Racing Electric Vehicle – REV
Proposal
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1: Introduction .............................................................................................. 1
1.1: Purpose .............................................................................................. 1
1.2: Team Goals ........................................................................................ 2
1.3: Background ........................................................................................ 2
2: Design Objectives .................................................................................... 3
3: Design ...................................................................................................... 4
3.1: Chassis and Body .............................................................................. 4
3.1.1: Frame ........................................................................................... 4
3.1.1.1: Engineering Specifications ..................................................... 4
3.1.1.2: Design History ........................................................................ 5
3.1.1.3: Engineering Analysis.............................................................. 8
3.1.1.4: Material Study ........................................................................ 12
3.1.2: Body ............................................................................................. 13
3.1.2.1: Engineering Specifications ..................................................... 13
3.1.2.2: Design History ........................................................................ 13
3.1.2.3 Material Study ......................................................................... 15
3.2: Vehicle Dynamics ............................................................................... 15
3.2.1: Suspension and Steering Geometry ............................................. 15
3.2.1.1: Introduction ............................................................................ 15
3.2.1.2: Track Width ............................................................................ 15
3.2.1.3: Geometry ............................................................................... 15
3.2.1.4: Material Study ........................................................................ 22
3.2.2: Braking ......................................................................................... 23
3.2.2.1: Engineering Specifications ..................................................... 23
3.2.2.2: Design History ........................................................................ 24
3.2.2.3: Engineering Analysis.............................................................. 24
3.2.2.4: Material Study ........................................................................ 24
3.2.3: Wheels, Tires, and Uprights ......................................................... 25
3.2.3.1: Engineering Specifications ..................................................... 25
3.2.3.2: Design History ........................................................................ 26
3.2.3.3: Engineering Analysis.............................................................. 27
3.2.3.4: Material Study ........................................................................ 27
3.3: Drive System ...................................................................................... 28
3.3.1: Motor and Power Train ................................................................. 28
3.3.1.1: Engineering Specifications ..................................................... 28
3.3.1.2: Design History ........................................................................ 29
3.3.1.3: Engineering Analysis.............................................................. 30
3.3.1.4: Material Study ........................................................................ 35
3.3.2: Power Source ............................................................................... 37
3.3.2.1: Engineering Specifications ..................................................... 37
3.3.2.2: Design History ........................................................................ 38
3.3.2.3: Engineering Analysis.............................................................. 39
3.3.2.4: Material Study ........................................................................ 40
3.3.3: Cooling ......................................................................................... 41
3.3.3.1: Conceptual Design ................................................................. 41
3.3.3.2: Material Study ........................................................................ 41
3.4: Electrical System ................................................................................ 42
3.4.1: Electrical Specifications and Interface Requirements ................... 42
3.4.2: Technical Hardware Design ......................................................... 44
3.4.3: Technical Software Design ........................................................... 49
3.4.4: Schematic..................................................................................... 52
3.4.5: Reliability and Maintainability Assessments ................................. 53
3.4.6: Test Plan ...................................................................................... 53
3.5: Driver Interface/Ergonomics ............................................................... 55
3.5.1: Acceleration Pedal ....................................................................... 55
3.5.1.1: Engineering Specifications ..................................................... 55
3.5.1.2: Design History ........................................................................ 55
3.5.1.3: Engineering Analysis.............................................................. 55
3.5.1.4: Material Study ........................................................................ 58
3.5.2: Steering Wheel ............................................................................. 58
3.5.2.1: Engineering Specifications ..................................................... 58
3.5.2.2: Design History ........................................................................ 59
3.5.2.3: Material Study ........................................................................ 59
3.5.3: Driver’s Seat ................................................................................. 60
3.5.3.1: Engineering Specifications ..................................................... 60
3.5.3.2: Material Study ........................................................................ 60
3.5.4: Safety Equipment ......................................................................... 60
3.5.4.1: Engineering Specifications ..................................................... 60
3.5.4.2: Material Study ........................................................................ 61
4: Budget ...................................................................................................... 62
4.1: Initial Budget....................................................................................... 62
4.2: Final Budget ....................................................................................... 63
5: Organization and Capabilities .................................................................. 64
6: Scheduling ............................................................................................... 66
6.1: Gantt Chart ......................................................................................... 66
6.1.1: Mechanical Task Schedule........................................................... 66
6.1.2: Electrical Task Schedule .............................................................. 67
6.2: Milestones and Deadlines .................................................................. 68
7: Appendix .................................................................................................. 69
7.1: Calculations ........................................................................................ 69
7.2: References ......................................................................................... 74
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REV, the Racing Electric Vehicle, comes out of a tradition of Motor Sports at Florida tech.
For years now Florida Tech has been competing in the Formula SAE and Mini Baja
competitions. This year the team hopes to begin a new tradition at Florida Tech as we
introduce Florida Tech to the growing field of electric racing.
The REV Team has decided on building an electrically driven, open-wheel, single-seat,
purpose-built vehicle optimized for Autocross racing. In Autocross, the drivers race through
a flat road course that is often setup in a large parking lot. It takes only a few minutes to
race through the tight, winding course in which the car must accelerate, decelerate, and
corner very quickly. To promote the idea of to keep the weight and cost down we will
design the battery setup and gearing for the short, high acceleration races, where the
speeds are usually between 20 and 40 mph. To decrease the cost and time of
development we plan to scavenge a few components from the 2001-2002 Florida Tech
Formula SAE car. This should help decrease the overall cost for the project and reduce
the design and fabrication time.
The REV team is composed of 13 members from various disciplines and is actively
working to overcome the design, management, and communication challenges of the
project.
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The ability to be powered by electricity generated from all types of alternative energy
sources has drawn much attention towards electric vehicles. The significant efficiency
advantage that electric motors have over internal combustion engines has determined their
place in the future of automotive engineering. With the pervasion of electric motor systems
in all design applications, an electric drive race car is exceedingly relevant.
The Racing Electric Vehicle (REV) project is a remarkable opportunity for students to
become a management, design, and production team. Every student is learning in a whole
new way as they must apply all their knowledge to this demanding practical challenge.
Together the students will learn to manage themselves and communicate in ways much
closer to the industry than any other experience during college. Invaluable experience and
knowledge will be gained by every student through this challenge, and with it one more
piece to the developing array of electric powered vehicle knowledge.
This project will also serve to highlight electric drive technologies on and off the Florida
Tech campus in a visible, personally dramatic way. The team looks to draw the public and
the campus community into the excitement of the project and the potential of electric
power systems for the future. In that they seek to further school spirit, Florida Tech’s
relations with the community, and public interest in electric vehicle technology.
Historical EV
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The goals of the REV project are:
To design and build an electric vehicle for Autocross style racing that would
be capable of being competitive in the Formula SAE Hybrid races.
To diminish the challenges commonly associated with electric vehicles
including Power to Weight Ratio and total cost.
To build effective management, communication, and teamwork skills amongst
the student team, mentors, sponsors, and the community.
To allow time for thorough testing and optimization of the completed vehicle
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In the 1830’s the first electric vehicle was invented by Robert Anderson in Scotland. This
was a crude vehicle that was basically an electric carriage.
Electric vehicles, or EVs,
began to gain notice in America in the 1890’s and after this interest
increased in vehicles such as the one in Figure 1 below. This is
the is the 1902 Wood's Phaeton it is a typical electric vehicle of the
time, and had a cost of $2,000, a max speed of 14 miles per hour,
and a range of 18 miles [1].
These vehicles were widely used
because they lacked the noise and hand cranks of the gasoline
vehicles, and most people only went around town so the small range was ideal for them.
As gasoline vehicles made advances a decline began in electric vehicles. Electric vehicles
made a return in the 1960’s and 1970’s when there was a push for environmentally
friendly automobiles. Between this time and 1990 there were several companies that
accomplished making vehicles that fit certain needs and most had ranges from 50-60
miles at around 40 miles per hour. These vehicles ranged from service trucks to city cars,
and while these vehicles may not be
widely known they
laid the ground work for the more modern
electric
During the 1990’s there was another
push
vehicles because of new regulations in
pollution
Figure 2. EV Pick-Up Truck
vehicles.
for
electric
control.
The US Department of Energy and several other companies began creating new vehicles
from the ground up and converting existing vehicles to run on electricity. These new and
converted vehicles, like in figure 2, were trucks, vans, and even sports cars that could run
at highway speeds with larger ranges than any of the previous electric vehicles. The
downside to these vehicles was that they cost up to $40,000, but improvements in
production are making these vehicles have prices on the same lines as gasoline vehicles.
Electrical vehicles are used in various applications: working in industrial plants, on golf
courses, and on college campuses. Today these quiet, pollution-free vehicles are no
longer overgrown golf carts. There has been a recent emergence of high performance
open-wheeled electric race vehicles. These vehicles compete at well known race venues
across the USA and demonstrate that electric vehicles are no longer slow lumbering
vehicles. In the past there have been specific races catering to electric race cars. One of
the more widely spread races among colleges was the Formula Lightning. Numerous
universities across the country have joined in serious designs to compete every year. In
this race, most cars ran between 350 and 400 V at about 240 amps.
Also, other organizations, such as the National Electric Drag Racing Association (NEDRA),
are interested in promoting the sport of EV drag racing. For the vehicle design at hand the
best competition is the Sports Car Club of America (SCCA) Solo Autocross competition. It
is made up of short (under 5 minute) races involving obstacles and straight track runs.
Drivers race through a flat road course that is often setup in a large parking lot. It takes
only a few minutes to race through the tight, winding course in which the car must
accelerate, decelerate, and corner very quickly. The electric vehicle design would race in a
modified category and would run against all types of vehicles.
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For our electric vehicle, we have decided on specific design objectives we intend to reach.
These objectives include:
Acceleration from 0 to 60 mph in under 5 seconds
Top speed of 85 mph
Maximum power available between 20 and 40 mph.
Lightweight (under 650lb)
15 minute battery life running at high performance speeds
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We have decided on building an electrically driven, open wheel, single seat, purpose built,
vehicle optimized for Autocross racing. To keep the weight and cost down, we designed
the battery setup and gearing for the short, high acceleration races, where the speeds are
usually between 20 and 40 mph. To save the cost and time of development, we
scavenged some of the components from the 2001-2002 Florida Tech Formula SAE Car.
This decreases the overall cost and reduces the design and fabrication time.
An important aspect to designing a racing vehicle is balancing the power and weight. The
basic point to this idea is a light car will have more power than a heavier car. The heavier
the car, the more torque will be applied to the motor, thus making the car slower.
Therefore, keeping the weight of the car as low as possible is an important design factor
that we considered. Many aspects of a high performance vehicle consider high velocity and
the vehicle’s integrity at those velocities.
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3.1.1: Frame
3.1.1.1: Engineering Specifications
Table 3.1: Engineering Specifications
Label
Value
Length Minimum (front-to-rear wheels)
60 inch
Main Hoop Angle
10 º from vertical
Driver Head Clearance
2.00 inch
Vehicle Track (S = Small Track, L = Large Track)
S>.75L
Deflection
.333 inch
Torsional Rigidity
>1600 ft-lbs/deg
Maximum Stress from Static Loading
21,300 psi
Maximum Stress on Weld Points
26,600 psi
Maximum Stress from Dynamic Loading
21,300 psi
The specifications for the chassis are based on ideal characteristics for space frame
design. A list of optimal values for this frame is provided in Table 3.1 above. The Formula
SAE rules [15] regulate length, main hoop, driver clearance, and vehicle track. Deflection
and maximum stresses are based on the material used in the frame. As stated in Material
Study, 3.1.1.4, AISI 4130, Chromoly, will be used. According to the American Society for
Nondestructive Testing [5], a factor of safety for vehicles in this size is 3. This factor of
safety will be applied to all the stresses acting on the vehicle. Stress on the weld points is
important factor because these areas may be a weaker point on the frame. We need to
determine a maximum stress these welds will be able to handle to be sure our frame does
not buckle at these weaker points. Torsional rigidity is based on data from other similar
frame designs, see references [19]. If the frame is too flexible (i.e. below 1600 ftlbs/deg), the vehicle will not handle well.
3.1.1.2: Design History
The chassis went through a set of iterations of conceptual design then another set of
iterations of optimization. For conceptual design, the idea was using basic geometry
patterns to construct the support on the chassis. Triangulation patterns show the most
strength in geometric patterns, so many of the development of our chassis is more
triangulated in the final conceptual chassis design. We first attempted a basic structure,
shown in Figure 3. This concept represented what size we had for the vehicle, but would
not adequately support the different weights.
Figure 3. First Conceptual Chassis Design
Next conceptual revisions we added more triangulation for support and more idealistic
structure to support the various components. We took into consideration a two motor setup
in this chassis, as shown in Figure 4a. The rear section of the frame shown in Figure 4b
is a differential mounting box. This design was changed due to manufacturability.
(a)
(b)
Figure 4. (a) Chassis Two Motor Setup, (b) Chassis with Differential Mounting Box
The next revision was developed with changes to the rear end. The chassis has more
triangulation on the bottom surface and the rear differential support is now constructed with
square tubing, as shown in Figure 5. This will be lighter than the metal box and will
adequately support the differential.
(a)
(b)
Figure 5. (a) Side view, (b) Top view
The first concept of the side pod was a simple rendition of
a general layout of columns of batteries and side pod
layout. The general idea was to have the side pod open to
the outside to allow air to pass through the pod over
staggered columns of batteries. This general design was
enhanced in later concepts.
Front
Open
to air
flow
Figure 6. Top view of side pod (First concept)
The second side pod conceptual design. The changes were to allow the side pod to have
a side profile similar to that of the existing frame. This would ease the construction and
analysis of the side pods. The front view had not been decided on until the next design
concept.
Figure 7: Top view of side pod
The final conceptual design revision of the chassis has some small changes from the
previous layout. The variation of placement of the square tubing, bending of roll hoops for
ease of manufacturability, and measurement changes to support the components are some
changes made to this chassis. The side pod design was also incorporated into this
revision. Also, the chassis was evaluated to incorporate the rules set by Formula SAE
[15]. This latest version can be seen in Figure 8.
(a)
(c)
(b)
Figure 8. (a) Isometric view, (b) Side view, (c) Front view
Next, the final conceptual frame design was optimized for handling and support. Cross
members were added and variable wall thicknesses were changed for better support or
lightening the frame. Table 3.2 lists the optimization iterations the frame underwent from
finite element analysis.
Table 3.2: Optimization Iterations
Deflection
Stresses
(inch)
(psi)
1
.0068
8493
No change, conceptual chassis
2
.0054
4341
Add cross member in chassis, see Figure 9
3
.0070
5706
Some member wall thicknesses changed to .049”
4
.0020
8934
Driver side support changed angle, see Figure 10
5
.0055
4348
Suspension perches wall thickness changed to .095”
6
.0054
4342
Suspension perches wall thickness changed to .049”
Iteration
Cause of Change
For the side pod, the cross member was added to decrease the overall deflection. This
added support also stiffened the frame, increasing the torsional rigidity.
Figure 9. Side Pod cross members, iteration 2
To optimize deflections and stresses, the driver side support angle was changed. The
increase in stresses was too great to justify the decrease in deflection.
(a)
(b)
Figure 10. (a) Driver side support, iteration 4, (b) Final design choice
3.1.1.3: Engineering Analysis
All of the design modifications were taken into consideration in the analysis performed on
the chassis. The chassis was evaluated using finite element analysis packages ANSYS
and ADAMS.
For static loading, the frame was generated in ANSYS and all heavy weight components
were applied to the frame. The frame is constrained at the attachment points of the
suspension to accurately represent the actual model. As shown in Figure 11, the frame
holds and maximum load is placed at the back of the frame near the motor.
Figure 11. ANSYS Static Loading
Figure 12. ADAMS Model of Vehicle
Figure 12 shows the Adams model that was used for analysis. In ADAMS we used a
breaking, cornering and lane change analysis. In doing so the forces at each A-arm
attachment point were acquired. These forces, seen in figures 13 and 14, were then
applied to the ANSYS model at the attachment point locations to see the stress that
occurred in the frame. For breaking the analysis was done for a 50mph to 0 maneuver in
6 seconds. The lane change was done at 50mph and the cornering analysis was done for
a radius of 328 feet at a constant speed of 60mph. The analysis is considered
conservative because in each case the max loads were applied to each point, even if the
max loads did not occur at the same time point. Therefore, a larger load is applied to the
frame. Table 3.3 shows the results of each case. The magnitude forces were applied
along the same angle of each A-arm.
Figure 13. Loads applied on front A-arm geometry
Figure 14. Loads applied on rear A-arm geometry
The resultant stresses of cornering and braking yield higher values than that of stresses
seen from a lane change. Since the frame can handle the loads of cornering and braking,
therefore frame will be able to handle loads taken from a lane change.
point:
1
2
3
4
5
6
7
8
9
Table 3.3: Forces of Dynamic Loads
Cornering
Braking
Magnitude(lbf)
Magnitude(lbf)
39.31908413
-220.762382
34.84538617
272.4684389
-20.45761381
-64.2953577
66.09382924
245.9409836
-70.81481705
216.7158211
89.9235772
-275.3909552
-100.7144065
-62.7216951
184.3433333
-243.2432763
-41.58965446
-135.7846016
Lane change
Magnitude(lbf)
44.06531124
47.82222352
-38.73700112
81.59505554
-102.8143983
-60.85219105
52.19380568
152.5279732
-67.56172658
10
11
12
13
14
15
16
stress(psi)
-38.89194714
-37.99271137
-10.56602032
-51.03163006
-59.5743699
56.42704469
66.99306501
15120
98.01669915
-87.00106094
80.48160159
137.8078821
-98.69112598
85.65220728
-80.93121948
15506
83.10058853
22.21823304
19.74244951
-19.39881046
46.417467
-31.84996122
25.4444498
NA
The max Von mises stress that we experienced was 15506 psi, which is below our
engineering specification of 21300 psi.
To prove that the ADAMS analysis is correct, hand calculations were done for the lateral
acceleration, and lateral tire forces and then compared to the ADAMS output.
v 2 26.284 2 m / s

