasian Formula SAE competition at in 2002. The previous attempts at the competition had yielded average results. However, the quality of the racecars of those teams at the forefront of the competition in 2002 was world class and therefore, the need for drastic improvement, if the University of Queensland were to be competitive in the future became evident.

With regard to the suspension and steering systems there was still a lot of room for improvement. The 2002 racecar still maintained many of the handling characteristics of the 2001 vehicle. The terminal under steering nature of the 2001 racecar continued in the 2002 vehicle, only to a slightly less extent.

The limitations to the modifications able to be made with regard to chassis geometry somewhat restricted Riseley [6] from achieving anything other than small changes in suspension geometry. The main focus of the redesign in the 2002 University of Queensland racecar was to rectify the critical errors made in 2001. Errors that were imposed due to the inexperience of the team, being it’s first year. Some of these errors were roll centres that ended under the ground, drive shafts at a large angle in the static rest position and both wishbones and rod end bearings in bending.

The complete reanalysis of the suspension design for the 2003 racecar was undertaken and is the basis of this thesis. The rectification of the terminal under steering nature from the previous racecars, for the 2003 racecar, was obviously essential for the success of the 2003 formula SAE team.

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2. COMPETITION & D ESIGN O BJECTIVES 2.1 Formula SAE Competition The Formula SAE competition is a international competition, run by the Society of Automotive Engineers, for students to conceive, design, fabricate, and compete with a small formula -style racing car. The competition is comprised of static and dynamic events. The competition events test the vehicles engineering design, and performance.

EVENT

Points

Static Events Presentation Engineering Design Cost Analysis

75 150 100

Dynamic Events Acceleration Skid -Pad Event Autocross Event Fuel Economy Event Endurance Event

75 50 150 50 350

Total Points Table 2.1 Competition Points

1000

A full version of the rules can be downloaded from the Australasian Society of Engineers web page at www.sae-a.com.au/fsae/rules. Some extracts of the important parts of the rules, pertaining to the suspension/steering design are explained in the following.

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2.2 Formula SAE Rules - Suspension/Steering The strict rules pertaining to the suspension and steering system of the car are as follows.

3.1.2 Wheelbase and Vehicle Configuration

The car must have a wheelbase of at least 1525 mm (60 inches). The wheelbase is measured from the center of ground contact of the front and rear tires with the wheels pointed straight ahead. The vehicle must have four wheels that are not in a straight line.

3.1.3 Vehicle Track

The smaller track of the vehicle (front or rear) must be no less than 75% of the larger track.

3.2.1 Ground Clearance

Ground Clearance must be sufficient to prevent any portion of the car (other than tires) from touching the ground during track events.

3.2.2 Wheels and Tires

The wheels of the car must be 203.2 mm (8.0 inches) or more in diameter. The tires can be any size or type. Tire or wheel type, compound or size may not be changed after the static judging has begun. Tire warmers are not allowed. No traction enhancers may be applied to the tires after the static judging has begun.

3.2.3 Suspension

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 -3-

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mm (1 inch) jounce and 25.4 mm (1 inch) rebound, with driver seated. The judges reserve the right to disqualify cars which do not represent a serious attempt at an operational suspension system or which demonstrate unsafe handling.

3.2.4 Steering

The steering system must affect at least two wheels. The steering system must have positive steering stops that prevent the steering linkages from

2003 Formula SAE® Rules

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locking up (the inversion of a four-bar linkage at one of the pivots). The stops may be placed on the uprights or on the rack and must prevent the tires from contacting suspension, body, or frame members during the track events. Allowable steering free play will be limited to 7 degrees total measured at the steering wheel. 3.4.8 Roll Over Stability

The track and center of gravity of the car must combine to provide adequate rollover stability.

3.4.8.1 Tilt Table Test Rollover stability will be evaluated using a pass/fail test. The vehicle must not roll when tilted at an angle of 57 degrees to the horizontal in either direction corresponding to 1.5 G’s. The tilt 2003 Formula SAE® Rules 31 test will be conducted with the tallest driver in the normal driving position.

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2.3 Formula SAE Rules - Dynamic Events There are 5 different dynamic events in the competition. Four of which rely on the dynamic performance of the racecar. Those being, the acceleration, skid -pad, autocross and endurance events.

Since formula SAE racecars are, in essence, engine power restricted due to a regulations air intake restrictor sizing, the suspension package is paramount to the success of a formula SAE vehicles dynamic performance.

The successful design of a performance suspension sys tem is based on performance compromises, which will be discussed in later chapters. However, before these compromises can be assessed in the interest of performance in the competitions dynamic events, these dynamic events must be well understood. A breakdown of what these events involve follows.

5.4 ACCELERATION EVENT 5.4.1 Acceleration Objective The acceleration event evaluates the car’s acceleration in a straight line on flat pavement. 5.4.2 Acceleration Procedure The cars will accelerate from a standing start over a distance of 75 m (82 yards) on a flat surface. The foremost part of the car will be staged at 0.30 m (11.8 inches) behind the starting line. A green flag will be used to indicate the approval to begin, however, time starts only after the vehicle crosses the start line. There will be no particular order of the cars in each heat. A driver has the option to take a second run immediately after the first.

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5.5 SKID-PAD EVENT 5.5.1 Skid -Pad Objective The concept of the skid-pad event is to measure the cornering ability of the car on a flat surface while making a constant-radius turn. 5.5.4 Skid -Pad Layout There will be two circles of 15.25 m (50.03 feet) diameter in a figure eight pattern. The circle centers will be separated by 18.25 m (59.88 feet), and a driving path 3.0 m (9.84 feet) in width will be marked with pylons and a chalk line just outside the pylons. The start/stop line is defined by the centers of the two (2) circles. A lap is defined as traveling around one (1) of the circles from the start/ stop line and returning to the start/stop line. 5.5.6 Skid -Pad Procedure The cars will enter perpendicular to the figure eight and will take one full lap on the right circle to establish the turn. The next lap will be on the right circle and will be timed. Immediately following the second lap, the car will enter the left circle for the third lap. The fourth lap will be on the left circle and will be timed. Immediately upon finishing the fourth lap, the car will exit the track. The car will exit at the intersection moving in the same direction as entered. A driver has the option to take a second run immediately after the first.

5.6 AUTOCROSS EVENT 5.6.1 Autocross Concept The concept of the autocross event is to evaluate the car's maneuverability and handling qualities on a tight course without the hindrance of competing cars. The autocross course will combine the performance features of acceleration, braking, and cornering into one event.

5.6.2 Autocross Procedure There will be two Autocross-style heats, with each heat having a different driver. The car will be staged such that the front wheels are 2 m behind the starting line. The timer starts only after the car crosses the start line. There will be no particular order of the cars to run each heat but a driver has the option to take a second run immediately after the first. -6-

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Two (2) timed laps will be run (weather and time permitting) by each driver and the best lap time will stand as the time for that heat. The organizer will determine the allowable windows for each heat and retains the right to adjust for weather or technical delays. Cars that have not run by the end of the heat will be disqualified for that heat.

