1. EDITH COWAN UNIVERSITY – SCHOOL OF ENGINEERING
Design and Manufacture
of a Formula SAE Engine
Bachelor of Engineering (Mechanical) -
Thesis
Thomas James Ayres – ST-10083916
November 2013
Supervised by Dr Nando Guzomi and Dr Kevin Hayward
2. 2
Abstract
In 2012 and 2013, the Edith Cowan University Formula SAE team investigated power-train
options that were available to replace the Honda CBR-600-RR, and decided to design and build
their own bespoke Formula SAE engine. This engine, designated the ER-600-C1, was based
around internal components and cylinder head taken from the 2006 Honda CBR-600-RR
motorcycle engine.
This report outlines the motivation for the decision to undertake this project, and documents
the process of design and manufacture of the major components for the ER-600-C1 engine.
Figure 1: The ER-600-C1
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Copyright and Access
Use of Thesis:
This copy is the property of Edith Cowan University. However the literary rights of the author
must also be respected. If any passage from this thesis is quoted or closely paraphrased in a
paper or written work prepared by the user, the source of the passage must be acknowledged
in the work. If the user desires to publish a paper or written work containing passages copied
or closely paraphrased from this thesis, which passages would in total constitute an infringing
copy for the purposes of the copyright act, he or she must first obtain the written permission
of the author to do so.
4. 4
Declaration
I certify that this thesis does not, to the best of my knowledge and belief:
(i) Incorporate without acknowledgement any material previously submitted for a degree
or diploma in any institution of higher education;
(ii) Contain any material previously published or written by another person except where
due reference is made in the text; or
(iii) Contain any defamatory material.
Name: Thomas James Ayres
Signature: Date: 7/11/2013
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Acknowledgements
After almost six years at University studying first the Bachelor of Technology Motorsports
course and the moving into Mechanical Engineering, it is hard to believe that it is all coming to
an end. To all those who have supported me throughout this journey, I would like to sincerely
thank.
To my family, thank you for supporting me through University and putting a roof over my
head. Without your support and encouragement, life at University would have been so much
more difficult.
To the ECU Formula SAE team, thank you for putting up with me and allowing me to be a
member of the team for the past few years. I have learnt so much about engineering and race
cars in the team during my time in the team and will sorely miss being in the workshop with
you guys.
To my Faculty Advisors and mentors through this project Dr. Kevin Hayward, Dr. Nando
Guzomi and John Hurney, thank you so much for your support and advice through this project
and with all the other projects you have helped me with. Your ongoing support for the Formula
SAE team is amazing and the success that the team has achieved over the years could not have
happened without you.
Finally, many thanks to the sponsors and companies who have provided assistance with the
design and manufacturing involved with this project. Your contributions are greatly
appreciated.
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Table of Contents
Abstract........................................................................................................................................2
Copyright and Access ...................................................................................................................3
Declaration...................................................................................................................................4
Acknowledgements......................................................................................................................5
Table of Figures............................................................................................................................9
Chapter 1 – Introduction............................................................................................................12
1.1 Background ................................................................................................................12
1.2 Report Contents .........................................................................................................13
Chapter 2 – Literature Review ...................................................................................................13
2.1 Regulations.................................................................................................................13
2.2 Honda CBR-600-RR Engine .........................................................................................14
2.3 Review of Competitors’ Engine Choices .....................................................................16
2.3.1 Single Cylinder Engines.......................................................................................16
2.3.2 Twin Cylinder Engines.........................................................................................17
2.3.3 Four Cylinder Engines.........................................................................................17
2.3.4 Other Engine Choices .........................................................................................18
2.4 Other Custom FSAE Engines.......................................................................................18
2.4.1 Western Washington V8.....................................................................................18
2.4.2 Melbourne University 2 Cylinder........................................................................20
2.4.3 Auckland University............................................................................................21
2.4.4 Mahle 3 Cylinder Engine.....................................................................................22
Chapter 3 – Design and Construction of the ER-600-C1 Formula SAE Engine ............................24
3.1 Rationale ....................................................................................................................24
3.2 Packaging ...................................................................................................................25
3.3 Team Responsibilities.................................................................................................29
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8.3 ECU.............................................................................................................................91
8.4 Wiring Loom...............................................................................................................91
8.5 Alternator and Starter Motor.....................................................................................91
8.6 Battery........................................................................................................................91
Chapter 9 – Engine Internal Components ..................................................................................92
9.1 High Compression Piston and Con-Rod Selection.......................................................92
9.2 Stock Honda CBR-600-RR ...........................................................................................93
Chapter 10 – Recommendations................................................................................................95
10.1 Mass Reduction..........................................................................................................95
10.2 Performance Modifications........................................................................................96
10.3 General Improvements ..............................................................................................96
Chapter 11 – Conclusions...........................................................................................................97
Bibliography...............................................................................................................................98
Appendix A – Engine Block Images...........................................................................................104
Appendix B – Motec Wiring Termination Tables......................................................................108
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Table of Figures
Figure 1: The ER-600-C1...............................................................................................................2
Figure 2: The Honda CBR-600-RR...............................................................................................16
Figure 3: The Western Wahington University Viking 30 V8........................................................19
Figure 4: The Western Washington University V8 engine installed in the Viking 30 Formula SAE
car ..............................................................................................................................................20
Figure 5: The turbocharged variant of the WATTARD Formula SAE engine ...............................21
Figure 6: The University of Auckland single cylinder engine and transmission ..........................22
Figure 7: Layout of the Honda CBR600RR engine in F-SAE vehicle.............................................27
Figure 8: Layout of ER-600-C1 engine in F-SAE vehicle ..............................................................27
Figure 9: Comparison between the 2012 ECU F-SAE car with Honda engine and the 2013 ECU F-
SAE car with the ER-600-C1 engine............................................................................................28
Figure 10: Photograph comparing the Honda CBR-600-RR and the ER-600-C1 side by side (note
that the Honda engine also would have a chain driven final drive when installed in a car).......28
Figure 11: Preliminary design of the cylinder bore region of the engine block..........................32
Figure 12: FEA results showing a cross section view of the stress plot ......................................35
Figure 13: Main cap FEA results showing stresses over 90MPa only .........................................35
Figure 14: Drawing of the clutch cover design ...........................................................................37
Figure 15: Photograph of the finished clutch cover ...................................................................37
Figure 16: Clutch cover FEA stress plot (clutch actuation loads)................................................38
Figure 17: Drawing highlighting the major features of the gearbox cover.................................41
Figure 18: Gearbox cover FEA stress plot for the spool bearing loads .......................................42
Figure 19: Gearbox cover FEA stress plot for the final drive pinion loads..................................43
Figure 20: Gearbox cover FEA stress plot for braking loads .......................................................44
Figure 21: Photograph of the gearbox cover (outside) ..............................................................45
Figure 22: Photograph of the gearbox cover (inside).................................................................45
Figure 23: Alternator assembly exploded view..........................................................................46
Figure 24: Photograph of the alternator cover (inside)..............................................................47
Figure 25: Photograph of the alternator cover (outside) ...........................................................48
Figure 26: Photograph of the alternator cover with stator coil mounted ..................................48
Figure 27: Drawing of the sump showing the various features..................................................50
Figure 28: Photograph of sump (inside).....................................................................................50
Figure 29: Photograph of sump (outside) ..................................................................................51
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Figure 30: Photograph of cam cover (outside)...........................................................................52
Figure 31: Photograph of cam cover (inside) .............................................................................52
Figure 32: Graph of car speed against engine speed for final drive ratio design........................55
Figure 33: Illustration of various gear dimensions [33]..............................................................56
Figure 34: Bending stress S-N histogram for the final drive gear ...............................................57
Figure 35: Bending stress S-N histogram for the final drive pinion ............................................58
Figure 36: Contact stress S-N histogram for the final drive gear................................................58
Figure 37: Contact stress S-N histogram for the final drive pinion.............................................59
Figure 38: Technical drawing of the final drive pinion ...............................................................60
Figure 39: Technical drawing of the final drive gear ..................................................................61
Figure 40: Photograph of the final drive gear pair .....................................................................61
Figure 41: Gearbox assembly render .........................................................................................62
Figure 42: Gearbox assembly exploded view .............................................................................64
Figure 43: Gear selector barrel assembly...................................................................................65
Figure 44: Gear selector barrel...................................................................................................67
Figure 45: Spool assembly exploded view..................................................................................69
Figure 46: Spool FEA stress plot (side A) ....................................................................................70
Figure 47: Spool FEA stress plot (side B) ....................................................................................70
Figure 48: Photograph of spool and final drive gears.................................................................71
Figure 49: Photograph of spool assembled with rear brake rotor .............................................72
Figure 50: Honda CBR-600-RR clutch lever mechanism .............................................................74
Figure 51: Cross sectional view of the clutch and hydraulic slave cylinder ................................76
Figure 52: Photograph of clutch assembled with slave cylinder in a test setup.........................77
Figure 53: Gear selector assembly showing detent mechanism ................................................78
Figure 54: Oil system image, showing the major components...................................................79
Figure 55: Photograph of the copper oil lines being test fitted in the engine block...................81
Figure 56: Gearbox oil spray bar test apparatus ........................................................................82
Figure 57: Gearbox oil spray bar test sample (0.6mm diameter jet)..........................................83
Figure 58: Photograph showing test in progress (note the concentrated oil jet).......................84
Figure 59: Graph of gerbox oil jet test results, flow rate vs nozzle cross sectional area ............85
Figure 60: Oil filter/cooler manifold assembly exploded view ...................................................86
Figure 61: Drawing of the crankshaft main cap showing oil channels........................................87
Figure 62: Photograph of the main cap (note oil channels) .......................................................88
Figure 63: Photograph of engine block crankshaft journal oil channels and piston sprays........88
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Figure 64: Solid state power distribution module in protective housing ...................................89
Figure 65: Solid state power distribution module circuit board.................................................90
Figure 66: Photograph of high compression Carrillo con rod and JE piston ...............................92
Figure 67: High compression JE pistons with oil rings................................................................93
Figure 68: The Honda CBR-600-RR crankshaft (also used in the ER-600-C1)..............................94
Figure 69: Engine block showing the lubrication system in red ...............................................104
Figure 70: Engine block showing the water jackets/cooling system in blue.............................104
Figure 71: Engine block and gearbox cover assembly showing how the gear shaft can be
assemble within the engine block............................................................................................105
Figure 72: Engine block assembled with major internal components (alternator side) ...........105
Figure 73: Engine block assembled with major internal components (clutch side)..................106
Figure 74: Photograph of the engine block ..............................................................................106
Figure 75: Photograph of engine block (inside)........................................................................107
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Chapter 1 – Introduction
1.1 Background
Formula SAE (F-SAE) is a competition where teams from approximately 500 universities around
the world design and build their own formula style open wheel race cars and compete against
each other in a number of regional competitions. The teams compete against each other in a
variety of different categories including vehicle design, vehicle performance, vehicle cost, fuel
economy, and with a business presentation [1], [2].