 7.99m / s 2
r
100m
F  ma  226.79lbs * 7.99m / s 2  1812.0521N
a
a= angular acceleration
v= velocity
r= radius of turn
F= force
M= mass of car
Adams output:
a= 7.19m / s 2
F= 1555 N
The error between ADAMS and the hand calculations is because the hand calculations
don’t take into account the damping or the springs. So therefore, we consider that the
ADAMS analysis is accurate.
Figure 15. Torsional Rigidity
The torsional rigidity was calculated in ANSYS by constraining the back of the frame, and
1 center point at the front roll hoop (see figure 15). An upward force was applied on the
left side of the hoop and a downward force was applied at the right side. Using the
displacements that each node encountered, the torsional rigidity was calculated by:




100(9)


 a tan( .0051  .0063 ) 


Fl
2*9
k

 2142.858 ftlb / deg
y1  y 2
12
a tan(
)
2L
K= torsional rigidity
L= half the length of the hoop
Y1 and y2= displacements of node
3.1.1.4: Material Study
According to Formula Hybrid/Formula SAE Competition Rules, the main assembly of the
vehicle is to be made of round, mild or alloy, steel tubing (minimum 0.1% carbon). Other
alternative materials may be used, like aluminum or composite materials, but need to
follow these requirements listed in Formula SAE rules, section 3.3.3.2.1 [15] –
(A) The material must have equivalent (or greater) Buckling Modulus EI (where, E =
modulus of Elasticity, and I = area moment of inertia about the weakest axis)
(B) Tubing cannot be of thinner wall thickness than listed in 3.3.3.2.2 or 3.3.3.2.3.
(C) A “Structural Equivalency Form” must be submitted per Section 3.3.2. The teams must
submit calculations for the material they have chosen, demonstrating equivalence to the
minimum requirements found in Section 3.3.3.1 for yield and ultimate strengths in bending,
buckling and tension, for buckling modulus and for energy dissipation.
For the materials under consideration, we looked at strength, corrosion resistance,
machinability, weldability, availability, and cost.
Material Strength and Corrosion Resistance:
For material strength we looked at the yield strength of the material and the buckling
modulus. Table 3.4 shows a list of strengths of the various materials.
Table 3.4: Material Properties [3]
Material
Tensile Yield Strength
Modulus of Elasticity
Carbon Steel, ie AISI 1010
44200 psi
29700 ksi
Alloy Steel, ie AISI 4130
64000 psi
29700 ksi
Aluminum, ie 6061-T6
40000 psi
10000 ksi
Composite Material, ie carbon fiber
34800 psi
10700 ksi
Machinability and Weldability:
For alloy steel, specifically AISI 4130, the low carbon, about .30%, within the content of
tubing makes for easy welding. This material can be machined by conventional methods.
Machining is best under normalized, tempered conditions. Welding can easily be done by
all commercial methods [2].
For carbon steel, like AISI 1010, this alloy material is easily machined, welded, and
fabricated. It can be machined very well in cold worked condition. It can be welded using
any standard welding technique [2].
Machining and welding Aluminum 6061-T6 is an easy process. For this structure to
adequately hold the stresses with this type of material, the frame would undergo a heat
treatment process. This process would harden the material and therefore make it stronger
to uphold foreseen stresses.
Since carbon fiber material is a layered composite it is treated differently than metals. The
possible variations in layering can cause the material properties to change. Carbon fiber is
not flexible and is often pre-manufactured for custom sizes because cutting is not
recommended due to fibrous debris. Material is not welded; however, it is bonded together
with itself or other materials.
Availability and Cost:
The availability and cost of the each material played a big factor in our decision of frame
material. Although composite materials can have good yield strength these materials are
costly and not easily accessible. Aluminum and steels are commonly available materials.
However, Formula SAE rules state a thickness requirement of the tubing. This limits
availability because not all materials under consideration are available in the required tube
thicknesses. Cost of aluminum and steel are seen in Table 3.5.
Table 3.5: Material Cost [4]
Material
$ - .049” wall thick
$ - .065” wall thick
$ - .095” wall
thick
AISI 1010
$3.24/ft
$2.52/ft
$4.68/ft
AISI 4130
$3.24/ft
$2.52/ft
$4.68/ft
Aluminum 6061-T6
-
$3.24/ft
$3.72/ft
Although carbon fiber material is very lightweight, it is not as strong as steel. Also, it is
expensive and not widely available. Aluminum becomes a much stronger alloy after heattreating but a heat treatment is expensive in time and money. Carbon steel is strong steel,
but is not commonly used for tubing structure. This type of steel is used for bolts and
fasteners. Alloy steel is a strong material which is harder to weld than Aluminum but has
material properties to support our frame. Based on all these facts, we decided to use an
alloy steel, AISI 4130 chromoly steel tubing. It’s easily attainable, strong, and tolerable to
machine and weld.
3.1.2: Body
3.1.2.1: Engineering Specifications
The shell of the vehicle needs to hold with several specifications. The shell must fit over
the frame and the components of the car. Several of the components are mounted on the
outside of the frame and the shell must be formed so that these will fit. Also, the shell
needs to be removable from the frame so that work can be done on the car. A method of
quick removal should be designed for the rear of the car and the side pods in case there
is a problem with the drive system or the batteries. The shell should also be designed so
that there are air vents in the rear to supply cool air to the motor and other electronic
components. Finally the shell needs to be designed to have as little weight as possible so
that the total weight of the car will remain within the specified weight limit. The front end of
the body will be taken from the 2001-2002 Formula SAE car. The side pods and rear end
body design will be designed and fabricated. The body of the vehicle can relay the overall
beauty of the vehicle. Therefore it is important to choose a body design and layout that
will show the eminence of the vehicle.
3.1.2.2: Design History
The shell of the car is designed around the frame and any externally mounted
components.
Figure 16 shows several artists’ renderings of what the complete shell
should look like. The shape of the front body is known because it will be taken off of the
2001-2002 Formula SAE car. The rest of the shell will be created to accommodate any
future changes in the frame.
(a
)
(c
)
(b
)
(d
)
Figure 16. (a) Top View of Conceptual Body, (b) Isometric view of Body Design, (c) Top
view of Body Design, (d) Side view of Body Design
A final revision of the body takes into consideration air vents for cooling the motor and
controller, side pods to accommodate the batteries, and options for paint. These are shown
in Figure 17.
Figure 17. (a) Single color body, (b) Lightning design body
3.1.2.3: Material Study
For the shell of the vehicle there were two types of material that were considered. The
first material was sheet metal, and the second was fiberglass and resin. The sheet metal
was considered mainly because it would be easier to manufacture since fiberglass involves
more time and steps. The sheet metal would need to be bent to the shape of the frame
while the fiberglass would first need a form and then several layers of fiberglass would
have to be applied.
Fiberglass was chosen because while it will take more time to
manufacture it will have a lower weight, and the front body from the 2001-2002 Formula
SAE vehicle can be utilized.
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3.2.1: Suspension and Steering Geometry
3.2.1.1: Introduction
When designing a suspension for a high performance vehicle such as this many different
parameters must be taken into consideration. The wheelbase, track width, ground
clearance; suspension geometry and spring rates are just part of the equation.
Unfortunately, not all of these variables can be optimized at the same time. One
suspension designer described the process with the analogy of trying to reach for the
(a
(b
points of a triangle at the same time, the closer
you get to one point the further you get
)
) of a suspension system is made into an
from another. With this in mind the design
iterative process in which you make decisions with certain parameters in mind and check
to ensure other parameters are not greatly compromised. Finally the designer must choose
what is most important and make compromises accordingly. In this paper we will
investigate some of the parameters involved and the decisions made for the first iteration
of design.
3.2.1.2: Track Width
The definition of track width is the distance of the centerlines of the tires when viewed
from the front. This dimension greatly influences the amount of resistance the vehicle has
to the moment caused by the inertia forces acting at the center of gravity of the car.
Looking at previous year’s cars, as well as other highly competitive schools vehicles, the
track width is determined to be 50 inches. (Note: this method of design by utilizing
outside, as well as previous inside sources will be implemented for many initial baseline
values chosen for the suspension).
3.2.1.3: Geometry
Once the two previous components have been decided upon the geometry of the
suspension can now be addressed. The first is the amount of caster. Caster is the angle
of the steering axis when viewed from the side. Positive caster is defined by the top of the
steering axis being tilted back towards the rear of the car. By implementing positive caster
the outer wheel in a corner will camber negatively and thereby offset the positive camber
induced by the body roll experienced in a corner. A moderate amount of caster also
proves to be beneficial in providing the driver with feedback and good steering feel. For
the first iteration of our front suspension design the car has 5 degrees of positive caster
built into it.
The next parameters to be considered are
Kingpin Inclination (KPI) and Kingpin Offset
(KPO). Kingpin offset (also know as Scrub
Radius) is defined by the amount of
distance between the centerline of the tire
and the point of intersection between the
steering axis and the ground plane (see
Figure 18). This distance affects the amount
of steering force required by the driver.
Small amounts of KPO are beneficial again
for steering feedback, but need to be kept
Figure 18. Scrub Radius
minimal as to not require excessive steering forces. Kingpin inclination (defined as the
angle between the steering axis and the tire centerline) is used to help control the amount
of KPO. Because packaging constraints often forces the KPO to be too great; in this
scenario KPI can be used to offset this and bring the steering axis closer to the mid plane
point on the ground plane. KPI however has the negative drawback of adding positive
camber to the outside wheel in a corner. For our design the wheels we are using have an
appropriate amount of backspacing to allow the uprights to sit deep enough inside the
wheel and provide less than an inch (.915”) of KPO without the need of any KPI. For right
now this value of KPO seems reasonable in light of the lack of need of any KPI.
Next the amount of static camber that will be built into the front suspension geometry is
considered. Camber is defined by the angle that the wheel is offset from vertical when
viewed from the front and is considered negative when the top of the wheel is inclined
closer toward the center of the vehicle. By implementing negative static camber the