5.6.3 Autocross Course Specifications & Speeds The following specifications will suggest the maximum speeds that will be encountered on the course. Average speeds should be 40 km/hr (25 mph) to 48 km/hr (30 mph). Straights: No longer than 60 m (200 feet) with hairpins at both ends (or) no longer than 45 m (150 feet) with wide turns on the ends. Constant Turns: 23 m (75 feet) to 45 m (148 feet) diameter. Hairpin Turns: Minimum of 9 m (29.5 feet) outside diameter (of the turn). Slaloms: Cones in a straight line with 7.62 m (25 feet) to 12.19 m (40 feet) spacing. Miscellaneous : Chicanes, multiple turns, decreasing radius turns, etc. The minimum track width will be 3.5 m (11.5 feet). The length of each run will be approximately 0.805 km (1/2 mile) and the driver will complete a specified number of runs. The time required to complete each run will be recorded and the time of the best run will be used to determine the score.

5.7 ENDURANCE EVENT

5.7.2 Endurance Objective The Endurance Event is designed to evaluate the overall performance of the car and to test the car’s reliability. 5.7.4 Endurance Course Specifications & Speeds Course speeds can be estimated by the following course specifications. Average speed should be 48 km/hr (29.8 mph) to 57 km/hr (35.4 mph) with top speeds of approximately 105 km/hr (65.2 mph). Straights: No longer than 77.0 m (252.6 feet) with hairpins at both ends

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(or) no longer than 61.0 m (200.1 feet) with wide turns on the ends. There will be passing zones at several locations. Constant Turns: 30.0 m (98.4 feet) to 54.0 m (177.2 feet) diameter. Hairpin Turns: Minimum of 9.0 m (29.5 feet) outside diameter (of the turn). Slaloms: Cones in a straight line with 9.0 m (29.5 feet) to 15.0 m (49.2 feet) spacing. Miscellaneous: Chicanes, multiple turns, decreasing radius turns, etc. The minimum track width will be 4.5 m (14.76 feet). 5.7.5 Endurance General Procedure The event will be run as a single 22 km (13.66 mile) heat. Teams will not be allowed to work on their vehicles during the heat. A driver change must be made during a threeminute period at the mid point of the heat.

The aim in all events is to complete the circuit in the shortest elapsed time.

Notably the top speed in any event is 105 km/h, which is not considerably fast in the realm of motorsport. The lack of long straights, maximum of 77.0 m in Endurance and 60.0 m in Autocross, and abundance of tight corners mean that a successful formula SAE racecar is more likely to be one with a superior and refined suspension package rather than engine package.

Moreover, through observation in the previous formula SAE

competitions, the smallest racecars seemed to be more suitably sized for the circuits endured in the formula SAE competition. They were able to take different racing lines and in some instances effectively straight line chicanes thus always carrying more corner speed also increasing the speed down the straights.

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Figure 2.1: The incredibly small Tokyo Denki Formula SAE racecar was extremely competitive

In light of this, the racecar must be designed for the particular circuit in which it is to be raced. The formula SAE competition is of a tight and twisty nature and therefore, the racecar must be designed to perform on this type of circuit for success.

For example, if a formula 1 or a indy champ car was to be placed in the formula SAE competition it would barely be able to complete the circuit let alone with respectable times. Not that a formula SAE racecar is has anywhere near the cornering, braking or acceleration ability of one of these racecars, they are just not suited to a tight formula SAE type circuit.

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The racing of the 2002 formula SAE against the other hill climb cars at the mount cotton hill climb circuit displayed that the 2002 racecar was much more suited to a circuit of a larger size. It was more stable at the speeds achieved on this circuit as opposed to the much slower speeds achieved on a tight formula SAE circuit.

Figure 2.2: The 2002 UQ Formula SAE racecar at the Mount Cotton Hillclimb

The competition skidpad, autocross and enduro circuits are more like go-cart circuits, rather than circuits intended for a formula style racecar. Therefore, the ideal car for the competition is essentially to be a go-cart with mandatory suspension.

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2.4 Design Objectives In the previous 2 years the University of Queensland has entered a racecar aimed at just competing in the competition. This year, 2003 the University of Queensland formula SAE team is aiming to have a modest attempt at winning the Australasian competition.

The main design objectives for the 2003 formula SAE racecar are to produce a simple, lightweight and reliable vehicle capable of winning the 2003 Australasian Formula SAE competition.

The design objectives of this year’s suspension system are formed on somewhat a different perspective of the competition. The decision to omit the differential, for reasons discussed in later chapters, has prompted clever suspension design to overcome the potential problems in doing this. Also, having considered the dynamic events involved in the formula SAE competition the 2003 design philosophy is to build a go-cart with it’s mandatory suspension running on the shortest wheelbase limit of 1525 mm.

The effects of all suspension parameters have to be considered and the details of such design considerations are discussed here within.

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3. B RIEF LITERATURE R EVIEW

The purposes of this section is enlighten the reader as to some of the terms used in double wishbone suspension design and briefly explain the effects of each on the performance of the vehicle. Some of the terms that are related to double wishbone sus pension are as follows.

3.1 Wheelbase & Track The wheelbase is the distance between the front and rear wheels. In the formula SAE competition, the wheelbase is measured from the centre of ground contact of the front and rear tires with the wheels pointed straight ahead. The track is the distance between the centreline of the wheels on either side of the vehicle.

3.2 Roll Centres The roll centre is the instantaneous centre of rotation of the chassis about the ground. It is determined by the geometric layout of the suspension members. The roll centre is governed by the instantaneous centres of rotation of the wheels about the chassis. The picture below shows the instantaneous centres of rotation of the wheels about the chassis and the determination of the roll centre.

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Figure 3.1 Roll Centre

The roll centre placement has an effect on the weight transfer of the racecar, as well as the vehicle attitude on cornering.

There are two different types of weight transfer in roll that add to give the total weight transfer of the vehicle. They are elastic and geometric. The amount of each type of weight transfer is determined by the roll centre location. The formula for determining these weight transfers is shown below.

Lateral Suspended Mass Weight “Geometric” Transfer is found by the equation:

LGWT= SM*Lateral Acc*RC Height/Track

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Lateral Suspended Mass Weight “Elastic” Transfer is found by the equation:

LGWT= SM*Lateral Acc*(SM CG Height - RC Height)/Track

Where SM = Suspended Mass, RC = Roll Centre & CG = Centre of Gravity

The elastic weight transfer is the component of weight transfer that is transferred through the vehicles springs. The geometric weight transfer is transferred through the wishbones and consequently, does not contribute to chassis roll. That is to say, if the roll centre height were the same as the centre of gravity height, the chassis would not experience any roll at all during cornering.

Another important point is that geometric weight transfer is instantaneous, while elastic transfer is not due to the transient motion of the spring compression/extension and chassis roll.

The instantaneous centre of rotation of the wheel of about the chassis determines the track variation in different situations. It is displayed in the figure below.

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Figure 3.2 Track Variations

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3.3 Pitch Centre The pitch centre is just as important as the roll centre and is established in essentially the same manner as the roll centre.

Figure 3.3 Pitch Centre

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3.4 Anti-dive, Anti-lift & Anti-Squat The following diagram illustrates the calculation procedure of the anti-dive percentage.

Figure 3.4 Anti-dive

The angle B in the diagram is calculated by the evaluation of the forces that are applied to the tire contact patch as illustrated in the below diagram.