Edith Cowan University has been competing in Formula SAE Australasian competitions since
2008. Every year ECU has improved the design and performance of its race cars, and the
results in competition have mirrored this improvement with the team achieving second place
out of around thirty universities in the Australasian competition in 2012.
For the five previous cars that Edith Cowan University have built, the engine used was the four
cylinder 600cc 2006 Honda CBR-600-RR motorcycle engine. This engine was used for a number
of reasons including:
The engine has the largest capacity allowed in Formula SAE [2]
The engine produces high power for its size
The engine has relatively good power delivery over a wide rev range
The Honda engine has proven to be reliable for the team
Parts and replacement engines are readily available
Having used the Honda CBR-600-RR in previous years, the ECU team have developed
technologies that adapt the engine for use in F-SAE vehicle which can be carried on
and developed from year to year. Therefore changing engines would require this
process of development to start again.
During 2011 and 2012, the team have been finding it more and more difficult to make
significant improvements to the car working around the Honda engine. Reasons for this
include:
Difficulty packaging the Honda within the desired envelope of the car due to its
physical size and the need for a chain drive
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The exhaust exits the engine towards the front of the car resulting in difficulty in
packaging the exhaust system, and requiring the addition of heat shielding to the
chassis
The mass of the Honda engine and its associated subsystems is relatively high
The engine has a high vertical centre of gravity
To endeavour to find an alternative engine, which allows the evolution and improvement of
the vehicle concept behind the Edith Cowan University race cars, research was carried out
investigating alternative engine options. As a result of this investigation, the ECU F-SAE team
decided to design and manufacture their own bespoke engine which is based on parts from
the 2006 Honda CBR-600-RR engine.
1.2 Report Contents
This report provides an insight into the research carried out and the reasoning behind this
project of building a bespoke engine. Details of the various parts which make up the engine are
included in this report, with particular attention paid to the components and systems which
the author was directly involved with the design and/or manufacture of. A recommendations
section at the end of this report gives an insight for future teams developing this engine
concept into ways the author believes the ER-600-C1 may be improved.
Chapter 2 – Literature Review
The process of making the choice to design and build the ER-600-C1 began with reviewing the
options available to the team and analysing the choices that other teams have made regarding
their engines. The information gathered about alternative engine choices to help the team
make their decision is presented in this chapter.
2.1 Regulations
While teams are free to choose from a variety of different engines in Formula SAE, the
regulations for Formula SAE have several specific requirements for engines which limit these
choices. The significant aspects of these regulations are summarised below [2].
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Engines must have a capacity of less than 610cc
Engines must breathe through a 20mm inlet restrictor (Gasoline engines)
The engine must use a four stroke cycle
The engine must have an onboard electric starter
For forced induction engines, the inlet restrictor must be placed up-stream of the
turbocharger or supercharger
Any chain or belt drives must have a protective shield made from a minimum of
2.66mm thick steel
The noise output of the engine must not exceed 110dBA (fast weighting)
2.2 Honda CBR-600-RR Engine
Since the first Edith Cowan University Formula SAE car in 2008, the team has used a Honda
CBR-600-RR motorcycle engine as a power plant (see Figure 2). The Honda engine has proven
to be a successful engine for the team, with the team finishing second in Formula SAE
Australasia in 2012 to score a world ranking of twenty-third. The team has completed both
endurance events in all Australasian Formula SAE competitions since 2009.
The Edith Cowan University team, and many other successful Formula SAE teams, have
continued to use the Honda CBR-600-RR engine for a variety of reasons:
It has the largest capacity permitted in Formula SAE competition of 600cc
It produces relatively high power in Formula SAE configuration without significant
modification
It produces power over a wide range of engine speeds
It has proven to be reliable
Engines and parts are readily available
The ECU team has developed a number of ancillary engine systems which have been
improved year after year
There are however, disadvantages associated with the Honda CBR-600-RR engine which have
become more apparent as the ECU racing team have evolved the concept of their cars, and
points allocated in Formula SAE competition for fuel economy have been increased while
points for acceleration of the car have decreased [2]. The main disadvantages associated with
the Honda engine which the team have encountered are listed below.
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The engine is difficult to package compactly within the car due to its physical size and
the need for an external differential and chain drive
The engine exhaust system results in packaging difficulties due to the forward exhaust
exits and resulting heat dissipation issues
The engine mass, with its associated sub-systems is relatively high compared to other
engine options
The engine uses more fuel than other potential engine choices
The engine has a high crankshaft centreline and COG
Specifications of the standard Honda CBR-600-RR engine are included in the table below [3].
Table 1: Honda CBR-600-RR motorcycle engine specifacations
Capacity 599cc
Fuel Type Unleaded Petrol
Cylinders Inline 4
Bore/Stroke 67.0mm/42.5mm
Valve-train DOHC, 2 inlet, 2 exhaust valves per cylinder
Compression Ratio 12.0:1
Fuel Delivery Electronic fuel injection
Ignition Digital electronic ignition with individual coils
Cooling Liquid Cooled
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Figure 2: The Honda CBR-600-RR
2.3 Review of Competitors’ Engine Choices
A recent trend in Formula SAE following changes in the rules which reduced point allocations
for straight line acceleration and increased points for fuel usage [2] is for more teams to adopt
smaller capacity, fewer cylinder engines. The following sub-sections of this report review the
choices made by other Formula SAE teams regarding their engines and discusses the
advantages and disadvantages of these options in comparison with the Honda CBR-600-RR
engine.
2.3.1 Single Cylinder Engines
Many of the more successful teams in recent Formula SAE competitions have used single
cylinder engines. Smaller single cylinder engines have advantages over larger 600cc four
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cylinders such as being better for packaging in the vehicle, lighter than the larger four
cylinders, and using less fuel than the four cylinder engines. Major disadvantages of these
smaller engines are reduced power, reduced reliability, short power producing rpm range and
excessive vibration caused by the inherent imbalance of a single cylinder engine [4].
Teams which have been successful with different single cylinder engines include:
GFR (Global Formula Racing, an international partnership between the Duale
Hochschule University in Germany and Oregon State University, USA) – Honda
CRF450X
Monash University, Australia – KTM 450SXF
TU Graz, Austria – KTM EXC500/525
ETS, Canada - Yamaha WR450F
RMIT, Australia – Yamaha WR450F
Most of the common single cylinder engines used in Formula SAE are derived from Enduro
class off-road motorcycles and have similar characteristics and performance figures [5].
Perhaps the biggest advantage of using a single cylinder Enduro class motorcycle engine is the
saving in weight, which is quoted as being as much as 30Kg compared to a 600cc four cylinder
engine [5]. Another major advantage of a single cylinder engine is the simplified design of
intake and exhaust systems compared with four cylinder engines [6].
2.3.2 Twin Cylinder Engines
Twin cylinder engines have been integrated into Formula SAE race cars by several teams such
as the University of Texas at Arlington (UTA) in 2008[7], the US Naval Academy, the University
of Maine, and the South Dakota School of Mines and Technology[8]. The twin cylinder engine
primarily used in Formula SAE teams is the Aprilia 550cc SXV 77˚ V-Twin. The advantages of the
Aprilia V-Twin engines over a Honda CBR-600-RR engine are similar to the advantages of single
cylinder engines, with reduced weight and smaller physical size [7]. Being a highly stressed
engine, the major disadvantage with the Aprilia engine is the apparent lack of reliability, with
the engine having a reputation for having starter motor problems [9]. The Aprilias are also
difficult to source, with complete engines and spare parts relatively rare.
2.3.3 Four Cylinder Engines
The most common engine choices for Formula SAE teams are four cylinder 600cc Supersport
class motorcycle engines. Most of the engines of this class are manufactured by the “big four”
of the motorcycle companies, namely; Kawasaki (ZX-6R), Suzuki (GSXR-600), Yamaha (YZF-R6),
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and Honda (CBR-600-RR and R4) [5], [10]. Performance and characteristics of these engines are
similar [5], but the most commonly used and easiest engine to obtain is the Honda.