positive camber induced by the vehicle rolling in a corner will be counteracted. The amount
of static camber in our vehicle will be 1.5° negative but will be easily adjustable via shims
between the chassis and the suspension pickup points. By using unequal length a-arms
and manipulating the geometry, the amount of camber gain during suspension travel can
be controlled and utilized to ensure that the tire remains as flat as possible on the ground
during suspension travel. Therefore this ensures the contact patch of the tire is maximum
at all times and thereby supplies the maximum amount of grip in corners. Analyzing the
amount of camber change during suspension travel will be part of my tasks for this week.
Last, but far from least, (in fact some designers argue this to be the corner stone of
suspension design) is locating the Roll Center of the geometry. The roll axis (the
imaginary axis the vehicle rolls about in a corner) is defined by the front and rear roll
center points. The roll center point is more clearly defined in illustration than in words (see
Figure 19). After investigation of the illustration, the point is clearly defined by the
geometry of the a-arms. By choosing proper pick up points on both the uprights and the
chassis the geometry and therefore roll centers can be located according to the designer’s
wishes. Based on empirical data the location of the roll center is historically located in a
range just above or just below the ground plane. For our vehicle the front roll center will
be as close to the ground plane as possible. The designer must also keep in mind that
the roll center is an instant center and will move as the suspension travels. Keeping this
roll center in as fixed a position as possible is the designer’s goal. A final layout of the
front suspension geometry is shown in Figure 15.
Figure 19. Roll Center
Figure 20. Front View
In figure 21a, the Kingpin offset is seen in the distance between the red line and the gray
line just to the left of it. Note that these two lines are parallel showing the lack of kingpin
Inclination. Also seen in figure is the amount of static camber that is built into the
suspension. Figure 16b shows the amount of Caster built into the suspension. The first
iteration contains 5° of negative caster.
Figure 21b. Side View
Figure 21a. Front View
Now that the static geometry has been setup in the first iteration of the front suspension
design, geometric analysis can take place. The substantial variables are the Static Roll
Center, Roll center migration, and Camber Gain through suspension travel. Table 3.6
shows the break down of the suspension design and geometric analysis.
Table 3.6: Front Suspension Geometry
Static Camber
Camber Gain in Jounce
Camber Gain in Rebound
Caster
Kingpin Offset
-1.5˚
-1.0353˚/1"
.954˚/1"
5˚
.915"
Kingpin Inclination
0˚
Toe In
0˚
Ground Clearance
1.5"
Static Roll Center
-.17"
Roll Center @ 1" Jounce
-1.56"
Roll Center @ 1" Rebound
1.25"
Camber gain is also an important value in suspension design. Because the A-arms are not
equal in length, nor parallel, the amount of camber seen at the wheel will change as the
suspension (which can be analyzed as a simple four-bar mechanism) travels. As the tire
travels upwards relative to the chassis (Jounce) it experiences negative camber in the
order of 1.0353˚ for every inch of travel. Similarly, as the tire moves down relative to the
chassis (Rebound) it experiences positive camber in the order of .954˚ for every inch of
travel.
Suspension design is an iterative process. With the front suspension geometry values
determined, the data is evaluated to determine if values are in an acceptable range and if
values need to be modified. The amount of migration the roll center undergoes and the
fact that it crosses the ground plane during migration is a concern. The first iteration of the
front suspension geometry analysis was to raise the ground clearance height to 2” (which
is necessary to ensure the chassis never bottoms out in full rebound). This rise in ground
clearance will alter the location of all three listed Roll Centers (RC) and change the
equivalent four-bar’s location and thereby alter the amount of camber gain and loss
through suspension travel.
Figure 22. Suspension Geometric Layout
In figure 22, an equivalent four-bar mechanism (red lines) of the suspension and the
geometric layout. The two circles are the path that the upper and lower a-arms (the top
and bottom red lines) travel on. The left red line is the chassis and the right red line is the
upright. Using this geometry, the roll center migration and camber gain were analyzed.
Figure 22 more clearly presents the equivalent four-bars that represent the front
suspension. Figure 23a shows the suspension at equilibrium, figure 23b shows the
suspension at 1” of Jounce, and figure 23c shows the suspension at 1” of rebound.
(a) Static
(b) Jounce
(c) Rebound
Figure 23. (a) Equilibrium position, (b) Jounce position, (c) Rebound position
Before a complete geometric analysis of the rear suspension, several iterations of the
static Roll Center (RC) in the rear were evaluated to determine a satisfactory static RC.
The static RC for the rear is slightly higher than the front RC to allow the weight to
transfer forward onto the front wheels and thereby increasing the load of the front tires
slightly to improve traction. The only alteration made to the front was to increase the
ground clearance to 2”. This would allow the enough clearance for the amount of
suspension travel desired while still allowing the cars Center of Gravity (CG) to remain low
to improve vehicle dynamics. Tables 3.7 and 3.8 show the new suspension geometry and
geometric analysis of the front and rear suspension.
Table 3.7: Front Suspension Geometry
Static Camber
-1.5˚
Table 3.8: Rear Suspension Geometry
Camber Gain in Jounce
-.95˚/1"
Static Camber
Camber Gain in Rebound
.89˚/1"
Camber Gain in Jounce
-1.05˚/1"
Camber Gain in Rebound
1.02˚/1"
Caster
Kingpin Offset
5˚
.915"
Caster
0˚
3˚
Kingpin Inclination
0˚
Kingpin Offset
Toe In
0˚
Kingpin Inclination
0˚
Ground Clearance
2"
Toe In
1˚
Static Roll Center
1.23"
Ground Clearance
2"
Roll Center @ 1" Jounce
-.18"
Static Roll Center
1.34"
Roll Center @ 1" Rebound
2.65"
Roll Center @ 1" Jounce
0.4"
Roll Center @ 1" Rebound
2.75"
Front Track Width
50"
Rear Track Width
0.02"
48"
Using ADAMS the vertical migration of the roll center during suspension travel can be
determined for the full range of motion. Figure 24 below shows how the program runs the
suspension through its motion by moving the platforms that the tires rest on. The graph is
a plot of the position of the roll center (in mm) through the entire range of motion. The
total maximum displacement of the roll center is just over a half an inch. Note that this
value differs from the previous roll center displacement value because the analysis done
using the equivalent four-bars was for symmetric suspension motion where this analysis
was for the motion of the A-arms in opposite directions. Therefore, the equivalent four-bar
analysis is more relevant for pure heavy acceleration and braking where the ADAMS
analysis is more relevant for cornering. With that said the roll center location for cornering
is more important because it is during a corner that the vehicle is rolling. Keeping in mind
that minimizing the migration of the roll center to maintain predictable and near constant
handling is very important it becomes clear that the small value of roll center movement
achieved is ideal.
The Rear Suspension Geometry can be seen in figure 25 and figure 26. Notice clearance
was created for the suspension clevises to move up and down on the chassis rail to allow
easy manipulation of the geometry and therefore the ability to ideally locate the Roll
Center. When designing a suspension it is nearly impossible to predict just how the car
will handle once it is built. With this in mind it is wise to build in a certain amount of
adjustability whenever possible. For our vehicle we will use suspension pickup clevis that
are not welded directly to the chassis, but rather are bolted therefore allowing shims to be
placed between the clevis and the chassis to adjust the amount of camber built into the
suspension as well as adjust the static roll center, seen in figure 27. Furthermore, in the
rear the clevis are placed on the vertical rails (which will be made of square tubing for
ease of adjustability). In this way the clevises will not only be able to be shimmed, but
also move up or down along the chassis and therefore adjust roll center and camber gain
as well as static geometry (see figure 28).
Left: This figure shows
the suspension
geometry from above,
making the amount of
Toe In visible. (Toe in
Figure 26. Top View of Rear
shown by red lines)
Suspension
Figure 27. Side View of Rear
Suspension
Clevis
Left: This figure shows
Bearing
the amount of caster
of the upright.
(highlighted in red)
Suspension Rail
Right:: This figure
shows how the
clevises can be
shimmed away from
Figure 25. Rear View of Rear Suspension
the chassis.
3.2.1.4: Material Study
Following the logic behind our material use for the chassis, the suspension components
will primarily be made of chromoly tubing. The tube diameter will likely be 5/8 inch based
Figure
28. Shimmed
Clevisesof
on data from passed year’s vehicles. This assumed diameter
as well
as wall thickness
the tubing will be tested and either verified or changed following FEA analysis on the
suspension components. Before FEA can take place the suspension will be run through
some simulations in ADAMS to determine the forces that will be acting on the suspension
components and where they will be acting.
(a
)
Figure 29. (a) A-arm with flattened ends
at bearing rings (above), (b) pressed
bearing welded to the a-arm (below).
(b
)
The manufacturing of the a-arms will include flattening them at the ends to join to the
bearing rings where the bearings mount, as shown in figure 29. The spherical bearings we
will utilize will be mounted to the a-arms by press fitting them into the rings welded to the
ends of the tubes. Aurora Bearing [20], the supplier of the spherical bearings
recommends this process and illustrates it on their website. The bearings to be used will
allow 24˚ of misalignment which is sufficient for our design.
The spring and damper rates are to be determined through hand calculation and tested
using simulations in ADAMS.
3.2.2: Braking
3.2.2.1: Engineering Specifications
According to the SAE rules [15] the brake system must meet the following specifications.
“The car must be equipped with a braking system that acts on all four wheels and is
operated by a single control. It must have two independent hydraulic circuits such that in
the case of a leak or failure at any point in the system, effective braking power is
maintained on at least two wheels. Each hydraulic circuit must have its own fluid reserve,
either by the use of separate reservoirs or by the use of a dammed, OEM-style reservoir.
A single brake acting on a limited-slip differential is acceptable. The brake system must
be capable of locking all four (4) wheels during the test specified below. “Brake-by-wire”
systems are prohibited.
Unarmored plastic brake lines are prohibited.
The braking
systems must be protected with scatter shields from failure of the drive train (see 3.5.1.4)
or from minor collisions.”
At a Formula SAE competition, the brake system is dynamically tested and has to show
the ability of locking all four wheels and stopping the vehicle in a straight line at the end
of an acceleration run. A brake pedal over-travel switch is required on the car. In the
event of brake system failure, this switch will be activated and will stop the vehicle. This