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Figure 3.5 Forces applied to contact patch during braking

The anti-lift and anti-squat geometries are calculated in exactly the same manner.

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3.5 Camber Wheel camber is important to racecar design. It is illustrated in the following diagram.

Figure 3.6 Camber

Excessive camber gain during driving can detrimental to mechanical grip a racecar can generate. The camber variations in different driving situations are illustrated in the following diagram.

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Figure 3.7 Camber Change

The camber variation design involves a trade of between the bump and roll situations. As illustrated, for no camber gain in the bump situation the instantaneous centre of rotation of the wheel about the chassis must be at distance infinity from the centre of the vehicle. I.e. the virtual swing arm length, VSAL = 8 .For no camber gain in the roll situation, the virtual swing arm length should be half the track.

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3.6 Toe The variation of the vehicles wheels from parallel to each other is referred to as toe. Toe is illustrated in the below diagram.

Figure 3.8 Toe definitions

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3.7 Steering Geometry The steering geometry is one of the most important things to a racecars design.

3.7.1 Kingpin Inclination & Castor Kingpin inclination and castor are illustrated in the following diagram.

Figure 3.9 Castor & Kingpin Inclination

The effects of both kingpin inclination and castor are to camber the wheels during steering. Kingpin inclination cambers the outside wheel positively and castor does the opposite, cambering the outside wheel negatively. Castor and kingpin inclination trail both affect the torque experienced at the steering wheel to turn the wheels. If either or both are excessive the torque required to turn the steering wheel too may be excessive. Excessive steering input force is not ideal. However, a lack of torque is just as undesirable. Some torque is required for adequate steering “feel”.

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3.7.2 Ackerman Steering Ackerman is the term that describes the phenomenon of toe in or toe out during turning. Perfect Ackerman, or 100% is achieved when both wheel are rotating around the same instantaneous center as the vehicle. This is illustrated in the diagram below.

Figure 3.10 100% Ackerman

Consequently, the inside wheel has to turn more than the outside to achieve this (toe out). The more the wheels toe out during turning the greater the Ackerman percentage. Parallel steering is 0% Ackerman. The Ackerman angle is defined as the difference between the two steered angles and therefore, the percentage Ackerman is defined as the percentage of this Ackerman angle to the perfect Ackerman angle.

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4. CRITICAL ANALYSIS OF 2001/2002 S USPENSION D ESIGN/SETUP The previous formula SAE racecars of the University of Queensland exhibited many handling issues. The general consensus of all drivers for both the 2001 and 2002 competitions was that the racecar exhibited excessive under steering in both the transient and steady state cornering.

The first step to the improvement for the 2003 formula SAE racecar is to evaluate or identify the problems with the cur rent and previous designs/setups.

The suspension parameters in the 2001 and 2002 formula SAE Racecar as they were designed is as follows.

Parameters

2001

2002

Wheelbase (mm) 1700 1760 Track, front (mm) 1285 1285 Track, rear (mm) 1285 1285 Tire Hoosier 20.0x6.0 - 13, R25 compound Scrub Radius (mm) 42 38 Kingpin Inclination (deg) 0 2.8 Castor (deg) 4 4.8 Castor Trail (mm) 20 21 Roll Centre Height, front (mm) -39.6 18 Roll Centre Height, rear (mm) 15.24 32 Steering Parallel 70% Ackerman Weight Distribution 61 Front Left (kg) Front Right (kg) 60 Rear Left (kg) 90 Rear Right (kg) 90 Table 4.1 2001/2002 Suspension Parameters

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The first thing to notice is that the roll centres were under the ground in 2001. This was unintentional an occurred as a result of incorrect tire diameter and deflection.

The under steering apparent in both racecars was a result of a rear suspension that was working well partnered with a front suspension that was not. The castor on either of the racecars was not great enough to deal with the camber gain in roll that the car exhibited in the front geometry. The evidence is clear that the front tires were cambering excessively, when the used front tires are closely examined.

Figure 4.1 Poor Wear on Front Tires of 2002 Car

Some other issues that become apparent with the previous cars are that the roll centres lateral movement in roll is outrageous. They almost leave the vehicle’s track. This kind of roll centre movement makes the car difficult for an inexperienced driver to predict. Also, the manufacturing techniques to acquire the proposed geometry were appalling. The suspension pick up points were welded to the chassis with a student holding them in position with pliers, whilst the welder tacked them on. If this kind of inaccuracies are

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going to be introduced in manufacturing then the roll centre analysis and any other geometry may as well be forgotten about.

The spring and relative anti-roll rates were initially analysed in 2001, but not tuned with an anti- roll bar, as the anti-roll bar designed for use in 2001 was a failure. The weight distribution moved forward in 2002, however, the spring and relative anti-roll rates were not reanalysed.

The steering was supposed to be 70% Ackerman in the 2002 racecar. Upon extensive theoretical and practical analysis it was determined that the 2002 racecar actually realistically had close to parallel steering. The position of the steering rack meant that the onset of inversion of the inside steering arm was delayed to much to have any effect at all.

The Torsen differential used on both these racecars is a torque sensitive differential. It locks up when under torque or a torque bias. The viscosity of the oil inside it determines the bias in which it locks. Both the drive and braking torques are passed through the differential so when the racecar is either accelerating or braking the differential is effectively locked. The formula SAE competition comprises of a tight and twisty circuit and as such most of the time on corner entry or exit the differential is locked. This point was illustrated when the differential failed and was permanently locked during the endurance event of competition in 2002. Not only did this not slow the car at all, it actually made the car easier to predict on corner entry and exit, due to the lack of transients of the differential locking.

All these factors of miscalculation, inaccuracy or lack of calculation at all lead to the average suspension performance of the previous racecars.

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5. 2003 S USPENSION G EOMETRY DESIGN

The 2003 suspension geometry design is entirely new to that of the 2002/2001 racecars. The decision to run the car with no differential has had a great effect on the approach of the goals of the suspension setup.

The omitted differential, meant that if the car would attempt to “Drive through the front wheels” if the weight transfers weren’t setup such as to prevent it. The main points of the design of the 2003 suspension was to achieve a lot more than usual trans verse weight transfer (from inside rear to outside front). This was to be achieved by the use of lots of castor and Ackerman in the steering.

The parameters chosen for the 2003 geometry and an explanation of the effects of these are described below.

5.1 Wheelbase The wheelbase was chosen to be the smallest on the rule limit at 1525 mm. This was to increase the longitudinal weight transfer and increase the turn responsiveness of the car. The small wheelbase will also benefit the weight transfer, providing more traction in both braking and accelerating situations. Also, to some degree the car can be steered with the application of throttle/brake on the onset of under steer or over steer respectively.

The front wishbones are swept backwards rather than even to get the wheelbase down whist keeping the rear attachment on the front roll hoop. It also allowed the steer to be in behind the lower wishbone to keep the steering rack low in the car whist avoiding bump steer. -27-

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Figure 5.1 Front Wishbones are Swept Backwards

5.2 Track The track is at 1200 mm front and 1100 mm rear. This difference in track was chosen to create more weight transfer in the rear than in the front and lower the difference in cornering wheel speed in the rear, to compensate for the omitted diffe rential.