2.3.4 Other Engine Choices
Other less commonly used options for Formula SAE engines include snowmobile engines and
more common engines adapted for forced induction. Due to the difficulty in obtaining
snowmobile engines because of Edith Cowan University’s geographical location, no further
research was carried out on this option. Forced induction of more commonly available engines,
such as turbo-charging or supercharging, is an effective method of extracting more power from
smaller capacity engines [11].
Teams have been successful in adapting commonly available engines for forced induction, such
as the University of Sophia’s supercharged 4 cylinder engine [12], and Cornell University’s
turbo-charged Honda CBR-600-RR [13]. These projects have been successful in increasing the
power output and fuel economy of the original engines. Adapting a single cylinder engine for
forced induction could be an attractive prospect as some of the power deficit of the Enduro
class single cylinder engines could be reduced, while retaining the small engine size and weight
[14].
The major disadvantage of applying forced induction to engines in Formula SAE applications is
the potential for reduced reliability of the engine.
2.4 Other Custom FSAE Engines
In the past, other universities have developed their own bespoke engines for Formula SAE
competition with varying success. These engines are reviewed in the following sub-sections of
this report, highlighting the triumphs and failures in each project.
2.4.1 Western Washington V8
In 2001/2002 Western Washington University (WWU) manufactured a Formula SAE car, the
Viking 30 which featured a 554cc V8 engine (see Figure 3). The engine used cylinder heads and
pistons from two 4 cylinder 250cc Kawasaki motorcycle engines. A six speed transmission
taken from a Honda 600cc F1 motorcycle with a bespoke casing and final drive transmitted
power to the wheels. The engine and gearbox were fully stressed members with the rear
suspension mounted directly to the power-train. The WWU team designed and manufactured
the engine and gearbox casings from billet Aluminium, and also designed and manufactured
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the crankshaft [15]. Details of the Western Washington V8’s electrical charging, starting and
other sub systems are unknown.
Figure 3: The Western Wahington University Viking 30 V8
It is believed that the WWU V8 was relatively successful in that the engine functioned well and
produced relatively high power. The centre of gravity of the engine was low and being a
stressed member, the power-train likely had some weight advantage for the vehicle.
While the WWU V8 engine is an impressive feat of engineering, the conventional longitudinal
layout of the engine and transmission results in an engine package which takes up a relatively
large amount of space and unfortunately has no real advantage in terms of vehicle packaging
over a standard motorcycle engine. Figure 4 shows the WWU V8 installed in the Viking 30
Formula SAE car.
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Figure 4: The Western Washington University V8 engine installed in the Viking 30 Formula SAE car
2.4.2 Melbourne University 2 Cylinder
Melbourne University developed the Wattard engine, named after its chief designer William
Attard, for their 2003 Formula SAE car (see Figure 5). This engine was a 434cc in-line 2 cylinder
which was later turbocharged in 2004. The engine features duel overhead camshafts with four
valves per cylinder, a large capacity sump for minimal frictional losses, and a three speed
gearbox and chain drive to the rear wheels [16]. The engine was specifically designed to be
“...optimised for the needs of a Formula SAE car rather than a motorcycle” [17]. “The majority
of components were manufactured in-house (at Melbourne University), either specially cast,
fabricated or machined from billets” [18].
The Wattard engine succeeded in being a lightweight, high powered and well packaged engine
for a Formula SAE car. The engine was somewhat successful in 2003 and 2004 with the
Melbourne University team “completing the third fastest lap” in the endurance event, and
“matching the performance of all top four cylinder 600cc cars” in 2003 and winning the fuel
economy event in 2004 [17].
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Issues with the Wattard engine mainly relate to its poor reliability. It was suggested by Mauger
[5] that because so many of the engine components were custom made for the engine, the
Melbourne University team did not have sufficient time or resources to develop the engine to
a point where its reliability was satisfactory.
Figure 5: The turbocharged variant of the WATTARD Formula SAE engine
2.4.3 Auckland University
From 2009-2012 the University of Auckland have produced cars with custom single cylinder
engines based on the Yamaha WR450, YZF450, and WR450F off-road motorcycle engines. The
engines featured a four speed gearbox with gears taken from the Yamaha motorcycle gearbox,
and a transaxle style final drive with limited slip differential from a Yamaha Grizzly quad bike
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(see Figure 6). According to the University of Auckland; “The package significantly reduces
centre of gravity height, allowing a narrower track width and a more nimble car” [19].
The University of Auckland custom single cylinder engines succeeded in creating lightweight
and compact power-trains with low centre of gravity. The cars in which these engines were
installed, achieved a dry weight of 172Kg and were competitive with third place in the Skid-pad
event in 2009, third in the Autocross event in 2010, fifth in Autocross and Endurance in 2011,
and scoring consistently high Design event scores [19], [20].
Figure 6: The University of Auckland single cylinder engine and transmission
While Auckland’s custom engines were somewhat successful, they were plagued with
reliability issues. The engine suffered both problems with the engine and the gearbox. The
cause of this unreliability is possibly partly due to the internal components of the engine being
highly stressed. By making custom casings for already highly stressed components, and small
errors in design or manufacturing would be exaggerated.
2.4.4 Mahle 3 Cylinder Engine
The engineering company Mahle developed an inline 3 cylinder engine for the RWTH Aachen
Formula SAE team in 2003. The aim of the development of this engine was to showcase the
capabilities of the company [21]. The 609cc engine produces a quoted 60KW at 9,500rpm and
65N.m of torque at 7,000rpm, which is a respectable output for a restricted Formula SAE
engine [21].
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The Mahle engine was manufactured specifically for Formula SAE use from scratch, with all
engine internal components designed and manufactured for the engine. This level of
development is not yet a capability of the Edith Cowan University team.
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Chapter 3 – Design and Construction of the ER-600-C1
Formula SAE Engine
3.1 Rationale
Since its conception in 2008, the Edith Cowan University Formula SAE team has used a four
cylinder Honda CBR-600RR motorcycle engine as its power plant. Power has been transmitted
to the rear wheels by a chain drive to a differential unit which incorporates the drive shafts.
The design of the car has evolved and improved in each consecutive year since 2008 with the
packaging of the Honda engine and the drive train within the chassis envelope becoming more
compact and lightweight. By 2011/2012 the team reached a point where further improvement
of the packaging of the Honda engine was becoming more and more difficult therefore it was
decided that alternative power train options were to be explored.
From the point of view of the overall design of a Formula SAE vehicle, the requirements of the
power train are very specific.
The engine/power train should be of a minimal mass
The engine/power train should be reliable and serviceable by students
The power train should be as short as possible in the longitudinal direction to allow for
the vehicle to have a short wheel base and low polar moment of inertia (this is of
particular importance with new Formula SAE rules in 2013 stating that there must be a
minimum of 915mm between the seat back and the face of the pedals)
The engine should produce high power and torque throughout the rev range
The engine should be fuel efficient
The power train should have a low vertical centre of mass
After considering the power train options available, it was decided that the way to achieve the
best compromise between the requirements of the power train was to design and build a
bespoke engine adapting the internal components from the Honda CBR-600RR and
incorporating a final drive gear reduction and locked differential. This engine concept was
designated the ER-600-C1.
By building the engine around existing Honda internal components there are several
advantages:
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The number of parts to be designed and manufactured is decreased
There is a decreased risk of the engine having reliability issues
Time taken to design and build the engine is decreased
The Honda engine block could be used as a baseline for the design of the new block
Technologies developed by the team such as intake and exhaust systems, and high
compression piston/con rod assemblies can be carried over to the bespoke engine
Spare parts can be easily obtained
By building the ER-600-C1, the Edith Cowan University Racing team intends to combine the
advantageous peak power, smooth power delivery and reliability of a 600cc four cylinder
engine with the smaller package size, reduced weight and reduced fuel use of a smaller
capacity, single or twin cylinder engine.
While designing and manufacturing a bespoke engine can be considered a high-risk strategy
for the success of the team in competition in 2013, the technology and capabilities developed
by the team during the process of the development of this project will be valuable to the team
in years to come. The ER-600-C1, while fully intended to be a successful project is a crucial first
step in the future development of Formula SAE engines at Edith Cowan University.
3.2 Packaging
As stated previously, the design of the 2013 Edith Cowan University Formula-SAE vehicle’s
engine has been based around the use of Honda CBR-600RR motorcycle engine internal
components. The design for the Edith Cowan engine consists of a bespoke casing for these
internal components along with a final drive gear reduction, allowing the engine to be
efficiently packaged within the Formula-SAE vehicle.
Although other Formula-SAE engine options have been considered when designing the
bespoke engine, to illustrate the advantages of the design, comparisons will be made to the
standard Honda CBR-600RR engine. This is the engine which has been used by the Edith Cowan
University team in previous years and is also the engine of which many of the internal engine
components have been sourced. Some of the disadvantages of using the Honda CBR-600RR
engine for use in a Formula-SAE vehicle are listed below.
A chain drive is required to transmit power to the rear wheels.
Exhaust exits the engine towards the front of the car and the driver.
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Engine has a relatively high crankshaft location.
Cylinders are angled foreword, increasing the overall length of the engine.
The engine includes a six-speed gearbox, of which only three ratios are used.