switch will cut the power to all electrical devices. The switch needs to be a toggle switch
such that repeated actuation does not restore power and the driver cannot reset it. The
switch must be executed with analog components, and not through the programmable logic
controller. The car requires a red brake light with at least 15 watts which is clearly visible
from the rear. This light is mounted between the wheel centerline and driver’s shoulder
level on the vehicle centerline laterally.
Figure 30: Brake and
Figure 31: Brake and
Figure 32: Brake
upright Assembly
upright Assembly
Pedal and Master
Cylinder
The components of the brake system were taken from the 2001-2002 Formula SAE car.
The caliper and rotor are built into the upright, as seen in the figures above. This brake
system includes a Wilwood combination “remote” tandem master cylinder, which meets the
Formula SAE specifications [15], calipers with brake pads, rotors, brake lights, and steel
braided Teflon hoses.
3.2.2.2: Design History
The brake systems was collected and inspected to verify all the parts we present and still
in working order. It was determined that the only parts that would need to be purchased
would be steel hard brake lines and brake fluid.
3.2.2.3: Engineering Analysis
Calculations were done to verify that the brakes could provide adequate stopping distance
for the vehicle. It was found that with no sliding the brakes could bring the car moving 80
mph to a stop in 76.45 feet. Detailed calculations can be found in the appendix.
3.2.2.4: Material Study
The rotors are made from hardened steel and meet the specifications for handling the
forces applied during breaking.
Because our vehicle is designed to run primarily in autocross competitions it will need to
be able to accelerate very quickly and stop very quickly to achieve fast times through the
tight and winding course. Therefore, we will use disc brakes with cross drilled rotors on all
four wheels of the vehicle.
3.2.3: Wheels, Tires, and Uprights
3.2.3.1: Engineering Specifications
According Formula SAE rules [15], 10” and a 13” wheels can be used. To reduce the
budget, wheels from the 2001-2002 Formula SAE car are being reused. The wheel shells
are 13” three piece all aluminum shells, from Keizer company [21]. The shell consists of
two pieces that are bolted together along with the magnesium centers shown in figure 33
and 34.
Figure 33. Front view, Wheel shell
Figure 34. Rear view, Wheel shell
and center
and center
The uprights were also taken from the 2001-2002 Formula SAE car. The uprights are
made from aluminum and were manufactured to fit inside on the shells, as shown in
figures 35 and 36. These uprights are acceptable for the suspension design because
the uprights provide the system with correct values of inclination. The uprights’ unique
design includes the calipers and rotors for the braking system.
Figure 35. Front Upright brake
assembly
Figure 36. Rear Upright, brake,
wheel, tire assembly
Tires selection is based on a number of different factors. The diameter of the tire is
chosen based on the selected wheel size. Since the wheel shells are reused, the tires are
constrained to 13” tires. However, this is an optimal size for the vehicle because of the
overall weight of the car. Another factor is the width of the tire. The width of the tire is
dependent on operating temperatures. Once the tires are at operating temperature, the
tires will reach its full handling potential. The wider the tire, the more mass it has, thus the
longer it will take for the tire temperature to rise. Since power conservation is a concern
with limiting battery run time, “warming” the tires before a race will not be an option. To
compensate, a thinner tire is used to reduce the time it would take the tire the reach its
operating temperature. These are Goodyear 13” by 6.5” (D1383, R065 - 18.0x6.5-10)
racing tires (Figure 37). This tire has an operating temperature of approximately 60-70
degrees Celsius.
Figure 37. Goodyear D1385, R065 - 20.0x6.5-13 [7]
3.2.3.2: Design History
Originally two tire widths had been chosen – 7.5” and 6.5”. The 7.5” tire was a possibility
because it was readily available. The preferred tire is the 6.5” tire since it weighs less and
can reach its operating temperature faster than the 7.5” tire. The first obstacle was to
ensure the 6.5” tires would fit on our wheel shells. Wheel shells can handle slightly
different tire dimensions. The Keizer wheel shells can handle both 6.5” and 7.5” tire width
and sustain pressure. A simple design calculation was applied to determine the distance it
takes a tire to reach its operating temperature was applied. It was found that the 7.5” tire
will take 1115.98 feet and the 6.5” tire only 772.54 feet. Based on these conditions, the
6.5” width tires are preferred.
3.2.3.3: Engineering Analysis
The following calculations show the distance it will take for the different tires to reach their
operating temperature. These calculations prove the 6.5” width tires can reach full
potential at a shorter distance than the 7.5” width tires. A complete calculation can be
found in the appendix.
d
m tireCv(Tdesired  Tamb )
mcar g
7.5" Tire :
(5.896 7 kg)(1600 J/kg - K )(343.15 K  298.15 Kelvin )
 340m  0.21miles  1115.98 ft
1.7(74.842 7 kg)(9.81 m/sec 2 )
6.5" Tire :
d
d
(4.082 3 kg)(1600 J/kg - K )(343.15 K  298.15 Kelvin )
 240m  0.15miles  772.54 ft
1.7(74.842 7 kg)(9.81 m/sec 2 )
3.2.3.4: Material Study
Tires selection is based on weather, dry or wet, compound, and size. For this application,
tires will be dry slicks. Since these tires are not very diverse, the material is made on only
one type of compound – R065 compound. Sizes of the tires vary, but for this application a
13” rim with a 6.5” tire width is necessary. These types of tires are not commonly used
and only sold from two companies, Goodyear and Hoosier. The prices for these tires are
comparable, $153 from Goodyear [7] and $133 from Hoosier [6]. From research, it was
found that the performances between Goodyear and Hoosier tires were comparable. With
good contacts and the possibility of tire donations, the Goodyear tires have been chosen.
3
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3.3.1: Motor and Power Train
In autocross racing the vehicle is most often in the range of 30-40 mph with peak speeds
not much higher than 60mph. The largest factor in autocross performance (regarding the
power train) is the ability of the car to accelerate to that high speed of 60mph. A
calculation of time needed to accelerate from 0 to 60mph for a given motor configuration
is taken as a representative estimate of the performance of that configuration in autocross.
This calculation also gives the added bonus of giving the average person a common point
of comparison with the quickness of the REV.
On courses with longer straight-aways speeds of as much as 85mph may be achieved. To
take advantage of these the car also needs to be able to reach 85mph. The requirement
for the 0-60mph time was set at 5 seconds; however, as a race car, the faster it was
capable of the better.
3.3.1.1: Engineering Specifications
There are several standards that the motor mount must conform to. The first standard is
that the mount must be designed to fit within the frame, and to fit the specifications of the
motor (set attachment points). Secondly, the mount must be designed to hold the engine
weight as if the motor were acting as a cantilever beam as shown in figure 38.
This
standard is set so that if the rear motor mount were to fail the motor would not be subject
to damage that would result in failure of both mounts. The third standard for the mount is
that it must be able to withstand the maximum torque that the engine puts out.
This
standard is set so that if the drive shaft becomes jammed and the engine itself is
attempting to rotate the motor mount will be able to withstand the torque placed on it.
The final standard is put in place for the rear motor mount, and it is that the mount must
be sized to create a restraint at the rear of the motor so that it is not acting as a
cantilever beam off of the front mount.
Front Motor
Mount
Motor
Attachment
Points
Weight of
Motor
Figure 38. Motor Diagram Without Rear Motor Mount
3.3.1.2: Design History
The motor and power train as a system is absolutely crucial to the design of the car. The
motor selection process began with evaluating the motors by their power, torque, and
efficiencies in comparison with the weight, battery, and power train tradeoffs. After some
basic research into light weight, powerful motors four were given serious consideration.
They were three sizes of motors by Advanced DC (from smallest to largest - the A00, the
203-06-4001, and the FB01) and one motor by Netgain Technologies (the Warp 9).
Qualitative factors such as durability, ease of drive train implementation, and configuration
flexibility were considered, but quantitative comparisons needed to be made. To do so a
performance calculator was created.
The design for the motor mount went through several stages cumulating in the present
design.
The design originated from a combination of several different electric motor
mounts that were found during research on the subject. This design involved an vertically
mounted motor and a damper system to reduce vibrations of the motor.
The original
design was then altered for two reasons. First, the vertical mounting would cause the
center of gravity of the car to become too high. Secondly, for our purposes the vibrations
were not going to be significant enough to warrant a damper.
These changed the
mounting to its second phase which was placed in CAD and attached to the skeleton of
the car. Calculations were applied to this mount and it was found have a factor of safety
of over 10. The present mounting configuration was created in a group meeting when it
was suggested that two vertical bars might be enough to satisfy the engineering
specifications. So, the forces that would be applied to these bars were determined (shown
in Engineering Analysis Section) and then the same calculations were applied to this
configuration.
This configuration was found to support the same loading as the second
design, and so after comparing the weight of each of these designs the newest one was
found to be lighter.
So, the two bars were added as the front motor mount, and two
straps were placed as the rear mounting (as shown in figures 39 and 40).
Figure 39. Front Motor Mount Shown With Partial Frame
Figure 40. Rear Motor Mount Shown With Partial Frame
3.3.1.3: Engineering Analysis
Performance Calculations
Below is an example of the performance calculations used to evaluate the motors with the
FB01 9” dc motor at 144V and a maximum current of 550 amps used as an example.
Weights, distances, and torques are originally known in lbs, feet or inches, and ft-lbs and
are later converted to SI units for the actual acceleration calculations.
First the average torque is found over the rpm range needed to achieve 60mph. The plot
of the speed torque data used can be seen below This is a plot of the data points taken
from the supplier’s speed-torque curve, with lines connecting the points to visualize the
area over which the torque is averaged.
FB01 motor, 0-60mph Speed Torque curve
140
120
Torque (ft-lbs)
100
80
60
40
20
0
0
500
1000
1500
2000
2500
3000
3500
4000
4500
5000
RPM
The area under the curve is found by finding the average torque for each line segment
(the height) and then multiplying by the rpm range (the width) which gives the area of
that section. The sections are then summed and divided by the rpm range (the total width)
to get the average torque (aka. the average height).
 T1  T 2 