The wheelbase to track ratio is then a lot smaller, at 1.27, than 2001 and 2002 with 1.33 and 1.37 respectively. This means that the racecar will possess a lot more agility with it’s decreased yaw resistance and yawing inertia. This value is more typical of a go-cart and therefore, will be more suited to the go-cart like circuits this racecar will endure.

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5.3 Roll Centres Driver feedback was one off the major concerns involved in the placement of roll centres. It was considered just as important as the effects of the roll centres in weight transfers and racecar cornering attitude. To achieve accurate and informative feedback to the driver not only the roll centre placement has to be considered but also the roll centre movement.

The roll centres are above and as close to the ground as feasible making sure they never pass through the ground, to control weight transfer by making it mainly elastic and avoiding jacking/change in jacking confusing the driver. The lateral roll centre movement was also cons idered and almost zero movement was achieved with a ratio of width of chassis in top to width of chassis in bottom of 1.5. The front roll centre is statically at 35mm and the rear at 38mm. The rear is slightly higher than the front to give the feeling of under steer initially on turn in. This is due to the track variation in roll. The higher roll centre in the rear means that the increase in track on the outside of the vehicle in the rear is greater than that of the front with a lower roll centre. The overall effect is that the racecar’s chassis rotates with respect to the wheels, ever so slightly, to face the outside of the turn. Thus creating an under steering type feeling.

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Figure 5.2 Front roll Centre Movement with 1° of roll (All measurements are in inches)

Figure 5.3 Rear roll Centre Movement with 1° of roll (All measurements are in inches) -30-

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5.4 Anti-dive, Anti-squat, Anti-lift & Pitch Centre Just as important to driver feedback is the pitch centre and anti-dive, anti-lift and antisquat geometries. The anti-dive and anti-squat was chosen to stop movement of the pitch centre. There is 10% of anti-dive and 0% of anti-squat. The reason for adding anti-dive and not anti-squat is because of the considerably larger spring rate in the rear combined with less weight transfer (i.e. better braking than accelerating). The pitch centre is 1 mm above the ground and 600 mm forward of mid wheel base and moves only 3 mm up and 384 mm forward when under full braking of 1.5g.

5.5 Anti-roll Rates The anti-roll rate not only determines the amount the chassis rolls during cornering but the relative anti-roll rates, front to rear, determine the weight transfer characteristics of the racecar. Many racecar engineers refer to the relative roll stiffness as the “magic number”. Changing the relative anti-roll rate front to rear is the single most effective way of establishing a balanced racecar. By changing the relative anti-roll rate and hence the relative weight transfer, the overall mechanical grip can be sacrificed at one end of the racecar to improve the other, until a balance is achieved.

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Figure 5.4 Anti-roll bar Design (only half CAD modelled)

The relative anti-roll rate can be adjusted by the single rear anti-roll bar. The adjustment in the anti-roll bar goes from 87 N.m/deg rotation to 424 N.m/deg rotation. This adjustment is achieved by adjustment in the lever arm length to the anti-roll bar. This allows a range of relative anti-roll rate ratio of 44% (softest anti-roll bar setting) to 32% (stiffest anti-roll bar setting) or 49% with out anti-roll bar.

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Roll Stiffness

% Front Anti-roll Stiffness

0.5 0.48 0.46 0.44 0.42 0.4 0.38 0.36 0.34 0.32 0.3 0

100

200

300

400

500

Rear Anti-roll Bar Stiffness (Nm/deg roll)

Figure 5.5 Relative Anti-roll Stiffness with Rear Anti-roll Bar Adjustment

5.6 Steering Geometry The choice of the percentage of Ackerman steering geometry to run is complicated. It is well documented that the addition of pro Ackerman (over 100%) or toe out will achieve better turn in response. Turn in response was something that was severely lacking in the previous UQ formula SAE racecars. However, from tire slip angle analysis, the perfect Ackerman to run would be approximately 70% for steady state, depending on the tire. Yet, in the opinion of many racecar engineers, the unloading of the inside tire in turning is so great that having the inside tire at the optimum slip angle is not really that important. In fact, trading down from the optimum lateral grip on the inside tire, which may in be almost nothing due to it’s severe unloading, is not much to compromise for the gains in turn in response achieved with pro Ackerman. With this in mind the steering is adjustable from 80% to 120% Ackerman. Interchangeable steering arms on the uprights achieve this adjustment.

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5.7 Kingpin/Castor The kingpin was chosen to be 0° as kingpin positively cambers the outside wheel when steered and an acceptable scrub radius of 30 mm could be achieved with out the use of kingpin. 7° of castor was chosen to promote a considerable amount of transverse weight transfer.

5.8 Camber Gain The camber gain in roll and dive was chosen considering the kingpin, castor and spring/anti-roll rates as well as the trade -off for different driving manoeuvres. The trade off between roll and dive was evaluated front and rear. It was decided that roll was more important than dive to improve cornering ability whilst lowering the need for negative static camber to keep the loaded wheel not positively cambered. Eliminating substantial static camber also helps with accelerating and braking situations. The front is less biased towards favouring roll as the castor helps in steering to maintain good camber attitude whilst cornering.

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The following values of camber gain where obtained in the new geometry. Front

Rear

Camber coefficient in bump (° / 10 mm bump)

0.5

0.6

Camber coefficient in roll (° / g)

0.42

0.38

Table 5.1 Camber Gain Coefficients

These values were analysed in suspension analyser to achieve 0° camber on the two outside, loaded wheels in corner. This was achieved by estimation of the steering angle in various radius corners at different lateral acceleration levels. Ultimately, the values were entered into a excel spreadsheet for ease of evaluation of the different situations. The above values were deemed to be the best compromise based on the evaluations of the spreadsheets results.

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6. OVERVIEW OF 2003 S USPENSION COMPONENT DESIGN 6.1 Suspension Loading One of the main goals of the 2003 design was to reduce the mass of the vehicle as much as possible without compromising structural integrity. With a competitive formula SAE racecar being approximately 220 kg (almost 2/3 of the existing 2002 University of Queensland racecar), a lot of weight was to be removed. This was only going to be achieved by making all components of the vehicle designed to be on the limit of failure for their maximum potential loading. Before a design on the limit of failure could be established, with respect to the suspension components, the exact details of the loading needed to be understood with great detail and accuracy.

The forces on the suspension members as well as upright and rocker loading could be evaluated by considering the maximum accelerating, braking and cornering loadings. However, the evaluation of the distribution of the loading to various wheels in different driving manoeuvres and the factor of safety to apply to this loading with respect to shock and bumpy road surfaces still needed to be established or explored in great detail.

6.1.1 Strain Gauge Testing The evaluation of these factors was established, or at least better understood, by the use of strain gauges measurements on all suspe nsion members on the left hand side of the existing 2002 University of Queensland racecar during testing sessions conducted on the 16th and 24th of April, 2003.

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Figure 6.1 Strain Gauge Testing

The strain gauge testing was conducted using an existing strain gauge amplifier built by Barry Alsop. It had 2 channels with Dataforth SCM5B38 strain gauge modules and a low pass filter. The two-pole low pass filter was of 3 Hz and the sample rate was 30 samples per second. The data was recorded by a laptop with a 8 bit A/D converter National Instruments DAQ card. The software doing the logging was a Lab view module. The strain gauges were dual element rosettes with one gauge in the axis of strain and one across the axis.