To address the issues with the Honda CBR-600RR stated above, the ER-600-C1 engine features:
1. Vertical cylinder bores to minimise both the total length of the engine and the vertical
height of the crankshaft centreline;
2. Two gear ratios rather than the six in the Honda CBR-600RR (1st
and 3rd
gears);
3. An in-built final drive gear reduction and differential or spool to minimise the total
drive-train length;
4. Final drive gear ratio is optimised for the use of only two gears;
5. The cylinder head is rotated 180˚ relative to the Honda CBR-600RR so that the exhaust
exits rearwards;
As the final drive reduction in the ER-600-C1 features meshing gears rather than a chain driven
sprocket arrangement, the output drive to the wheels spins in the opposite direction in
relation to the crankshaft. To rectify this, the crankshaft and cylinder head was rotated 180˚
with the added benefit of allowing the exhaust to exit towards the rear of the car away from
the chassis and driver.
The sketches in Figure 7, Figure 8 and Figure 9 show a comparison between a Honda CBR-
600RR installation in a Formula-SAE car and the proposed layout of the bespoke Edith Cowan
engine in a similar vehicle. The contrast in overall size between the Honda CBR-600-RR and th
ER-600-C1 can be seen in the ptotograph in Figure 10.
27. 27
Figure 7: Layout of the Honda CBR600RR engine in F-SAE vehicle
Figure 8: Layout of ER-600-C1 engine in F-SAE vehicle
28. 28
Figure 9: Comparison between the 2012 ECU F-SAE car with Honda engine and the 2013 ECU F-SAE car with the
ER-600-C1 engine
Figure 10: Photograph comparing the Honda CBR-600-RR and the ER-600-C1 side by side (note that the Honda
engine also would have a chain driven final drive when installed in a car)
29. 29
To further enhance performance and to ensure reliability and compatibility of the engine to
suit use in a Formula SAE vehicle, various features are incorporated into the ER-600-C1 engine
design. These features include, but are not limited to:
An enlarged crank-case volume to reduce windage on the rotating engine components
Built-in chain drive and attachment points for Dailey Engineering multi-stage dry sump
oil scavenge and pressure pumps [22].
Mounting points for oil filter and oil/water heat exchanger
Oil galleries integrated with the engine block with feeds to the crankshaft journals,
gears, cylinder head, pistons, and gearbox bearings
A hydraulic clutch actuation slave cylinder to allow for a variety of cockpit clutch
actuation concepts
Brake calliper mounting bracket and rotor mounts for an unsprung, inboard braking
system
Hard-points for mounting the engine within the Formula-SAE vehicle
Multiple hard-points for mounting ancillaries such as oil tank, wiring, electronics,
coolant lines, external oil lines and other unforseen items
An internal combustion engine is a complex piece of machinery and comprises of a number of
different systems, parts, and assemblies. While many of the components of the Edith Cowan
University Formula-SAE engine have been sourced from donor Honda CBR-600RR engines,
there remain many components which need to be designed and manufactured or purchased to
complete the project.
3.3 Team Responsibilities
The project of designing and building the Edith Cowan University engine has been a team
effort. A list of the main engine team members is included below along with each member’s
general responsibilities in the project.
Sean Supiers – Design and manufacture of main engine block, post machining of main engine
block, official engine team leader
Tom Ayres – Formula SAE team co-technical director, design and manufacture of gearbox and
final drive components, design and manufacture of engine covers, co-design and manufacture
of oil system
30. 30
Cheng Chao Khor – Design of throttle and intake system, senior lathe machinist
Peter-John Grigson –Dynamometer development for ER-600-C1
Phillip Le – Formula SAE team co-technical director, research and purchasing
Didi Hardianto – Manufacture of intake
Alex Ayres – Manufacture of exhaust system
3.4 Cost Event
Part of the Formula SAE competition is to present a report which lists the components of the
Formula SAE vehicle and calculates the cost of the car based on standardised costs of
individual components. The point score from this cost event make up a possible 100 out of a
possible 1000 points from the entire Formula SAE competition, with the overall calculated cost
of the vehicle making up 40 of the 100 cost event points [2].
One advantage of entering the competition with a custom engine with inbuilt final drive, oil
pump and other components is that components such as these inbuilt parts do not need to be
costed in the cost event, which results in an increased point score for the car cost. The cost
saving (according to the cost event pricing) due to the use of the ER-600-C1 as opposed to a
Honda CBR-600RR engine with associated sub-systems has been calculated to be $1,500 to
$2,000 which, through analysis of past Australasian Formula SAE competitions, is estimated to
be worth 10 to 15 points.
While the direct competition point advantage of the ER-600-C1 is relatively modest, any
advantage is always welcome and the improved vehicle packaging made possible by the ER-
600-C1 has resulted in further cost event savings and additional cost event points.
31. 31
Chapter 4 – Engine Block & Covers
4.1 Engine Block Layout
The engine block itself is the most complex component of the engine. It must accurately locate
the moving parts within the engine in the desired layout. The engine block also incorporates
the cooling system and lubrication system in addition to providing adequate mating surfaces
for the cylinder head, sump and gearbox, alternator and clutch cover.
Design of the engine block has been a long and complex process with more than 90 design
iterations made before the final design was arrived at. The first stage of the process was to
reverse engineer the geometry of the Honda CBR-600RR engine block and the components
which were intended to be used in the bespoke engine such as the crankshaft, gear shafts and
clutch. The measurement of the Honda engine block and components were carried out using a
combination of manual measurements, and through the use of a coordinate measuring
machine arm (CMM). The general layout of the fundamental parts of the engine was then
decided with consideration to compact packaging, low centre of gravity, appropriate drive
shaft output height, and correct spacing between components.
The engine block has been designed to be machined from a solid block of aluminium on a CNC
milling machine. This manufacturing technique was chosen over casting for assurance of
homogenous material properties in the finished product, potential reduced manufacturing
costs for the small number of units required, and for relative ease of design. Milling the engine
block was chosen at the expense of potentially increased weight, less efficient use of material
and increased material wastage.
Computer modelling of the bespoke engine block was carried out in Solidworks computer
aided design software [23]. Three dimensional modelling began by producing the cylinder bore
portion of the engine which mates with the Honda cylinder head and locates the crankshaft
and pistons. This part of the engine block is critical for the correct function of the engine and
incorporates water jackets for cooling and oil feed and drains for the cylinder head.
32. 32
Figure 11: Preliminary design of the cylinder bore region of the engine block
Following the initial modelling of the cylinder bore part of the engine block the gearbox area of
the block was laid out. Consideration needed to be made to the appropriate height of drive
shaft outputs from the final drive and the final drive gear diameter and pinion/gear spacing. In
addition, the gear shaft centreline separations reverse engineered from the Honda CBR-600RR
engine had to be maintained and the correct axial alignment of the shafts ensured.
Once the layout of the fundamental engine components had been decided and modelled, then
consideration was given to how the engine could be designed to be easily assembled. Several
ideas were investigated of how this could be achieved including splitting the block into two
pieces along the centreline of the various shafts within the engine. This idea was however
rejected due to the shafts not being aligned along a single plane resulting in a block to be split
in a ‘V’ shape. This would have introduced a weakness into the engine block at the apex of this
‘V’, been complex and expensive to machine and difficult to achieve an oil tight seal around
the mating surface. The idea was therefore not pursued along with a number of other
concepts. The method of assembly finally decided upon was to have individual bearing caps for
each crankshaft journal, and a removable section of the gearbox casing which houses the
bearings for one end of the gear shafts, while the other side is located within the engine block.
This solution was found to be the most straightforward to manufacture, provides ideal sealing
surfaces, and allows assembly of the engine.
33. 33
Mounting points for the starter system and the oil pump were then located. These
components were constrained to how they may be located due to the decision to use the
standard Honda CBR-600RR starter motor and drive gears, and the Honda water pump drive to
power the aftermarket oil scavenge and pressure pumps.
The engine block incorporates oil galleries providing lubrication to the moving parts within the
engine. Galleries will be machined into the block such that oil is fed in from the pressure pump
via the water/oil heat exchanger and the sump, into an oil filter mounted to the block, and
then into a main gallery which provides feeds to the crankshaft journal bearings, sprays to the
bottom of the pistons to aid cooling, the cylinder head, a gearbox spray bar, and a feed to the
inside of the secondary gear shaft. The feed to the secondary gear shaft provides lubrication to
gearbox bearings and further lubrication to the gearbox through oil flowing through holed in
the gear shaft as it rotates. To reduce losses in power due to windage on the crankshaft [24],
the crankcase has been enlarged to allow a volume for oil to be displaced into.
Hard-points have been incorporated into the engine block design to facilitate the engine being
mounted to a vehicle. Mounting points are also provided for location of ancillary and
miscellaneous items to the engine block such as wiring, oil tanks, battery etc.
For photographs and computer generated images highlighting some of the various points of
interest and systems associated with the engine block, see Appendix A.
4.2 Crankshaft Main Bearing Caps
In order for the crankshaft to be installed in or removed from the engine block, it is necessary
for the casing around the crankshaft main bearings to be able to be split and removed. There
are two primary methods for this to be achieved. The first, method is for the engine block to
be split in two with one half of the crankshaft main bearing housings in the “top” part of the
engine block and for the other half of the bearing housings in the “bottom” part of the block.
The alternative method is for each of the main bearings to have their own individual caps
which bolt to the main engine block [25].
The method chosen for the ER-600-C1 main bearings was to have individual main bearing caps
unlike the Honda CBR-600RR engine due to concerns with the difficulty and cost associated
with manufacturing the alternative. Difficulties with manufacturing a split engine block would
have been concerned with the integration of oil galleries and machining accuracy of the
bearing bores. Due to the fact that the main engine block has the main bearings located on a
34. 34
different plane than the bottom of the block, it could potentially be difficult to deck the mating
surface of the main bearings in order to be able to remove more material from the bearing
bores without increasing their diameters. With individual caps however, the mating surface
can be easily (and individually) decked in order to remove more material from the bores.