T2 T3
T3T4 
T4 T5
T5T6 
 2  *1000   2  *1000   2  *1000   2  *1000   2  * 300 / 4300  Tavg











 127  127 

 127  127 
 127  127 
 127  72 
 72  55 
 *1000  
 *1000  
 *1000  
 *1000  
 * 300 / 4300  Tavg  116.17 ftlbs

2
2
2
2
2











Ti (i=1-6) represents the torque (in ftlbs) at each rpm starting at 0 and going though
4300 rpm (4300rpm is approximately the rpm that coincides with 60mph). Tavg is the
average torque at the motor shaft.
Tavg * GR  Tw
116.17 ftlbs * 4.375  Tw  508.26 ftlbs
Tw
 Fw
  win 1 ft 
*


12in 
 2
508.26 ftlbs
4.448 N
 592.15lbs *
 Fw  2633.88newtons
20
.
6
in
1
ft
1lb


*


12in 
 2
Using GR (the gear reduction in the differential) the average torque at the wheels, Tw, is
obtained. Then using Өw, the tire outer diameter, the average force at the wheel tread, Fw,
is obtained.
Before further performance calculations can be done, drag, weight, and rolling resistance
need to be calculated.
 b  .00327 * C d * A F * U 2  
 
dU 
 

3

 
a

 Davg
ba
 60mph .00327 * .4 *12ft 2 * U 2  


 
  