The one across the axis was for temperature compensation of the

wheatstone bridge. They were placed on one side of each member only. The rod ends and spherical bearings meant there was theoretically no bending in any of the members. So strain gauges on one side of the beam were adequate for measuring the tension/compression the beams experienced.

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The other two resistors, which completed the wheatstone bridge where chosen with the same temperature coefficient by heating the resistors and measuring the change in resistance.

The offset nulling for the wheatstone bridge was done at the track with the car on the ground. Therefore the vehicles resting weight was considered 0 strain for all members including pushrods.

Calibration of the strain gauges was performed by hanging various weights on one end of each member, with the other end suspended from the roof.

A testing circuit was designed to test the various driving manoeuvres of the vehicle were performed in a range of various testing runs. The maximum recording time of the datalogger was 3mins.

6.1.2 Bump and Braking Shock Factors The evaluation of the braking and shock factors with the use of the strain gauges was quite effective. The comparison of the actual loading in driving situations to the theoretical loading in these situations was extremely informative. The steady state cornering loads were as expected, illustrating the fact that little or no lateral grip generated on the inside wheel. However, the most interesting facts to come from this testing were that there was always a peak in the wishbone loading on the application of heavy braking. This shock peak in most cases being 2 times that of the steady state braking loading. Also, the circuit in which testing was conducted on this particular day had a few bumps and potholes, which inevitably got driven over at considerable speed. The pothole generated loadings 5 times that of the steady state in the pushrods and steering links.

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6.1.3 Fatigue Design The evaluation of fatigue frequencies for the design of the aluminium uprights and rockers (aluminium components designed against fatigue failure) was also to be established by the strain gauge data.

However, the strain gauge amplifier utilized, had a 3 Hz low pass filter incorporated. The problem being that the suspension fatigue da mage frequencies would be of a considerably larger than the 3 Hz filter. To cut a long story short the frequency spectrum plot was established only to realize that the results of the analysis were entirely useless.

The frequencies seen in this frequency spectrum plot for a cornering manoeuvre are ridiculous and entirely useless in the analysis of fatigue of any of the suspension components.

In future, if the loading frequencies are to be considered in design, frequencies of at least up to 100 Hz should be recorded. As the wheel rotational frequency at the vehicles top speed of 140 km/h is 23.6 Hz or 148.28 rad/s, giving a factor of four to overcome aliasing effects.

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6.2 Wishbone Construction Wishbones were constructed of two sections of circular hollow tubing, two rod end inserts, two rod ends and a spherical bearing housing with spherical bearing. The wishbone that has the pushrod/pullrod connected to it also has a connecting plate assembly.

Figure 6.2 Rear Lower Wishbone

The single most important consideration with the construction of the wishbones is that the direction of all members passes through the node of the spherical bearing housing, especially that of the pushrod/pullrod. The failure to do so will leave wishbone members in bending and consequently lead to excessive loading and failure. This fact was realized in 2001 with the failure of the rear lower wishbones in bending.

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6.2.1 Tubing Selection The Selection of tubing was based on the maximum loading both in tension and compression. The ma terial chosen for the wishbones was racetech 4130 chromoly, with a yield strength of 650 MPa. The list of available tubing in this series was restricted. ¾” tube was to be retained from the previous years designs to give the required stiffness of the suspe nsion wishbones, so the wall thickness required was all that needed to be determined. In tension, simple normal stress calculation was used and in compression, Euler’s buckling theorem was used to establish the onset of buckling. Considering the maximum loading of 10468.8 N in tension, 18183.3 N in compression and maximum rod length of 0.5 m in the 2003 geometry design incorporating the before mentioned shock factors, the tubing was chosen to be ¾” CHS with 0.9 mm wall thickness. This tubing size would see a stress of 278.622 MPa and with a critical buckling load of 40155 N, the onset of buckling was likely to occur.

6.2.2 Rod end Selection There are many different rod end manufacturers, producing products of various qualities. Due to budget restrictions in the project and the kind sponsorship of linear bearings for rod end and spherical bearings the choice was narrowed to 2 different series, PMG or RMT.

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6.2.2.1 Testing On the 29th of May 2003 rod end testing was undertaken in the materials department. The testing involved the two previously mentioned rod end bearings series (PMG and RMT). Intuition and previous design lead to the initial selection of 5/16” rod end bearings were considered the approximate size required for this application and as such were the subject of the initial rod end testing.

Specimen 1 (PM5G) was tested with a high tensile bolt through the eye, which was held by the instron at one end and 4 grade 8 nuts clamped by the instron at the other. See figure 1. The tensile bolt bent as the load was applied, and therefore the data from this test is uncharacteristic of the rod end bearing. The deflection measurement taken, by the instron machine, being that of the rod end and the shaft. This test was deemed a failure and consequently, the shaft deflection problem had to be rectified for further testing.

Figure 6.3 Initial Setup of Rod End Bearing Testing

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The other specimens (2-5) were tested with a hardened steel shaft. Also, the nuts at the other end were not simply clamped in the instron as before. Two (2) grade 8 nuts were placed on the end of the rod end bearing and in a frame that restrained only the nuts. See figure 2. The idea here was to only have the amount of thread in the testing as would be seen as the rod end would be used, so if thread failure was the critical mode of failure it would be seen in the testing.

Figure 6.4 Revised Setup of Rod End Bearing Testing

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Rod End Testing Results 3000

2500

2000

Load (kgf)

Specimen 1 (PM5G) Specimen 2 (PM5G) 1500

Specimen 3 (PM5G) Specimen 4 (RMT5X5) Specimen 5 (RMT5X5)

1000

500

0 0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

Displacement (mm)

Figure 6.5 Results of Rod end Bearing Testing

A smaller rod end in the RMT series could be considered. However, the smallest in this series is the 5/16” rod end. This being considered, the PM5G 5/16” UNF rod ends were chosen, as they are strong enough for this application.

6.2.3 Insert Design The tubing insert is the piece in the end of the tubing that the rod end bearings thread into. The inserts in the previous two years designs were silver soldered into the tubing. The design of the insert was of a considerable length due to the need for great surface area for the silver solder to be effective. A better solution, in my opinion, is to just weld in the insert. The insert can then be half the length and consequently half the weight. Consequently, as the tubing is 4130 chromoly, the insert being welded into the tubing was to be of a similar material. 4340 alloy steel was the only material available.

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Figure 6.6 Wishbone Insert

Note: Hexagonal section was used for ease of rod end adjustment.

6.3

Rocker and Upright Design

The rockers and uprights design were to be made from machined billet aluminium 2024T351. This alloy was chosen for it’s high strength relative to other aluminium’s, with a yield strength of 455 MPa, however, the presence of the alloying element copper makes it suitable for machining.

With the lack of fatigue frequency data, which was to be determined from the strain gauge data, another type of estimation of fatigue life of the components had to be evaluated. The estimation of a loading with an alternating amplitude equal to the mean amplitude was deemed an appropriate approximation. With the evaluation of this amplitude ratio in the appropriate fatigue strength diagram for 2024-T351 for infinite life gave a maximum stress value of 14.5 ksi or 100 MPa. -45-

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The parts were designed to be as light as possible whilst maintaining stress levels below 100 MPa in the stress concentrators in all loading situations.