Another reason for choosing individual main caps was to minimise the vertical height of the
crankshaft vertical centreline. The advantage of the split block concept is increased structural
strength and stiffness of the engine block structure supporting the crankshaft.
Once the decision was made to use individual main bearing caps rather than a split block
design, a rough size envelope was determined by examination of the Honda CBR-600RR engine
block. Identical bolt sizes and spacing were used in the bearing caps to the Honda engine. Also
reverse engineered from the Honda were the oil channels leading to and around the main
bearings. So that oil feeds could be supplied to the bearing caps from the internal oil lines from
the sides in order to minimise the vertical height of the system, a small protrusion was added
to the bottom a the bearing caps to accept 1/8 NPT pipe fittings. To increase the lateral
support of the bearing caps, hollow dowel pins were incorporated in the design which the two
bolts pass through.
Finite element analysis was carried out on the crankshaft main bearing caps to ensure that the
design could endure the loads generated through the operation of the engine at peak power.
To calculate reasonable loads for the analysis, calculations were based on peak cylinder
pressures of 1500psi and are included below.
Maximum force generated by each piston:
Where; F=force (N), P=peak cylinder pressure (Pa), A=cylinder area (m2
)
For the purpose of the analysis, the loads generated by the piston acceleration as the
crankshaft rotates are neglected due to these loads being balanced to a degree by the other
pistons in the engine. The calculated force of approximately 35KN was used in the analysis
even though this force would actually be shared by two crankshaft main journal bearings.
Overall the use of the 35KN load results in a conservative analysis.
The 35KN load was applied vertically downward to the bearing surface as a bearing load with a
sinusoidal distribution. The two surfaces at the top of the bolt counter bores were defined as
fixed.
35. 35
Figure 12 and Figure 13 show the results of the analysis. Figure 13 shows regions of the part
which are stressed over 90Mpa, and Figure 12 shows a cross section of the stress plot. The
peak stresses in the analysis which exceed the material yield stress are concentrated around
the bolt landing surfaces which were represented as fixed entities in the analysis and can be
assumed to be lower than the analysis suggests. The other relatively highly stressed regions
are concentrated around the bearing surface. In operation, the crankshaft will be supported by
an oil film and by the bearing shells which will both help to more evenly distribute the stresses
around the main bearing cap.
Figure 12: FEA results showing a cross section view of the stress plot
Figure 13: Main cap FEA results showing stresses over 90MPa only
36. 36
4.3 Clutch Cover
The clutch cover is a component which is located on the left hand side of the engine and allows
access to the clutch and starter systems. The clutch cover also locates various components of
the starter and clutch systems. A list of the functions and requirements of the clutch cover are
listed below.
1. Accurately locate the starter gear shaft and starter idler gear
2. Accurately locate the crank angle sensor
3. Allow the crank angle sensor wiring to pass through to the outside of the engine
4. Provide a pathway/mounting for the oil level sight tube
5. Provide a mounting for the hydraulic clutch slave cylinder
6. Resist loads generated by the activation of the clutch
7. Provide oil tight sealing surfaces with the engine block and the sump
8. Be of minimal mass
9. Provide external mounting points for undetermined components (such as wiring etc)
The general envelope of the clutch cover was determined by the design of the side of the
engine block. To determine the location of the various mounting points incorporated in the
design of the cover, accurate modelling of the relevant engine components was necessary.
Attachment of the clutch cover to the main engine block is through a series of M5 socket head
cap screws arranged in a pattern around the mating surface. Accurate location of the clutch
cover in relation to the main engine block is achieved through the use of M6 shoulder screws
incorporated into the bolting pattern. Both the M5 socket head cap screws and the M6
shoulder screws are threaded into the main engine block. Attachment to the sump is with M5
socket head cap screws, where the bolts are threaded into the clutch cover.
To provide stiffness and strength and to ensure that the cover could withstand the loads
generated by the activation of the clutch for minimal mass, internal ribs are included in the
design of the clutch cover radiating from the mounting point of the clutch hydraulic slave
cylinder.
The clutch cover was manufactured in-house at ECU on the Okuma CNC vertical milling centre.
The component was machined from 50mm thick 5083 aluminium plate. Figure 14 and Figure
15 show the design and a photograph of the clutch cover respectively.
37. 37
Figure 14: Drawing of the clutch cover design
Figure 15: Photograph of the finished clutch cover
To check that the cover could withstand cycles of loading from clutch applications, FEA analysis
was performed on the component. The loads applied in the analysis due to clutch application
38. 38
were calculated in two different ways. The first method was to measure the spring rate of the
clutch springs and to calculate the load required to compress the five springs the distance
required to activate the clutch. The second method was to calculate the load on the clutch
based on measurement of the load on the clutch lever multiplied by the mechanical advantage
achieved through the lever/cable system used on the 2012 ECU Formula SAE car.
The load generated by the actuation of the clutch was determined to be approximately 1300N
from each of the two methods of calculation. For the finite element analysis, a load of 1500N
was applied to the circular area on the outside of the cover which corresponds to the
dimensions of the clutch slave cylinder. The mating flanges of the clutch cover were assumed
to be fixed in all directions for the analysis. Figure 16 below shows the results from the analysis
performed in the Solidworks F.E.A. software package.
Figure 16: Clutch cover FEA stress plot (clutch actuation loads)
As shown in Fig XX, stresses generated in the clutch cover by the action of the slave cylinder
are effectively dissipated and the component is under relatively low stress in relation to the
material yield strength of around 255MPa. There is a region of the component where the
stress peaks to around 80MPa (fillet where the rib joins a bolt hole at the upper middle of the
part). This stress is less than the material yield stress and is not a concern. The region of
relatively high stress may however be a point where a crack may propagate from after
prolonged repetitive use of the clutch and should be monitored should the engine be used for
an extended service life.
39. 39
4.4 Gearbox Cover
To enable assembly of the gearbox and final drive section of the engine, the design of the
engine required that the right hand side of the gearbox housing was a removable component
separate to the main engine block. This design allows the gearbox components to be
assembled within the main engine block and then held in place by the gearbox cover.
It was also decided that to reduce the unsprung mass (and overall mass) of the Formula SAE
vehicle and to take full advantage of the decision to use a locked differential, a single inboard
brake would be used. The best method of mounting the calliper for this braking system was to
incorporate hard points into the gearbox cover.
Because the gearbox cover incorporates many features into the single component, it is
perhaps the most complex components of the engine second to the main engine block. The
functions which it is required to perform are listed below.
1. Securely and accurately locate the primary gearbox shaft bearing (clutch shaft)
2. Securely and accurately locate the secondary gearbox shaft bearing
3. Securely and accurately locate the spool bearing and oil seal
4. Resist radial reaction forces generated from torque transfer through meshing gears
5. Locate the gear selector barrel bearing and oil seal
6. Locate the gear selector barrel detention roller and spring
7. Locate the gear selector fork slider/support
8. Locate the gearbox oil spray bar
9. Provide an oil channel to the primary gear shaft (clutch shaft)
10. Provide an oil tight seal against the main engine block
11. Be of a minimal mass
Overall layout of the various bearings and shafts located by the gearbox cover (in side view)
were determined during the design phase of the main engine block. To determine the axial
positioning of these components required accurate modelling of the relevant parts which are a
combination of components taken from the Honda CBR-600RR engine and in-house designed
components. These components include the primary and secondary gear shafts, gear selector
barrel, spool and all other associated bearings, gears, spacers, washers, circlip grooves etc. To
allow for a small margin of error and for adjustments of the axial location of the various gear
shafts to ensure correct meshing of the gears, allowance has been made to use spacers at both
40. 40
ends of the shafts which will be ground down to size during the assembly of the engine and
gearbox.
To prevent rotation of the bearing shells within their respective bores, two primary methods
have been used. For the primary gear shaft, spool and selector barrel bearings the respective
bores have been specified to be machined to press fit tolerances. For the secondary gear shaft,
a pin which protrudes from the bearing is located within a corresponding grove in the bore in
the gearbox cover, eliminating the chance of rotation. This method has been used in this case
due to the fact that the secondary gear shaft bearing is pressed onto its shaft.
An Achilles heel of many engines produced in small numbers is the difficulty of sealing oil
inside the engine around rotating shafts which pass through the engine casing, according to
John Coxon [26]. In order to minimise the possibility of oil leaks around both the spool bearing
and the selector barrel bearing, two lines of defence have been utilised. Firstly a “clip-in” nylon
seal has been used in both the spool and gear selector barrel deep groove ball bearings.
Secondly, a wiper seal specified by the seal and bearing supplier is used to completely seal oil
inside the engine.
To provide lubrication to the clutch and to the bearing which supports the oil pump drive
sprocket, oil was required to be fed through the clutch shaft to these areas. This was achieved
by having a connection to the internal pressurised oil system (copper tubing) to the gearbox
cover. A compact system of galleries was machined into the gearbox cover with an o-ring
sealed connection to the pressurised system, which delivers oil to the centre of the clutch
shaft bearing.
The gearbox cover is attached to the main engine block using M5 socket head cap screws
arranged around the perimeter of the cover. To ensure accurate location of the gear shaft
bearings in relation to the main engine block, 8mm dowel pins are arranged around the mating
surface.
As torque is transmitted through a pair of meshing spur gears, a radial reaction force is
generated [27], [28]. To provide support for the bearings and to resist any radial reaction
forces while adding minimal mass, thickness was added to the shells around the bearing bores
and ribs were included which radiate from the bearing bore shells to the bolting points around
the perimeter of the cover. An image of the design is shown below in Figure 17.