3
 
4.448 N
 0 mph
 Davg  18.84lbs *
 83.78 Newtons
60
1lb
Cd is the coefficient of drag, AF is the frontal Area, and U is the speed in miles per hour.
The integral of the drag from 0 to 60mph is found and then divided by the speed range to
get the average drag force, Davg,.
N s*N p*Wc  WB
48*12*.15lbs  WB  86.4lbs
WD  WM  WB  WRC  WCN  WO  WT
150  143  86.4  200  23  50  WT  674.2lbs
2.205lbs
 MT
1kg
2.205lbs
674.2lbs*
 M T  304.94kg
1kg
WT*
Ns is the number of battery cells in series, Np is the number of series sets in parallel, and
Wc is the weight of each cell. The weights are WD for driver, WM for motor, WB for all the
batteries, WRC for the rolling chassis, WCN for the controller, WO for all other, and WT for the
total. MT is the total mass. Both the weight and mass will be used in further calculations.
WT * .02  FR
674.2lbs * .02  13.484lbs *
4.448N
 FR  59.98newtons
1lb
FR
 FRavg
2
59.98
 FRavg  29.99newtons
2
FR is the rolling resistance at 50mph and rolling resistance is estimated as a linear
function of speed, thus the average rolling resistance (FRavg) for the 0-60mph acceleration
is approximated as ½ of FR.
Knowing the force at the wheels and the losses the net force (the force that provides
acceleration) can be found. Using this net force and the mass the acceleration and then
the time from 0 – 60mph can easily be found. A comparison between peak force and
traction is found at the end of these calculations. The net force is just slightly higher than
the traction, which will cause some wheel spin. This however may be canceled out by the
rotational inertia of motor and other components which have not been taken into account
in this calculation.
Fw  Davg  FRavg  Fnet
2633.88newtons  83.78newtons  29.99newtons  2520.11newtons
Fnet
 Aavg
MT
2520.11newtons
m
 8.264 2
304.64kg
s
VF
t
Aavg
0.44704
60mph *
m
s
mph
m
8.264 2
s
 t  3.246s
Aavg is the average acceleration, VF is the final velocity (60mph), and t is the time to that
final velocity. Thus the time needed to achieve 60mph from a standing start is calculated
3.246 seconds.
Friction Force vs. Motor Force Comparison
WT * R wd *  FNR
674.2lbs * .55  369.82lbs
FNR * CoF  FFR
369.82lbs *1.7  628.69lbs *
4.448newtons
 2796.41newtons  FFR
1lb
T P * GR
 FP
  w in 1 ft 
*


 2 12in 
127 ftlbs * 4.375
4.448newtons
 FP  647.33lbs *
 2879.321newtons  FP
1lb
 20.6in 1 ft 
*


12in 
 2
Rwd is the rear weight distribution, FNR is the rear normal force (for both wheels together),
and CoF is the coefficient of friction (estimated number from users in the Formula SAE
forums for broken in Hoosier slicks, which has been assumed to be the value for the
Goodyear slicks as well). FFR is the rear friction force and FP is the peak force (provided
by the motor).
These calculations neglect the efficiency of the differential (which provides the gear
reduction, but should be around 97% or better for spiral bevel gears), efficiency of the CV
joints, and the effects of rotational inertia, all of which would increase the 0-60mph times
by between .25 and .5 seconds.
The analysis of the front motor mount consisted of several hand calculations and a weight
calculation preformed in Pro-Engineer Wildfire 2.0.
The first thing that had to be done
was a diagram showing the forces that the motor was going to apply to the bars of the
mounting. This is shown below in figure 41, and after these forces were found they were
resolved
into X and Y
components.
These
forces
find
the
in
the
were used to
shear
stress
bolts,
the
bearing
stress
on the bars,
and
the
stress
in
the
bar.
The
equations
for
these calculations are found in appendix 7.1.
Figure 41. Left hand image shows the forces the motor causes, torsional and weight.
Right hand image is the free body diagram with the reaction forces.
RyTop
RxTop
F1  T  d  (127 ft  lb) * (4.20in)  (1524in  lb)  (4.20in)  3093.72lb
 Fx   F1 sin( 45)  F2 sin( 45)
W
2
F1x
F1  F2
 Fx  0
F1y
 Fy   F1 cos( 45)  F2 cos( 45)  2W
F2x
W
2
F1  F2
 F1 cos( 45)  F2 cos( 45)   F
F2y
 143 
 Fy  2W  F  2
  (3093.72)  3165.22lb
 4 
RxBottom
RyBottom
Free Body Diagram
PL (3165.22)(8)

 2.5578in
AE
(.25)( 29700)

2.5578


 0.42629
L0
8

  E  29700(.42629)  12660.88
FS 
lb
in
 Y 63100

 4.78
 12660
3.3.1.4: Material Study
Motor Selection
The performance numbers took into account the variations in motor and battery weight
along with the variations in torque speed curves. Each motor was compared with a gear
ratio that made the peak force at the wheels just higher than the rear friction force so as
go get the best possible accelerating force while still allowing the ability for slight wheel
spin.
Two A00 motors were to be hook up one to each wheel with electronic differentiation,
while the other motors would have been each a single motor running to a differential. The
A00 setup was the lightest, but the drop off of in torque at higher rpms, due to the small
motor size, caused poor 0-60mph times. The 203-06-4001 motor did well, but not as
well as the FB01 and Warp 9. These two 9” motors did the best in the performance
comparison, and also provided the greatest efficiency, the ease of a single motor
configuration (which only requires a single gearbox for reduction), the greatest flexibility for
gear ratios (being larger they could provide more torque easily if a differential with a high
enough gear ratio could not be found), and the greatest continuous horsepower (the
amount of horsepower they are able to provide indefinitely without overheating).
The Warp 9 did not have data available for it at high
currents and voltages, but conservative extrapolations
showed that it would provide better performance than
Warp 9 motor by Netgain Technologies [11]
the FB01 with greater efficiency. This level of performance was confirmed by the
experiences of various sources in the electric vehicle community. The Warp 9 also had
larger commutators and advanced timing (which reduces arcing). NetGain Technologies,
LLC designed the Warp 9 with these advantages specifically for electric vehicles, while
Advanced DC builds motors only for general applications. Thus the Warp 9 was selected
for REV.
The Warp 9 motor weighs 156lbs, is rated for 32.3 continuous hp, will have an estimated
73hp peak in our configuration, and will output an estimated 127ftlbs at 550amps
(controller limited maximum) from 0 to 3000 rpm. The estimated 0-60mph time for the
REV with the Warp 9 motor and a 4.375:1 gear reduction was 3.237 seconds. After a
97% efficient spiral bevel gear reduction in the differential, some CV joint losses, and
accounting for the rotational inertia of the components, the 0-60mph time should still be
under to 3.75seconds. This greatly exceeds our acceleration requirements.
The ability to change gear ratio would help REV obtain a higher top speed, but would
have negligible advantages during an autocross race. A transmission would also add
significant weight (an obvious disadvantage). Thus the REV has no transmission. Based
on comparing the rear friction force to the peak motor force it was determined that a gear
reduction between 4 and 5 to 1 is needed in a differential. A differential with this gear ratio
provides the necessary gear reduction without the
addition of another gear box. For the sake of
racing performance a differential with limited slip
is preferred. When one wheel on a car with a
regular (open) differential slips all the torque
goes to the slipping wheel and none goes to the
wheel with traction. In racing conditions some
wheel
slippage
acceleration
and
can
be
expected
cornering.
A
in
limited
heavy
slip
differential causes more of the torque to go to the wheel with traction, thus increases
acceleration and control.
The front differential in Kawasaki 4x4 ATV’s (Bruteforce or Prairie of any size) 2002 or
newer has a 4.375:1 spiral bevel gear reduction and limited slip capabilities. It is also
lockable which allows full torque to be given to
both wheels regardless of whether one is slipping
Kawasaki Prairie 700 4x4 Front Differential [22]
or not. If activated, this will increase friction in corners, but increase control and
acceleration in drag race conditions, or 0-60mph time tests. This differential was the only
one to meet the above criteria for gear ratio and limited slip capacity. After inputting this
4.375:1 reduction into the performance calculations it provided just enough peak force over
rear friction force to allow the desired the ability for some wheel spin at peak force while
maximizing acceleration. This small margin of peak force over rear frictional force also
provides for some efficiency and rotational inertia losses while sill maintaining peak, or
near peak acceleration. The limiting factor for acceleration is the rear friction force, and
thus peak acceleration is when the peak force matches the rear friction force.
Using a
safe maximum motor speed of 6500 rpm (series wound motors have no fixed unloaded
speed, but their lifespan exponentially decreases with higher speeds) and the 4.375:1
reduction the top speed of the REV would be approximately 84mph. This also exceeds our
design requirements.
For the motor mount the materials chosen were limited by how they are going to be
attached to the frame. Originally an aluminum alloy was going to be used since it would
be resistant to the weather it would be exposed to, and it was a lightweight material. After
changing the design though aluminum became impractical since the front mount is going to
be welded to the frame which means that it should be the same material as the frame.
To keep the material consistent throughout the frame of the car the motor mount will be
made of AISI 4130.
Manufacturing of these pieces is going to be relatively simple
because it is simply drilling four holes, but they also need to be done with high tolerances
so that they will match up with the mounting points that exist on the engine.
3.3.2: Power Source
3.3.2.1: Engineering Specifications
Side-Pods/Batteries
-
Battery Temperature Range
o -30°C to +60°C
-
Side-pod Envelope
o Should not obstruct driving
o Be within a reasonable size

Not farther out than mid-plane of front wheels

Not higher or lower than driver’s area side beams, between main
hoop support and the next hoop forward.
-
Number of Batteries
o According to EE’s there will be approximately 600 batteries needed
o each side-pod will then have 300 batteries
-
Battery Pack Container
o completely isolate batteries and their connections from the rest of the
chassis and body
o hold 300-315 batteries
o batteries easily accessible
o reduce movement and vibration of batteries during usage of vehicle to
keep batteries in contact
o 48 batteries in series, 12-13 48-battery packs in parallel to get
approximately 144V and 25Ah
o Container easily removable from side-pods
-
Battery Connection
o Highly safe, no chance of electrocution
o connection between battery pack container and controller
o easy to connect and disconnect battery packs for faster battery pack
switching
3.3.2.2: Design History
Current Battery and side pod configurations
Battery packs
Side Pods
(Frame support)
Front
of car
Figure 42. Current Side pod and battery
Side-pods contain battery packs that
are 7 columns of batteries wide, 15
columns deep, and 3 tall. This will
get a total of 315 batteries per
battery pack. This configuration
will give 30 batteries more than the
600 total batteries needed.
layout
Batteries (3)
Connectors
PVC Containter
Figure 43. Battery column layout
Battery Configuration
The above battery packs will be isolated from the chassis and the rest of the car. This will
done using a box made of a non-conducting material (i.e. plastic). The box will be easily
removed from the side-pods per engineering specifications. This container has not been
designed as of yet.
Within each column of batteries there are three (3) batteries. Between two of the batteries
there will be disc connectors made from a copper disc surrounded by a non-conducting
(plastic) annulus. The annuluses are used to hold the copper connector in position. The
columns of batteries will be held in place by a tube of PVC with a slot cut out to allow for
thermocouples to be placed on the batteries. Each column of batteries will be held in
compression.
3.3.2.3: Engineering Analysis
Battery Configuration
The following calculations were used for an excel sheet that helped determine
whether or not the current battery configuration would meet heat transfer specifications. As
of right now the specification is that the rate of heat generated by the batteries must be
less than the rate of heat transfer. All equations are taken from Fundamentals of Heat and
Mass Transfer [17].
The governing equation for rate of
SL
heat transfer from a liquid flowing
through the bank of tubes is as
following:
q  N (hDTlm ) L
SL
V, Ti
(1)
Sd
Where N is the total number of tubes, h is the convection coefficient, D is the diameter of
the tubes, ΔTlm is the log mean temperature, and L is the length of the tubes.
The Reynolds number is calculated
Figure 6. Staggered Tube Arrangement
using the maximum fluid velocity. It is
used in calculating the convection coefficient, h.
Vmax D