Figure 6.7 Front Upright

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Figure 6.8 Rear Upright

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Figure 6.9 Front Rocker

Figure 6.10 Rear Rocker -48-

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6.4 Spring and Damper

6.4.1 Selection The selection of a mountain bike spring and damper unit is almost unanimous among formula SAE designs. This is not only because of the fact that they are almost the only thing commercially available that is suitable. They are compact, light weight, have interchangeable springs with varying stiffness and have sophisticated and adjustable damping in both compression and rebound.

Figure 6.11 Risse Racing Jupiter 5 Shock

The particular shock unit chosen for this application was a Risse Racing Jupiter 5. It has the following features: •

2 ¼ “ Travel

•

External Compression damping adjustment

•

External Rebound damping adjustment

•

Independent damping circuits for compression and rebound

•

Large piston allows for wide damping adjustment range -49-

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•

Gas charged eliminates cavitation

•

Adjustable preload for fine tuning sag

6.4.2 Damper Dynamometer The dampers were taken to the damper dyno at fulcrum suspensions to quantify their individual damping properties. This would allow informed damping set-up decisions. The damper dynamometer software produces various different graphs, however the graph of most interest to this project was the force versus velocity graphs. All 4 dampers were tested individually. The dampers had 14 clicks of compression adjustment and 6 clicks of rebound adjustment.

Figure 6.12 Damper Dyno Results of 1 Shock -50-

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Some interesting points to note are: •

All 4 different dampers yielded different results on the same settings

•

Compression damping is approximately linear

•

Rebound damping is approximately linear

•

The available range of damping available in these dampers will allow any where from 50% to 150% of critical damping on the racecar.

6.4.3 Spring Testing Spring testing was undertaken to establish the real spring rates of the springs obtained. This was deemed to be a sensible operation, to establish if the rates where as intended by the manufacturer, as correct spring rates are so critical to the vehicle performance.

Spring Rates 4 y = 0.0624x - 0.0162 3.5

3

Load (kN)

2.5 225 lb Risse y = 0.0411x - 0.0263

2

180 lb Thomas Marsh 225 lb Fox

y = 0.0342x + 0.0176

1.5

1

0.5

0 0

10

20

30

40

Displacement (mm)

Figure 6.13 Graph of Results

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Spring Rate

‘Formula SAE Suspension Design’ 180 lb Thomas Marsh

KN/mm 0.0342 N/m 34200 lb/" 195.3 Table 6.1 Actual Spring Rates Results

225 lb Risse

225 lb Fox

0.0624 62400 356.3

0.0411 41100 234.7

The Thomas marsh and fox spring rates were close to their designed value. However, the Risse spring was a lot stiffer than as designed.

6.4.4 X-Ray The placement of the spring and damper units in the vehicle required that a spherical bearing be placed in both ends of the damper unit. This presented a major problem as the bolt needed to secure the damper unit needed to be of adequate size to take the loading and yet a spherical bearing needed to be sourced that would fit into the tight hole without too much machining of the damper unit. The spherical bearing chosen was an SKF GE6C. It allowed a 6 mm high tensile bolt to be used and didn’t require too much machining for fitment, only 1 mm larger diameter. The machining was not blindly performed. X-Rays taken in the department of Veterinary Science illustrated just how close the machining came to interfering with internal passages and mechanical adjustment components.

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6.5 Steering Selection On researching steering rack availability of commercial steering racks. It was deemed economically viable to purchase a BRT steering rack at $400 US for a 750g it wasn’t even worth cons idering manufacturing a custom one.

Figure 6.14 Steering Rack in 2003 Racecar

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6.6 Accuracy & Adjustment One of the major considerations in the construction of the 2003 racecar was the accuracy to which the major components were constructed. Particular care was taken in the chassis construction to ensure that the suspension attachment points were in exactly, or as close to as possible, in the place as designed. Even the slightest inaccuracy of 1 mm can move the roll center by up to 5 mm.

Figure 6.15 Chassis on Jig for suspension Pickup Accuracy

All suspension parameters such as camber, castor, wheel alignment, ride height and toe can adjusted by the rod end thread tuning. In the case of the ride height and toe adjustments the pushrod/pullrod and toelink members function as turnbuckles with both a right and left hand thread rod end a for ease of adjustment on the car. -54-

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Kingpin inclination on this racecar is 0° and is not adjustable. Making the kingpin inclination adjustable is a difficult task, without compromising the design. In 2001, a rod end was used on the upper front wishbone connection to the upright. It was consequently in bending under braking and inevitably yielded during driving. The only other way to achieve adjustable kingpin inclination is to shim pack a bracket to the connection point of the upper wishbone. Both options compromise the lightweight, simple and effective design and therefore, kingpin inclination adjustment was sacrificed.

6.7 Component Placement The placement of inboard suspension components within the chassis is a difficult task and some of the main points to consider whilst doing so are: •

Aesthetics of Packaging

•

Linearity of movement

•

Chassis load paths

6.7.1 Rocker, Spring and Damper and Anti-roll bar Placement As in previous years the springs and dampers where placed with such a ratio to utilize all the damper travel. This is to maximize the efficiency of the damper and avoid potential cavitations. The other main concern of the rocker, spring and damper placement is the linearity of the movement of the spring compression to upright displacement. It is impossible to get exactly linear movement, however, close to linear movement can be achieved. The design methodology for this was to place the spring and damper unit with respect to the rocker in such a way that at full compression/travel they met at right angles. -55-

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This would mean that the movement would be as close to linear as possible and the suspension movement would always be getting progressively stiffer rather than softer. The rear anti-roll bar placement was placed using the same design considerations as the rocker, spring and damper placement.

Figure 6.16 Front Spring and Damper Placement

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Figure 6.17 Rear Spring/Damper and Anti-roll Bar Placement

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7. VEHICLE S ET UP

The racecar set up is just as important as the original design. It is in the set up that all the inaccuracies introduced in the construction of suspension components can be rectified to ensure the geometry as the vehicle is designed to have is in fact, the same as the geometry achieved on the vehicle.

A wheel alignment is the first set up operation that is to be performed on the vehicle.

The University of Queensland’s sponsorship with fulcrum suspensions allows free wheel alignment time on their wheel aligner. This particular wheel aligner performs laser measurements through many steering angle operations to determine suspension alignment.

Through the use of mathematics and some logical thinking, all links that are out in the alignment of the suspension can be rectified in 1 adjustment iteration. The wheel alignment is performed by the adjustment of rod end bearings at the end of all members as discussed in the previous chapter.

Some interesting points to note about this wheel alignment set up are: •

Toe is extremely sensitive and must be adjusted after all changes.

•

There are no chassis alignment measurements - chassis alignment to straight suspension still questionable.

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A well-designed suspension system will have plenty of room for adjustment. The possibilities for adjustment allowable in this racecar are: •

Ride Height

•

Castor

•

Camber

•

Toe

•

Spring Rates

•

Rear Anti-roll Bar Rates

•

Compression Damping

•

Rebound Damping

•

Anti-Dive Geometry

•

% Ackerman on steering

•

Tire Pressure

Each setup change has an effect on all other suspension parameters and this must be considered whilst setting up the racecar.