41. 41
Figure 17: Drawing highlighting the major features of the gearbox cover
In order to ensure that the design of the gearbox cover is sufficient, the potential magnitudes
of these forces were calculated. This calculation is shown below for the pinion of the final drive
gear pair which was found to be the greatest of the radial forces generated in the gearbox.
The torque generated at the wheels for 1g acceleration of a 300Kg car:
Where; m=vehicle mass (Kg), a=vehicle acceleration (m/s2
), d=wheel diameter (m)
The force tangential to the gear:
Where; Wt=tangential force (N), T=Torque (N.m), d=gear diameter (m), R=final drive gear ratio
The reaction force to this tangential force acts on the secondary gear shaft bearings normal to
the gear contact point.
A second force is generated at 90˚ to the tangential reaction force, known as the radial force.
This force is calculated below.
Where FR=radial force (N), Wt=tangential force (N), Ø=gear pressure angle
42. 42
The total reaction force is therefore the sum of these two forces:
This total force acts at 115˚ from the gear contact point (90˚+Ø).
The stresses resulting from this reaction force in the gearbox cover were analysed using finite
element analysis software, the results of which are shown below in Figure 18 and Figure 19.
Two separate analyses were carried out for both the pinion and gear of the final drive pair. The
force was applied as a bearing force with sinusoidal distribution and the mating flange around
the perimeter of the part was fixed. Note that the reaction forces generated by the other gears
on the secondary shaft oppose the calculated force and partially cancel its magnitude,
however in order to provide a margin of safety and to not over-complicate the process the
effect of the other gear pairs were ignored for this analysis. It can be seen from the results of
the analysis that the stresses are effectively dissipated through the part.
Figure 18: Gearbox cover FEA stress plot for the spool bearing loads
43. 43
Figure 19: Gearbox cover FEA stress plot for the final drive pinion loads
As it was decided that the 2013 Edith Cowan University Formula SAE vehicle would feature a
single inboard rear disk brake, the mounting of the calliper for this system was incorporated
into the gearbox cover. The rotor of this system is mounted directly to an end of the spool. To
provide mounting for this calliper two “ears” were added to the gearbox cover with M8
threaded mounting holes positioned so that the calliper was positioned correctly for a 240mm
diameter brake rotor.
Forces generated by braking can be some of the highest seen on a racing car [29], [30], [31]. In
order to ensure that the gearbox cover could withstand the stresses resulting from the braking
loads finite element analysis was carried out. The loads that can be potentially generated by
the rear brake are calculated below based on the assumption of a 300Kg car, 50% longitudinal
braking balance and 2.5g braking acceleration.
Where; T=braking torque (N.m), m=vehicle mass (Kg), a=braking acceleration (m/s2
), d=tyre
diameter (m)
The result of the FEA analysis is shown in Figure 20, where it can be seen that the stresses
generated are less than the material yield stress of around 250MPa. The majority of the
44. 44
stressed areas are at approximately 100MPa and little deflection was seen. The analysis was
carried out with the braking load applied to the calliper mounting “ears” as an 800N.m torque,
with the spool bearing surface represented as a fixed but sliding bearing surface (only rotation
around the bearing surface allowed), and a fixed surface at the opposite end of the gearbox
cover to balance the applied torque and prevent rotation around the bearing surface. There is
one small region where the stress in the analysis reaches 216MPa, however this can be
ignored because it is located at the edge of a surface which was fixed in the analysis for the
purpose of preventing rotation of the component. This application of loads and supports is not
realistic, but results in higher stresses than a real-life situation, providing a further margin of
safety.
Figure 20: Gearbox cover FEA stress plot for braking loads
The gearbox cover was manufactured externally by local company Robert Cameron & Co.
Manufacture was carried out on a 4-axis CNC vertical milling machine, and the part was
machined from 35mm thick 5083 Aluminium plate. Photographs of the finished gearbox cover
are included in Figure 21 and Figure 22 below.
45. 45
Figure 21: Photograph of the gearbox cover (outside)
Figure 22: Photograph of the gearbox cover (inside)
46. 46
4.5 Alternator Cover
The Edith Cowan University engine uses the standard Honda CBR alternator assembly. This
consists primarily of a rotor that is bolted to an end of the crankshaft, and a stator coil which is
mounted so that the concave rotor spins over it (see Figure 23). This design necessitates a
component which securely and accurately locates the stator. The alternator cover performs
this function and includes a sheet metal clamp which secures the alternator wiring and a port
for this wiring to pass through to the outside of the engine. The assembly of the ER-600-C1
alternator cover and associated components is shown below in Figure 23 (not including wiring
and wiring clamp).
Figure 23: Alternator assembly exploded view
Due to the complexity of calculating the loads generated by magnetic forces between the
alternator rotor and stator, not finite element analysis was carried out on the alternator cover.
To ensure that the alternator cover has sufficient strength and stiffness, comparisons were
made to the Honda CBR-600RR alternator cover during the design phase. The ER-600-C1
alternator cover was designed to have thicker wall thicknesses, a smaller mating surface with
the engine block, and larger stiffening ribs than the Honda equivalent.
Accurate and secure mounting of the stator coils within the rotor are of high importance with
this component therefore M6 shoulder screws have been used to precisely locate the
alternator cover in relation to the engine block. M5 socket head cap screws were also used to
mount the alternator cover. Two of the bolts in the cover are shared with a bracket which
47. 47
locates a bearing which is part of the gear shifting system. This is why the bolting flange is
raised for two of the bolts.
The alternator cover was manufactured externally by local company Robert Cameron & Co.
Manufacture was carried out on a 4-axis CNC vertical milling machine, and the part was
machined from 50mm thick 5083 Aluminium plate. Photographs of the finished alternator
cover are included in Figure 24 and Figure 25 below. A Photograph of the assembled alternator
cover with the stator coils is also provided in Figure 26.
Figure 24: Photograph of the alternator cover (inside)
48. 48
Figure 25: Photograph of the alternator cover (outside)
Figure 26: Photograph of the alternator cover with stator coil mounted
49. 49
4.6 Sump
Standard Honda CBR-600RR engines have a wet sump design where oil collects at the bottom
of the engine where it is picked up by the oil pump and re-circulated around the engine. This
design has two major drawbacks for a Formula vehicle. While a wet sump works well for a
motorcycle where the engine is not subjected to any significant lateral accelerations (as the
motorcycle corners, the rider leans into the corner and effectively cancels the lateral
accelerations), oil starvation can occur with this system in a formula car as the body of oil
moves around the sump as the car corners and lateral accelerations are generated. The wet
sump design also requires a relatively large volume beneath the crankshaft and results in a
raised vertical centre of mass of the engine.
In order to maintain a steady flow of oil to the engine with frequent, high lateral acceleration
cornering, and to keep the vertical centre of mass as low as possible, a dry sump oiling system
was decided to be used. A dry sump oil system involves pumping (scavenging) oil from the
sump and transferring it into a tank where air is separated, before being re-circulated around
the engine by a pressure pump.
The sump consists primarily of a flat plate with two wells with oil scavenge pick-ups toward the
front of the engine. Due to the direction of rotation of the crankshaft, oil inside the crank case
will be forced towards the front of the engine, and into the wells and oil scavenge pick-ups.
The space between the wells provides a passage for the oil scavenge lines to pass beneath the
engine to the Dailey Engineering oil pump at the rear. A drawing illustrating the main features
of the sump is provided below in Figure 27. Further information about the oil system can be
found in Chapter 7 of this report.
50. 50
Figure 27: Drawing of the sump showing the various features
The ER-600-C1 sump was manufactured externally by local company Robert Cameron & Co.
Manufacture was carried out on a 4-axis CNC vertical milling machine, and the part was
machined from 35mm thick 5083 Aluminium plate. Top and bottom view photographs of the
finished sump are included in Figure 28 and Figure 29 below.
Figure 28: Photograph of sump (inside)
51. 51
Figure 29: Photograph of sump (outside)
4.7 Camshaft Cover
Although the custom Edith Cowan University engine uses the Honda CBR-600RR cylinder head,
the standard camshaft cover is bulky and contains baffles and ventilation passageways which
are not required in this application. To allow the engine to be more neatly packaged within the
chassis of the 2013 Edith Cowan University Formula SAE car, a bespoke camshaft cover was
designed and manufactured.
The ER-600-C1 cam cover features:
Ports for the coil pack/spark plugs to pass through
Locating grooves for the standard Honda rubber gasket to be located
Four hard-points for bolting the cover to the cylinder head with allowance for
standard Honda rubber sealing washers
Individual internal ports for sealing the coil packs/spark plugs against oil with locating
grooves for the standard Honda rubber gasket
A vent to the oil overflow tank for pressure relief
52. 52
External mounting points for ancillary systems
The camshaft cover was manufactured in-house on the Okuma CNC vertical milling centre with
programming and operation carried out by students (see section 4.8). The part was
manufactured from 50mm thick 5083 Aluminium plate. Figure 30 and Figure 31 show
photographs of the finished camshaft cover.
Figure 30: Photograph of cam cover (outside)
Figure 31: Photograph of cam cover (inside)
53. 53
4.8 CNC Machining
Due to the fact that the engine team had some access to Edith Cowan University’s 4-axis
vertical CNC milling machine, and in order to save as much of the engine project budget for
purposes where it was most needed, as much of the CNC machining work was carried out in-
house as possible. The operation of the CNC milling machine was carried out by workshop staff
and by students.