ST

V
ST  D
Re D ,max 
(2)
Vmax
(3)
Vmax is the maximum fluid velocity, in our case air, within the tube bank. V is the velocity
of the air flowing into the tube bank, in our case it would be approximately the speed of
the car. The constant Pr is the Prandlt number for air at the inlet temperature Prs is the
Prandlt number at the highest possible temperature, the surface temperature of the tubes.
C and m are constants given in a table [17].
Nu D  C Re
h  Nu D
m
D , max
Pr
0.36
 Pr

 Prs



1/ 4
(4)
k
D
(5)
Now all that is needed to find is the log mean temperature, Tlm, to find the rate of heat
transfer. To find log mean temperature the outlet temperature must be estimated by the
following equation:

DNh
Ts  To  (Ts  Ti ) exp  
 VN S c
T T p

(T  Ti )  (Ts  To )
Tlm  s
 T  Ti 

ln  s
T

T
o 
 s




(6)
(7)
The final step is to compare the heat generated by the batteries and the rate of heat
transfer of the flow of air through the side pod.
(8)
qbank  qelec
These calculations have only been roughly done. We assumed that since the operating
temperature of the batteries is between -30˚C and 60˚C that the highest surface
temperature of the batteries will be 60˚C. Using this temperature the rate of convection
heat transfer per unit length is approximately 235 W/m with approximately .5 inch
between batteries and the inlet air speed at 30 mph.
The design that these calculations were made for was abandoned due to the fact that the
batteries would be held together through soldering between batteries. The manufacturer of
the batteries advised against any soldering on the batteries due to heat from the soldering
damaging the batteries.
3.3.2.4: Material Study
The decision on what would be the power
source was an important decision. The power
source had to be able to handle the high
current and voltage pulls that was required for
a large enough engine to perform up to our
goals. We looked seriously into two types of
batteries, Pb-acid and Li-ion. The Pb-acid
would
be
suitable
because
they
were
dependable and easily available. They were
also the cheapest of the batteries that were
looked out. The draw back to them would be
the weight of the Pb-acid batteries. To have
enough current and voltage the electrical
Figure 44. A123Systems lithium-ion
rechargeable ANR26650M1 cell
engineers deemed that upwards of 10-12 batteries with each weighing approximately 40
lbs would be needed. With that much weight and just the size of 10 batteries the size of
car batteries the project would have to be changed quite dramatically. The vehicle would
have to increase in size until the power to weight ratio became higher.
With further searching into the possibility of lithium-ion batteries, a relatively new battery
that had been used in other high current and voltage applications was found. The
A123Systems lithium-ion rechargeable ANR26650M1 cell was light weight and had high
amperage. Each cell only weighs 70 grams and only approximately 600 cells would be
needed. The draw back to these batteries was the high price. Each cell costs
approximately $18-$20 depending on where they were purchased. The power to weight
ratio of the batteries was considered great enough to warrant the price and so these
batteries were finally chosen.
3.3.3: Cooling
3.3.3.1: Conceptual Design
In the initial design phase the idea was to air cool the controller. This decision was revised
after further investigation of the manual of the controller. It suggested that air-cooling was
possible but only with intermittened use and low amperage usage. Since our usage is
beyond those specification and the fact the controller is initially designed for water-cooling
the decision was made to investigate the possibility of water-cooling.
The system would have to be small, light weighted, simple and cheap. It had to be able to
provide 2 gallons per minute (120 Gal/hr) flow rate across the component. Also the pump
has to run low voltage (12-36V). The cooling for this system will be provided by forced air
thru a radiator at opening in the vehicles chassis. The internal component ideally is to be
kept at below 55 C due to manufacturers specifications. The heat dissipated is about 2
watts per amp of current.
3.3.3.2: Material Study
The components found suitable for this project are mostly based off of computer processor
water-cooling systems. These meet the specifications given and provide adequate heat
dissipation to cool the water. Also a closed overflow container with a pressure valve will be
used to contain the extra water. The system will be closed to avoid contact of water with
the electronic components in direct vicinity. The estimated budget is around $150.
3
3..4
4:: E
Elleeccttrriiccaall
3.4.1: Electrical Specifications and Interface Requirements
High Voltage (HV) Requirements
There must be no connection between the frame of the vehicle (or any other conductive
surface that might be inadvertently touched by a crew member or spectator), and any part
of any HV circuits.
HV and low-voltage circuits must be physically segregated:
• Not run through the same conduit.
• Where both are present within an enclosure, separated by insulating barriers.
• Both may be on the same circuit board.
No Exposed Connections
No HV connections may be exposed. Non-conductive covers must prevent inadvertent
human contact. This would include crew members working on or inside the vehicle. HV
systems and containers must be protected from moisture in the form of rain or puddles for
any car that is certified to run rain or wet conditions. There will be no HV connections
behind the instrument panel or side switch panels. All controls, indicators and data
acquisition connections must be isolated using optical isolation, transformers or the
equivalent.
HV Insulation, Wiring, Insulation, and Conduit
All insulation materials used in HV systems must be rated for the maximum temperatures
expected. Insulated wires must be commercially marked with a temperature rating. Other
insulation materials must be documented.
All HV wiring must be done to professional standards with appropriately sized conductors
and terminals and with adequate strain relief and protection from loosening due to vibration
etc.
All HV wiring that runs outside of electrical enclosures must be enclosed in orange nonconductive conduit. The conduit must be securely anchored at least at each end, and must
be located out of the way of possible snagging or damage.
Contactors (Drive Current)
Contactors shall be enclosed in a fireproof shield and shall not be located in the driver's
compartment.
Fusing
All electrical systems must be appropriately fused. Any wiring protected by a fuse must be
adequately sized and rated for current equal to the fuse rating. A separate main fuse shall
be placed in series with the Drive Battery output. The fuse rating shall not exceed two
hundred percent (200%) of the maximum drive current requirement. The fuse shall have
an interrupt rating of at least 20,000 amps. Fuses shall be rated at a higher DC voltage
than the nominal system voltage.
Safety Equipment
The team must have the following:
• Insulated cable cutters, rated for at least the voltage in the HV system.
• Insulated gloves, rated for at least the voltage in the HV system.
Master Switches
The vehicle must be equipped with two master switches. Each switch must stop the
engine. The international electrical symbol consisting of a red spark on a white-edged blue
triangle must be affixed in close proximity to each switch with the “OFF” position of the
switch clearly marked.
Primary Master Switch
The primary master switch must be located on the (driver’s) right side of the vehicle, in
proximity to the Main Hoop, at shoulder height and be easily actuated from outside the
car. This switch must disable ALL electrical circuits, including the battery, alternator, lights,
fuel pump, ignition and electrical controls. The primary master switch must be of a rotary
type and must be direct acting, i.e. it cannot act through a relay. All battery current must
flow through this switch.
Cockpit-mounted Master Switch
The type and location of the cockpit-mounted master switch must provide for easy
actuation by the driver in an emergency or panic situation. The cockpit-mounted master
switch must cut power to the ignition. The cockpit-mounted master switch may act through
a relay.
Quick Disconnect
The steering wheel must be attached to the column with a quick disconnect. The driver
must be able to operate quick disconnect while in normal driving position with gloves on.
Sensors
The PLC shall have the capacity for the input of the following sensors: Thermal, Voltage,
Current, and Encoder.
Interface System Requirements:
Menu
The EZTouch PLC will provide access to system data through a touch tab menu system.
This menu will provide vital information on the batteries, motor, speed as well as monitor,
limit and shut down the system if necessary.
under software design details.
Warning Lights
The menu is discussed in greater detail
Programmable Logic Controller (PLC) must provide operator with battery disconnect, check
engine, and check battery lights. The PLC monitors the temperature of the HV battery
pack and motor, and controls yellow warning light on the instrument panel.
3.4.2: Technical Hardware Design
The block diagram below describes the general layout of our electrical system. The main
components of our system include the batteries, motor, controller, programmable logic
controller, monitor, user interface, wireless interface, voltage/current divider, and electrical
shutoff. We have two sets of batteries: the 144V main power supply and the 24V auxiliary
batteries. These battery sets are made up of many Lithium Ion cells in series and parallel.
Our DC series wound motor comes from NetGain Technologies, LLC, and is specifically
designed for electrical vehicles. The programmable logic controller (PLC) and motor
controller monitor and control all aspects of the car system. The monitor displays on the
steering wheel showing status of the system with warnings. The wireless interface enables
team members to monitor the system remotely. A laptop serves as the user interface to
log and record the performance and status of components. Sensors include a
voltage/current divider, thermocouples, encoders, and light sensors. The voltage/current
divider monitors the battery voltage at different points. Thermocouples monitor the
temperature in the batteries, motor, and controller. The encoder measures the revolutions
of the motor which enables calculation of the speed and distance traveled of the vehicle by
the PLC in real time. The light sensors communicate the status and any errors of the
controller to the PLC. For safety, the electrical shutoff turns off all power from the batteries
to the controller when the digital input for the electrical shutoff is sent to the PLC.
T
Batteries 144V