For example, a ride height change will change the roll centres; Tire pressure will change the roll centres; Castor will change the roll centres; spring rates will change anti-roll rates; even damper adjustment will affect anti-roll rates, however, only in the transient phase. These are only to name a few of the effects of a few possible changes.

Considering this and the fact that the driver’s style is just as an important and integral part of the vehicles performance, racecar setup is not an easy task and can be a very daunting task. It requires a lot of experience and the use of intuition and inference.

Unfortunately, due to the late completion of the racecar this year, not enough time has been allowed to go into too much detail with the setup of the 2003 university of Queensland formula SAE racecar as was originally intended with this thesis.

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8. RESULTS

The resulting weight distribution on the racecar was somewhat different to as designed. The original design was for a weight distribution of 50:50. However, the prediction of the weight distribution of a racecar that didn’t exist at the time is near impossible. The resulting weight distribution of 53.5:46.5 front heavy, with a 90 kg driver.

This problem had to be rectified before the racecar could be driven/tested successfully.

The spring rates had to be changed from those original designed for (180 lb front, 225 lb rear), to 235 lb front and 195 lb rear to achieve similar sprung mass frequencies.

Once this was rectified, extensive testing was undertaken.

The first testing session was conducted immediately following the racecar construction completion on the 8/10/03. This short testing session revealed a lot of minor problems as well as toe control issues. This was followed by a complete suspension and drive train strip and crack test. None of the parts had been damaged or showed crack initiation. A second and much longer testing session was conducted on 10/10/03.

The 2003 car pulls extremely hard with its spool differential and almost flat torque from 5000 rpm. Its turn in response was amazing, and with no suspension tuning other than a comprehensive wheel alignment at this stage the results were more than pleasing. The vehicle wasn’t perfectly balanced, being slightly prone to over steer on power application. However, at this stage the rear anti-roll bar had not been implemented.

With preliminary testing completed the evaluation of the rear toe control and inevitably redesign and reconstruction of toe link mounting location and toe link. The testing was simple and involved the application of a torque to the wheel whilst measuring the

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deflection through use of a wheel-mounted laser. The results of the testing revealed that toe control was in fact an issue.

Rear Toe Deflection 1.8 1.6 1.4 1.2 1 Toe (deg) 0.8 0.6 0.4 0.2 0

y = 0.0054x + 0.3192

0

50

100

150

200

250

Torque (Nm)

Figure 8.1 Rear Toe Deflection

With the steady state cornering torque on the wheel in the order of 100 Nm which relates to a deflection of almost 1 deg, which is outrageous. The problem or reason for the excessive deflection was the small moment arm to the toe link from the kingpin of 50 mm. The most suitable solution to this problem was to relocate the toe link further up the upright and chassis, where a much larger moment could be achieved without clashing.

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Figure 8.2 Toe Control Solution

The implementation of the rear anti-roll bar and fixed rear toe control for a testing session on the 19/10/03 saw a dramatic improvement in the cornering ability of the racecar. The performance of the 2003 formula SAE racecar is astounding. In comparison to the 2001/2002 racecar it is in another league altogether. With myself as driver skid pad testing was undertaken. The racecar was setup as it would be in competition, however, with formula ford tires instead of Hoosier slicks. The skid pad was of an inside diameter of 15.25 m as in the competition, however, on undulating off camber asphalt. Consistent times of 5.7s a lap were achieved. Evaluation of the average lateral acceleration for this yielded a result of 0.99 g’s. Considering that the formula ford tires are no where near as soft a compound as the Hoosier R25A compound tires as used in previous years and to be used in this years competition, and that with the Hoosier tires the 2002 racecar achieved a maximum of 1.14 g’s, as seen in Riseley [6], the resulting performance is going to be very interesting when the Hoosier R25A compound tires are used. -62-

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The vehicle appears to be performing as designed, sometimes elevating the rear inside wheel on corner entry. The camber of all wheels seem to be optimal with tire temperatures, as monitored in the pits, a consistent 41°C across the front tires and 43 °C across the rear.

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9. CONCLUSIONS & R ECOMMENDATIONS

In conclusion, the 2003 formula SAE racecar has proven to be quite successful so far in testing. The team spirits are high for the upcoming competition with many testing days scheduled in the month left before the competition.

With regard to the outcome of the suspension design, the results so far look promising and so far the design is quite successful. There is no right or wrong answer in suspension design. All suspension parameters have a unique relationship with all the others and as such they must be addressed as a complete suspension package. With more time tuning these suspension parameter the car will hopefully go on to be competitive at the formula SAE competition in Adelaide on the 4th to 7th of December this year.

In all honestly, the successful use of data acquisition systems on racing cars is difficult to achieve. Measuring the shock position is good for calculating wheel weights and developing shock speed histograms for shock setup. Tire temperature sensors can tell if the tires are cambered excessively or over/under inflated. Strain gauges on the suspension members can help evaluate the shock and bump loading of the member as well as possibly evaluated fatigue loading frequencies for component design.

Other than this, the evaluation of other parameters with data acquisition is of not such a straightforward manner.

I guess the point is that the problem is not only in how do you achieve accurate measurement of other parameters, but also what changes should be made upon knowing such information. For example, if optical slip angle sensors were to be used in the future, they can only be used for academic purposes, considering it’s already known that pro Ackerman works best on formula SAE cars. The result will tell you that you should be running around 70%

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Ackerman for best performance in the steady state. So the point being why measure it if it’s not going to prompt a setup change in the car, or achieve anything towards the cars performance.

The driver is just as much an integral part of the cars performance as the car itself and must be treated as such. Often driver feedback is a lot more useful than any data ever could be.

The use of the program suspension analyzer has some potential problems. It does not take into account tire deformation whilst simulating vehicle steady state dynamics. The static deformation of tire can be compensated for, however, a certain degree of inaccuracy is introduced in the roll center, camber analysis without considering the change in tire deformation in the various dynamic situations being modeled.

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10. B IBLIOGRAPHY 1. Gillespie, T.D.,1992, Fundamentals of Vehicle Dynamics, Society of Automotive Engineers, Warrendale, Pa.

2. Juvinal, R.C., Marshek, K.M., 2000, Fundamentals of Machine Component Design (3rd Edition), New York, John Wiley and Sons.

3. Manhire, O., 2001, Suspension Geometry Design of the 2001 University of Queensland Formula SAE Racecar, BE thesis, University of Queensland.

4. Maria, P., 2001, Suspension Component Design for Formula SAE, BE thesis, University of Queens land.

5. Milliken, W.F and Milliken, D.L., 1995, Racecar Vehicle Dynamics, Society of Automotive Engineers, Warrendale, Pa. USA

6. Riseley, C., 2002, Suspension Optimisation of the 2001 University of Queensland Formula SAE Racecar, BE thesis, University of Que ensland.