In order to create code for the operation of the Okuma CNC milling machine, SolidCAM
software was invested in which integrates with Solidworks solid modelling software which is
predominantly used for 3D design of components by the ECU racing team. An integrated
CAD/CAM approach to manufacturing “...has emerged as one of the most effective tools for
improving the overall efficiency and productivity of manufacturing” [32]. The
Solidworks/SolidCAM integrated CAD/CAM system allows the designer/machinist to generate
machine G-code files in the same program window as the part was designed in, and with
machine tool paths automatically updated when the geometry of the part is modified. The
SolidCAM software automatically generates tool paths outlined by the user is capable of
simulation of tool paths before the code is sent to the machine, allowing the user to check for
mistakes and machining time and efficiency before any potential damage is done.
Although there were initially costs in terms of time and money in setting up the integrated
CAD/CAM system and learning to use the software and CNC machinery, the savings to the
team’s budget during this process outweighs these costs. During this process of moving
manufacture of complex CNC parts in-house, the team has gained experience and knowledge
which will result in significant future cost savings.
Parts programmed and CNC machined in-house by engine team students were the camshaft
cover, crankshaft main bearing caps, and gear selector barrel detent wheel.
54. 54
Chapter 5 – Transmission
5.1 Simulation of Selection of Gears and Final Drive Ratio
The final drive gear pair was one of the first components designed and manufactured for the
project. Because many aspects of the bespoke engine concept rely on the final drive gear
reduction, it was important to acquire these components early in the project.
It was decided early in the project that only two gear ratios would be used in the engine from
the Honda CBR-600RR gearbox to minimise the rotational inertia of the gears and to minimise
the complexity of the gear shifting mechanism. Through analysis of vehicle data from use of
the Honda CBR-600RR in Formula-SAE competition, the Edith Cowan University Formula-SAE
Team has found that only the first three of the six gear ratios are used. The decision was made
to use the 1st
and 3rd
gear ratios in the bespoke engine due to the ability to shift between the
two ratios with only one selector fork, the ability to use unmodified gears from the Honda
gearbox, and the relatively large difference between the two ratios. The relatively large
difference between the 1st
and 3rd
gear ratios allow for the maximum speed range that the
engine can power the vehicle through in the range of engine speeds which produce optimum
torque.
Design of the final drive began with determining the appropriate gear ratio to suit the
Formula-SAE vehicle in the conditions it is expected to encounter in competition. Vehicle data
was reviewed from previous competitions at the Australasian Formula-SAE venue along with
studies of various international venues to determine the top speed, average speed, and
minimum corner speeds expected to be encountered in Formula-SAE competition. Factors
such as the torque curve of the engine, fuel usage, and tyre diameter were also taken into
account when determining the final drive ratio. The graph in Figure 32 below was used as a
tool as part of a spreadsheet to determine the appropriate final drive ratio.
55. 55
Figure 32: Graph of car speed against engine speed for final drive ratio design
The parameters used to select the final drive reduction of around 2.7 were:
Maximum engine speed of 12000rpm
Tyre diameter of 0.5m
Vehicle top speed of at least 120kph
Engine speed greater than 6000rpm at 40kph for low speed corner exit performance
Engine speed between 6000rpm and 8000rpm at an average speed of 65kph for fuel
economy
5.2 Final Drive Gear – Design, Manufacture
Once the desired gear ratio had been selected, it was decided that spur gears were to be used
due to the lack of axial thrust generated by spur gears in comparison to helical gears, because
there was no requirement for quiet running gears, for ease and lower cost of manufacture,
and because spur gears were already in use in the other gear ratios in the gearbox. Initial
calculations were carried out to determine the appropriate size and number of teeth for the
gears. These initial calculations were carried out using a process recommended by Shigley’s
Mechanical Engineering Design [28]. It was calculated initially that the gears should have 16
teeth on the pinion and 43 teeth on the gear to give a ratio of 2.6875, and a module of around
3.5 and a face width of around 30mm. The gears have an even number of teeth on the pinion
and an odd number of teeth on the gear so that there is a “hunting tooth”. A “hunting tooth”
results in even wear of the gears due to there being no tooth on the pinion which repeatedly
contacts a particular tooth on the gear, ensuring that small manufacturing imperfections are
not magnified over long running periods.
0
2000
4000
6000
8000
10000
12000
14000
0 25 50 75 100 125
EngineRPM
Speed, kph
Car Speed vs Engine RPM
1st Gear
3rd Gear
56. 56
Figure 33: Illustration of various gear dimensions [33]
The module of a gear is determined using the following equation.
Where m=module, D=pitch diameter, and z=number of teeth.
To verify the calculations recommended by Shigley’s Mechanical Engineering Design [28],
American Gear Manufacturing Association (AGMA) standards were researched. Alternative
methods of calculating bending stresses and contact stresses in gear teeth were investigated in
the standards; AGMA 908-B89 - Geometry Factors for Determining the Pitting Resistance and
Bending Strength of Spur, Helical and Herringbone Gear Teeth [33], and ANSI/AGMA 2001-D04
– Fundamental Rating Factors and Calculation Methods for Involute Spur and Helical Gear
Teeth [34]. Calculations were performed based on these standards, but there was still
uncertainty in the results due to the number of factors which needed to be estimated due to
lack of reliable information, such as machining tolerances and heat treatment processes.
The calculation method finally used to verify the appropriate size of the gears for the
application was found in ANSI/AGMA 6002-B93 – Design Guide for Vehicle Spur and Helical
Gears [27]. This method involves plotting different loading conditions onto an S-N histogram
for both bending stresses and contact stresses. Due to lack of material fatigue properties of
57. 57
the material used to make the gears, S-N curves were plotted for a range of materials
described in ANSI/AGMA 6002-B93 [27]. The gears from the Honda CBR-600RR gearbox were
also reverse engineered and analysed in the same method to ensure that the calculation
results were realistic (using the assumption that the Honda gears would be of similar material
properties to the final drive gears). The design of the final drive gears were adjusted to yield
similar numbers of stress cycles as the Honda gears and a final design of a 3mm module, 16
tooth pinion, 43 tooth gear, with a 25mm face width and 25˚ pressure angle was arrived at.
The resulting contact and bending stress histograms for the final drive design are included
below in Figure 34, Figure 35, Figure 36 and Figure 37.
Figure 34: Bending stress S-N histogram for the final drive gear
0
25000
50000
75000
100000
125000
150000
175000
200000
1.E+03 1.E+04 1.E+05 1.E+06 1.E+07
BendingStress(lb/in^2)
Stress Cycles
Gear Bending Stress Histogram
worst case
Clutch grab
wheel slip
full load low gear
full load high
GRADE 1, L1
GRADE 1, L10
GRADE 2, L1
GRADE 2, L10
58. 58
Figure 35: Bending stress S-N histogram for the final drive pinion
Figure 36: Contact stress S-N histogram for the final drive gear
0
25000
50000
75000
100000
125000
150000
175000
200000
1.E+03 1.E+04 1.E+05 1.E+06 1.E+07
BendingStress(lb/in^2)
Stress Cycles
Pinion Bending Stress Histogram
worst case
Clutch grab
wheel slip
full load low
full load high
GRADE 1, L1
GRADE 1, L10
GRADE 2, L1
GRADE 2, L10
0
100000
200000
300000
400000
500000
600000
700000
1.E+03 1.E+04 1.E+05 1.E+06 1.E+07
ContactStress(lb/in^2)
Stress Cycles
Gear Contact Stress Histogram
worst case
Clutch grab
wheel slip
full load low
full load high
GRADE 1, L1
GRADE 1, L10
GRADE 2, L1
GRADE 2, L10
59. 59
Figure 37: Contact stress S-N histogram for the final drive pinion
The histograms in Figure 34, Figure 35, Figure 36 and Figure 37 indicate that the weakest gear
of the final drive pair is the pinion, and would be most likely to fail in bending. To further
minimise the likelihood of this type of failure occurring, the gears are made from the best gear
material which was readily available, EN36A – Case Hardening Steel, heat treated to achieve
case depth of 0.8-1.0mm and quenched and tempered to a hardness of 58-60 HRC. To help
bending stresses to be distributed through the pinion teeth and to ensure that the full width of
the gear pair has full contact across the face width of the gears, the pinion is 2mm wider than
the gear. The pinion also has a tip relief applied to reduce bending stresses.
The final drive gears were also designed to be effectively integrated with the bespoke engine
and the parts taken from the Honda CBR-600RR. The spline of the CBR-600RR secondary gear
shaft was reverse engineered and cut into the pinion of the final drive gear pair so that it could
be located on the aforementioned shaft. The gear of the final drive gear pair was designed
with an internal 18-hole PCD and a precision ground internal bore for accurate location in a yet
to be designed differential or spool assembly. The 18-hole PCD allows for an evenly spaced 3,
6, 9, or 18 point circular bolting pattern.
Manufacture and heat treatments of the final drive gear pair were carried out by CAMCO
Engineering using a hobbing process with final grinding of the gears performed after heat
treatment. CAMCO Engineering also provided design advice, and supplied the material for the
0
100000
200000
300000
400000
500000
600000
700000
1.E+03 1.E+04 1.E+05 1.E+06 1.E+07
ContactStress(lb/in^2)
Stress Cycles
Pinion Contact Stress Histogram
worst case
Clutch grab
wheel slip
full load low
full load high
GRADE 1, L1
GRADE 1, L10
GRADE 2, L1
GRADE 2, L10
60. 60
gears. Technical drawings which were supplied to the gear manufacturers and a photograph of
the finished gears are included below in Figure 38, Figure 39 and Figure 40.
Figure 38: Technical drawing of the final drive pinion
61. 61
Figure 39: Technical drawing of the final drive gear
Figure 40: Photograph of the final drive gear pair
62. 62
5.3 Gearbox Assembly
Figure 41: Gearbox assembly render
The two speed constant mesh ER-600-C1 gearbox is made up of a combination of standard
2006 Honda CBR-600-RR components and bespoke in-house designed parts (see Figure 41).
The decision was made to use as many Honda parts as possible in an unmodified state in order
to minimise the number of parts required to be designed and manufactured, to carry over the
reliability of the Honda engine/gearbox to the ER-600-C1, to integrate effectively with the
Honda crankshaft and clutch which were already decided to be used, and to be able to source
spare parts easily.
Through investigation of the Honda gearbox and by carrying out simulations of the final drive
gear ratio (see section 5.1) it was decided to use the first and third gear pairs from the Honda
gearbox. Aside from the ratios of the gears themselves, there were a number of other reasons
that the first and third gear pairs were chosen.
The pinion of first gear is manufactured as a single part with the clutch shaft.
63. 63
Both first and third gears are freely rotating on the secondary shaft and a standard
Honda splined dog-toothed sliding selector is able to transmit power from either gear.
Only a single sliding dog-toothed selector is required, simplifying design of the gear
selector mechanism.
The first and third gear pairs could be located in standard locations on the standard
Honda gear-shafts which are on one side of the gearbox. This packaging allows for
sufficient room in the gearbox casing to fit the final drive gear pair.
The first and third gear pairs from the Honda gearbox are capable of transmitting the
most torque compared with higher gear pairs.
Other parts taken from the Honda gearbox to simplify the design process of the ER-600-C1
gearbox include:
Both primary (clutch) and secondary gear shafts
All four primary and secondary gear shaft bearings
The gear selector fork
Freely rotating first and third gear bearings
Spacers, washers, circlips
Having taken as many parts as possible from the Honda CBR-600-RR gearbox, there remained
some parts which needed to be designed and manufactured for the ER-600-C1. These parts
mainly consisted of the selector barrel and mechanism, the final drive gear pairs, and the
spool. More details of these parts can be found in sections 5.4 (selector barrel), 5.2 (final drive
gears) and 5.5 (spool).
To illustrate the parts of the gearbox assembly and how they interact, an exploded view of the
ER-600-C1 gearbox is included below in Figure 42. A rendered image of the gearbox assembly
is also shown in Figure 41.
65. 65
5.4 Selector Barrel
In order to select first gear, second gear or a neutral position, a mechanism is required to slide
the dog-toothed selector gear from side to side to engage the dogs with the freely rotating
first or second gear gears. To achieve this, a slotted selector barrel was designed to translate a
rotation of the barrel to a linear motion of the selector fork, which in turn slides the dog-
toothed selector gear. The slot in the selector barrel has three distinct positions; one for first
gear, one for second gear, and one in-between for a neutral position. The selector barrel is
shown in its corresponding assembly in Figure 43.
Figure 43: Gear selector barrel assembly
The design of the selector barrel was loosely based on the selector barrel from the Honda
gearbox. The Honda CBR selector barrel is relatively complex and functions in a slightly
different way than the ER-600-C1 because the Honda has six gears and a neutral position to
select between with three selector forks, while the ER-600-C1 only has two gears and a neutral
position with one selector fork. Due to the number of positions that the slots in the Honda
selector barrel require the barrel needs to use almost 360˚ of rotation and has a relatively
large diameter. To make it possible for the Honda selector barrel to be rotated almost 360˚ by
the action of a lever, it incorporates a ratchet mechanism which allows the barrel to be rotated
precisely one position per application of the gear shift lever.
To simplify the design and minimise the number of parts, and considering the fact that the
gearbox only has three positions to select between, it was decided to rotate the selector barrel
66. 66
by a direct lever action rather than a ratchet mechanism. This decision limited the angular
rotation of the selector barrel to around 90˚ to maintain reasonable torque applied to the
barrel through a gear selector lever.
Because of space constraints in the ER-600-C1 gearbox housing the diameter of the selector
barrel was made smaller than the Honda. This made packaging of the slot in the barrel,
including the three positions and the ramp between the positions, challenging with only 90˚ of
rotation. This packaging was able to be achieved by:
Making the distance between the gear positions as small as possible (7mm)
Sequencing the gears as 1st
– N – 2nd
rather than N – 1st
– 2nd
as was originally intended
Making the positions in the slots as short as possible
The compromises made in the design of the selector barrel mean that there is little room for
error in the installation of the barrel, and incorrect rear engagement, or jumping out of gear
could occur if not properly setup.
To ensure that the required torque to move the selector fork and dog-toothed selector gear
from side to side and the stresses on components would be comparable to the Honda gearbox,
the angle of the slot ramps between the positions on the selector barrel were reverse
engineered from the Honda selector barrel and adapted to the new diameter. In order to do
this, the following steps were taken.
The axial distance between positions was measured on the Honda selector barrel and
transferred to the ER-600-C1 barrel
The angular rotation between the two positions was measured
The angular distance between the two positions was multiplied by the ratio of the
smaller to the larger diameter
For example:
If the larger Honda barrel diameter = 42mm, and the ER-600-C1 barrel diameter = 32mm, and
the angle between the positions on the Honda barrel = 15˚, then the angle between the two
positions on the ER-600-C1 barrel = 15˚ x (42 / 32) = 19.7˚
The selector barrel is supported at both ends by 17-ID - 30-OD deep-groove ball bearings. The
barrel was manufactured from EN-26 steel alloy for its durability and relatively good corrosion
resistance. The barrel was made in two parts so that it could be made hollow in order to
67. 67
minimise its mass without the possibility of filling up with engine oil. Manufacturing was
carried out by local company, High Speed Engineering. A photograph of the finished part is
included in Figure 44.
Figure 44: Gear selector barrel
5.5 Final Drive –Spool, Tripods, Drive-shafts,
To transmit torque from the final drive gears to the wheels it was decided to use a locked
differential (spool) rather than a limited slip differential. From experience in 2012 Formula SAE
competition it was found that a spool gave no real disadvantage in performance or handling,
but provided a significant reduction in rotating mass, savings in time for design and
manufacture, and financial savings. Use of a spool also permits the change to a single rear
inboard brake from duel outboard rear brakes, resulting in reduced unsprung mass and
reduced rotating mass and inertia.
The design of the spool began with the requirements and constraints listed below.
Transmit torque from the engine to the drive shafts
Integrate with the final drive gear
68. 68
Have built-in C.V. joint housings to suit Taylor Race Engineering tripod C.V. joints
Locate C.V. joints axially with sufficient room to travel and seal in grease
Provide a mount for the rear brake rotor
Rotate on 70-ID – 90-OD deep groove ball bearings
Incorporate a seal to prevent oil from the engine leaking through the bearings
The spool was designed primarily as a two-piece unit which sandwiches the final drive ring
gear, and has built-in C.V. tripod housings at each end. In addition to the primary structure of
the spool, there are various other components which are part of the spool assembly which
enable the spool to function effectively. These additional parts are described below and an
exploded view of the spool assembly can be found in Figure 45.
Located inside the spool between the C.V. housings is an ABS rapid prototype internal
brace which prevents the C.V. joints from over travelling inside the spool.
At either end of the internal brace are pressed aluminium domes which prevent
wearing of the ABS brace, allow the C.V. joints/drive shafts to move
up/down/forwards/backwards smoothly as the wheel moves through its arc of travel
and the spool rotates, and seal grease inside the C.V. housings.
C.V. boot cups are bolted to the outside faces of the C.V. joint housings and prevent
the C.V. joints from over travelling out of the housings, seal grease inside the spool,
and provide a mount for silicone C.V. boots.
Small spaces at either end of the spool allow adjustment to the alignment of the final
drive gears
The brake disk carrier bolts to one end of the spool
A drive speed sensor trigger wheel attaches to one end of the spool, providing data
through a Hall effect sensor to the ECU for traction control operation and data for
analysis of the driver/track/vehicle
The main spool assembly is held together with nine 5/16”-18 UNC bolts with locking k-
nuts.
69. 69
Figure 45: Spool assembly exploded view
To ensure that the strength of the main spool components was sufficient to withstand the
loads generated through driving, FEA analysis was carried out on both the left and right sides
of the spool assembly. In order to perform this analysis, a maximum load expected to be
generated due to driving the wheels was calculated based on an acceleration of 1.3G, vehicle
mass of 300Kg, and a tyre diameter of 0.44m. The calculation performed is shown below.
Where; T=Torque (N.m)
m=car mass (Kg)
a=car acceleration (m/s2
)
d=Tyre diameter (m)
For the FEA analysis, this load was applied to the final drive gear flange while the tripod
housing faces were made fixed and vice versa, for both ends of the spool individually. The
results shown in Figure 46 and Figure 47 are from where the load is applied to the gear flange
and where the C.V. joint contact faces are fixed for both parts of the spool. The results from
this configuration of loads and constraints showed the highest stresses.
70. 70
Figure 46: Spool FEA stress plot (side A)
Figure 47: Spool FEA stress plot (side B)
The peak stresses for both of the spool ends occur immediately adjacent to the faces which are
fixed. Fixed faces in FEA analysis result in stress concentrations next to these fixed faces,
therefore these peak stresses in the results of the analyses can be largely ignored. The
remainder of the stresses in the spool are relatively evenly distributed and are far less than the
material yield stress. The results from the FEA analysis indicate that the spool is capable of
enduring the loads it is expected to be subjected to.