7. Smith C., 1996, Drive to Win , Carroll Smith Consulting Inc, Palos Verdes Estates, Ca USA.

8. Smith C., 1978, Tune to Win , Aero Publishers Inc, Fallbrook, Ca USA.

9. Smith C., 1975, Prepare to Win , Aero Publishers Inc, Fallbrook, Ca USA.

10. 2003, Hoosier Racing Tires Homepage, [Online], , [Accessed January 15 2003]

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APPENDIX A FSAE-A Design Spec Sheet

2003

Competitors: Please replace the sample specification values in the table below with those appropriate for your vehicle and submit this to with your design report. This information will be reviewed by the design judges and may be referred to during the event. --Please do not modify format of this sheet. Common formatting will help keep the judges happy! --The sample value s are fictional and may not represent appropriate design specs. --Submitted data will NOT be made public or shared with other teams.

Car No University

41 University of Queensland

Dimensions

Front

Rear

Overall Length, Width, Height Wheelbase

1525 mm

Track

1200 mm

1100 mm

Weight with 68 kg driver

154 kg

154 kg

Suspension Parameters Suspension Type

Tyre Size and Compound Type Wheels Design ride height (chassis to ground) Center of Gravity Design Height Suspension design travel Wheel rate Roll rate Sprung mass natural frequency (in vertical direction) Jounce Damping Rebound Damping Motion ratio Camber coefficient in bump

Front

Rear

Unequal length A-Arms. Pull rod actuated spring/damper unit 20.7x6-13 Hoosier R25A 3 Pce, Mag Centre 13"x6"-43mm o/s 40 mm

Unequal length A-Arms. Push rod actuated spring/damper unit. 20.7x6-13 Hoosier R25A 3 Pce, Mag Centre 13"x6"43mm o/s 40 mm

300 mm above ground 26 mm jounce/ 26 mm rebound 19.5 N/mm 0.86° / g, without anti-roll bars 2.76 HZ 90% of critical damping @ 30mm/sec 62% of critical damping @ 30mm/sec 0.9:1 0.5° / 10 mm bump

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26 mm jounce/ 26 mm rebound 25.4 N/mm 3.15 Hz 56% of critical damping @ 30mm/sec 54% of critical damping @ 30mm/sec 0.96:1 0.6° / 10 mm bump

Daniel Raymond Burt Camber coefficient in roll Static Toe and adjustment method Static camber and adjustment method Front Caster and adjustment method Front Kingpin Axis Kingpin offset and trail Static Ackerman and adjustment method

Anti dive / Anti Squat

Roll center position static Roll center position at 1g lateral acceleration

Steering System location

‘Formula SAE Suspension Design’ 0.42°/g 0mm toe out adj. by tie rods 0° by inboard rod ends 7° adjustable by rod ends 0° non-adjustable 30 mm offset, 32 mm trail 100% Ackerman @4.8m turn diameter Adjustment by interchangable steering link arms AD 10% @ 1.5 g's, adjustable by spacers ( 0%-10% ) 35 mm above ground, CL of the car 35 mm above ground, moves 6.35 mm toward inner wheel (with no steer), central with 15m turn diameter, 6.35mm toward outer wheel with 7m turn diameter Rear steer, not in line with steer

Brake System / Hub & Axle Rotors

Master Cylinder Calipers Hub Bearings

Upright Assembly

Axle type, size, and material

Front

0.38°/g 0mm toe in adj. by toe links 0° by inboard rod ends

0% (parallel)

38 mm above ground, CL of the car 38 mm above ground, moves 2.8 mm toward outer wheel

lower A-arm, yet no bump

Rear

Student designed, Student designed, lasercut lasercut from bis80 steel, from bis80 steel, spool hub mounted, 230mm mounted, 220mm dia. dia. PBR, Mechanical bias bar, for balance. 44.5 mm Wilwood Billet 44.5 mm Wilwood Billet Dynalite Dynalite (2) off SKF6006 (2) off SKF6007 (30/55/13), deep groove, (35/62/14), deep groove, 55mm spacing 40mm spacing CNC 2024_T351, "boltCNC 2024_T351, 60mm on" steering arms (adj toe link ackerman), 60mm toe link Stationary axle, 4340 HT Stub axle, 4340 HT 46 Rc. 46 Rc, Hollow 30mm OD Hub, 4340 HT 46 Rc , 20mm ID. Hub 2024_T351.

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Ergonomics Driver Size Adjustments Seat (materials, padding) Driver Visibility (angle of side view, mirrors?) Shift Actuator (type, location) Clutch Actuator (type, location)

Fixed seat and steering wheel. Pedal box adjust fore and aft 50mm Carbon fibre base, adapted to driver with individual padding (if req) 200° side visibility, rear view mirrors on cockpit sides L/H electromotive 'button' shifter Custom hand lever, below steering wheel.

Instrumentation

Frame Frame Construction Material

Steel tube space frame with bonded aluminium panels 4130 Chrome Moly tube

Joining method and material Targets (Torsional Stiffness or other) Torsional stiffness and validation method Bare frame weight with brackets and paint Crush zone material

TIG welded, 100% MPI to AS 1554.5 SP

Crush zone length Crush zone energy capacity

Powertrain Manufacture and Model

1998 Honda CBR600 F3 4 cylinder, custom dry sump, with integral scavenge pump.

Displacement Fuel Type Induction Max Power design RPM

599 cc Optimax Atmospheric induction 10,000 rpm. 56.5Kw (76Hp)

Max Torque design RPM Min RPM for 80% max torque

9,000 rpm. 57Nm (42ft.lbs) 5,850 rpm

Effective Intake Runner Length Effective Exhaust runner length

308 mm. 35mm ID. 690 mm Primary, 31.8 mm ID. 185mm Secondary, 44.4mm ID . 4-2-1 equal length, Student designed/built fuel injection, sequential IAT, CTS, CPS, TPS, EGO Above ports and pointing at back of inlet valves

Exhaust header design Fuel System (manf'r) Fuel System Sensors Injector location Intake Plenum volume

1575 cc (Butterfl y to airhorns). "Symetric" PlenumRunners. 12.0 :1 3 bar (static), EV1, 150g/min injectors MoTec M4

Compression ratio Fuel Pressure Ignition Timing Coolant location

System

and

Radiator

(2) Side pod mounted radiators, with electric fans, and water pum ps.

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Daniel Raymond Burt Fuel Tank Location, Type Muffler

‘Formula SAE Suspension Design’ Floor mounted aluminum tank between seat and firewall Carbon fibre canister.

Drivetrain Drive Type

Chain #520

Differential Type Final Drive Ratio Vehicle Speed @ max power (design) rpm 1st 2nd 3rd

Spool Diff 4.00 (13:52)

4th 5th

97 km/h 111 km/h

6th Half shaft size and material Joint type

123 km/h

45 km/h 65 km/h 81 km/h

GKN 21mm OD GKN

Aerodynamics (if applicable) Front Wing (lift/drag coef., material, weight) Rear Wing (lift/drag coef., material, weight) Undertray (downforce/speed)

N/A

Wing mounting

N/A

N/A N/A

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APPENDIX B Roll Centre Analysis

Front Geometry in Suspension Analyzer

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Front Camber Gain with 1° of roll (All measurements are in degrees)

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Front Camber Gain with 1” of dive

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Rear Geometry in Suspension Analyzer

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Rear Camber Gain with 1° of roll (All measurements are in degrees)

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Rear Camber Gain with 1” of dive

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Formula SAE Suspension Design

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