E shutoff
(power
mosfet)
Voltage/
Current Divider
T
E shutoff digital inputs
and reset
Motor
E
T
Programmable Logic
Controller
Motor Controller
E
L
S
Batteries 24V
Serial/Ethernet
Wireless
Interface
Monitor
Wireless
T
Thermocouple
E
Encoder
L
Light Sensors
S
Speed Inputs
User interface
Major hardware components and their technical specifications are listed below:
DC Motor
NetGain Technologies, LLC [11]
Part Number: WarP 9
Length: 15.70 in
Diameter: 9.25 in
Weight: 156.0 lbs
Input Voltage: 96-144V
Data at 144V input
Time On
Volts
Amps
RPM
HP
KW
5 min.
134.0
320
4200
48.80
36.80
1 hr.
138.0
185
5700
30.40
22.90
Continuous
139.0
170
6000
28.50
21.50
Peak Horsepower
100.00
Zilla Controller
Café Electric [8]
Part Number: Z1K-LV
Length: 9.00 in
Width: 7.00 in
Height: 4.63 in
Weight: 15.5 lbs
Maximum Motor Amps: 1000 A
Nominal Battery Voltage: 72 – 156 V
Peak Power: 320,000 Watts
Dimensions of the Hairball interface:
Length: 10.00 in
Width: 3.5 in
Height: 1.75 in
Hairball interface required to run the Zilla controller. It enables many driving and safety
features.
Curtis Throttle Control (Pot Box)
Curtis [8]
Part Number: PMC #PB6
Length: 1.875 in
Width: 4 in
Height: 3.75 in
Weight: 0.625 lbs
Resistance: 0 – 5 kΩ
Speed Sensor
Café Electric [8]
Part Number: 2171S
Nominal Voltage: 12V
Fuse
FERRAZ/SHAWMUT
Part Number: A30QS600-4
Ampere Rating: 600 A
Dimensions (in):
A: 3.13
B: 1.22
C: 1.63
D: 2.44
E: 2.31
F: 0.31
G: 1.00
H: 0.19
Fuse
FERRAZ/SHAWMUT
Part Number: A30ZS800-4
Ampere Rating: 800 A
Dimensions (in):
A: 3.13
B: 1.22
C: 1.63
D: 2.44
E: 2.31
F: 0.31
G: 1.00
H: 0.19
Thermocouple
Omega [10]
Part Number: SA1
Temperature Range: -60˚C to 175˚C
Insulation: Teflon
Size: 25 x 19 mm
Length: standard 1 m
Wireless Interface Adapter
New Micros, Inc [9]
Part Number: XBEE PlugaPod-S
Size: 1.3" x 1.5"
Weight: 0.4 oz
Small C – freeware, includes limited assembler
512 words Program Ram
24 General Purpose Digital I/O lines share functions with
4 wire SPI Interface
Power requirement for the PlugaPod is 6-9VDC @ 300mA or higher.
EZPLC
EZAutomation [12]
Part Number: EZPLC-D-96E
12 module slots
96 I/O
Communication: Ethernet and serial
Nominal Voltage: 24VDC
Lithium Ion Batteries
A123 Systems [18]
Part Number: ANR26650M1
Nominal capacity and voltage: 2.3 Ah, 3.3 V
Max continuous discharge: 70 A
Operating temperature range: -30˚C to +60˚C
Weight: 0.154 lbs (70g)
3.4.3: Technical Software Design
The REV software package will consist of 3 main parts that must be integrated at a user
terminal. The PLC, motor controller, and wireless transmitter will be on the vehicle. A
remote wireless receiver will be on a user interface terminal. The PLC and the motor
controller will connect to the wireless transmitter via serial communication and they will
also connect to the user interface terminal.
See PLC software description, the wireless
communication, and the motor controller program tree below.
EZPLC-D-96E:
The EZPLC from EZ Automation uses a simple software editor to create Relay Ladder
Logic (RLL). The Relay Ladder Logic integrates 12 I/O modules with the controllers and
monitors the race car’s vital drive train parts.
Power-up Initialization
At power-up, the CPU initializes the internal electronic hardware. It also checks if all the
memories are intact and the system bus is operational. It sets up all the communication
registers. It checks the status of the back up battery. If all registers are go, the CPU
begins its cyclic scan activity as described below.
Read Inputs
The CPU reads the status of all inputs, and stores them in an image table. IMAGE TABLE
is EZPLC’s internal storage location where it stores all the values of inputs/outputs for
ONE scan while it is executing ladder logic. CPU uses this image table data when it
solves the application logic program. After the CPU has read all the inputs from input
modules, it reads any input point data from the Specialty modules like High Speed
Counters.
Execute Logic Program
This segment is also called Ladder Scan. The CPU evaluates and executes each
instruction in the logic program during the ladder scan cycle. The rungs of a ladder
program are made with instructions that define the relationship between system inputs and
outputs. The CPU starts scanning the first rung of the ladder program, solving the
instructions from left to right. It continues, rung by rung, until it solves the last rung in the
Main logic. At this point, a new image table for the outputs is updated.
Write Outputs
After the CPU has solved the entire logic program, it updates the output image table. The
contents of this output image table are written to the corresponding output points in I/O
Modules. After the CPU has updated all discrete outputs in the base, it scans for the
specialty modules. The output point information is sent to the specialty I/O like counters.
Subroutines
The CPU executes subroutines when called for in the ladder program.
Monitor Display (Main Screen):
Monitor Battery Status:
PlugaPod XBee wireless:
The wireless controller software package includes the concise editor, complier, assembler
(ECA) program called Small C, and the virtual terminal (NMI Terminal). The user defined
program will be implemented in the virtual terminal and sent digitally through the special
JTAG cable to the XBEE Doggle, where it is wirelessly sent to the onboard system. We
opted to use the limited version of the ECA program instead of the full version, due to its
adequacy for our application requirements.
The flow chart developed thus far is as
follows:
Zilla Hairball II:
The Zilla Motor Controller Package comes with the Hairball 2 Interface. Through the serial
port of a computer, the interface uses abbreviated menus to allow the user to change
values in the controller.
3.4.4: Schematic
3.4.5: Reliability and Maintainability Assessments
Due to the scale of our project, the reliability and maintainability will be vital to ensure the
car is kept in race condition. Most of the electronics used in our car have a life limit. To
guarantee that all parts make it to their projected life time all safety precautions and
guidelines will be followed in the user manuals.
A123 System’s M1 – The car’s lithium ion cell is the next generation of lithium ion, but it
still is a battery.
After 1000 complete discharges the battery maintains only 75% of its
original power. To avoid losing this power after only 1000 lifecycles, the PLC will turn the
car to a low voltage setting until the discharged cells are replaced with charged cells.
Another potential danger to the lithium ions is overheating.
While this particular battery
does not contain any phosphorous, it is still at risk of overheating due to high discharge
rates.
This will be counter acted by 16 thermocouples that will monitor the batteries
temperature at all times during operation.
NetGain Technologies’s WarP 9 – The integrity of the motor should be maintained if the
motor is handled with care prior to installation and carefully limited voltages and currents
are applied to ensure the motor does not reach 7000 rpm. The series wound motor is
capable of overheating and damaging itself if proper care is not taken. All safety manuals
will be carefully review before motor testing and followed during motor testing and racing.
Café Electric Zilla 1K – The Zilla 1K is made specifically for racing or high output of series
wound motors.
The controller is notorious for overheating if proper procedures are not
taken to cool the controller.
implemented prior to testing.
The Zilla 1K calls for a liquid cooling which will be
The controller’s temperature will also be monitored by the
PLC and so warning can be sent to the driver if the controller begins to overheat.
3.4.6: Test Plan to Determine Compliance with Specifications/Interfaces
Dielectric Withstand Test
The isolation between the HV circuit and other parts of the vehicle will be tested at an rms
ac voltage equal to 1000 V plus 1.5 times the maximum expected peak voltage in the HV
circuit. The primary test will be between the HV system and the frame (which must be
connected to the ground of any low-voltage systems). Additional tests will be conducted
between the HV system and any other ungrounded conductive surfaces or objects, unless
they are protected from human contact. If any section of circuitry is completely isolated by
contactors (e.g., by having both dual contactors on the positive and negative terminals of
a battery bank), at least one contactor must be energized or jumpered during this test
such that the full HV system is energized during the test. A current of more than 4 mA will
constitute failure.
Leakage Test
For testing, a 1000 Ω resistor will be connected between points on the HV circuit and the
grounded frame. A current of greater than 1 mA through the 1000 Ω resistor will be
considered excessive.
Battery Testing
We will test the individual battery cells power output by demanding various and continuous
loads.
To test that the individual Lithium-ion cells charge evenly in the battery pack
configuration, we will measure the voltage and current levels using the PLC.
This test
will also be used for the battery pack temperature. As specified by the manufacturing
company, we will test our battery pack system for the optimal charging routine.
Motor/Motor Controller Testing
To test the configuration of the controller, we will supply 1000 amps and an excess of
144V to ensure that the Controller limits the voltage to a range of 72-144V and a max
current of 600A. Through the PLC, we will run the motor in two scenarios, race scenario
of high rpms and a distance scenario of nearly continuous rpms and monitor the
temperature of the motor to ensure that it does not overheat and does not cause
malfunction of other system components.
Emergency Switches
Because the contactors are essential to the safety of the system, we will first isolate the
components themselves to test for functionality. Two power sources will supply power to
the contactor leads and the field. To test the emergency switches with the user interface,
we will set up a system batteries, When the driver hits the E-Shutoff Button, the (register)
of the PLC will turn on and supply a voltage to the E-shutoff contactor, which shall then
turn on the main contactor.
PLC
Testing for the PLC will consist of testing the input and output logic of each module. Push
buttons and a voltage potentiometer will be used as discrete and analog inputs
respectively. Outputs will be measure using a simple circuit with a light.
Wireless Tests
The wireless system requires an input voltage of 6-9 V @ 300mA or higher. We will be
testing it for functionality only on those pins sets which we will be implementing for our
application of the system. Pins PA0-7 will be implemented as general purpose
Input/Output pins. We will also be using the serial I/O (RS-232 level) pins located in
the J1 set. Vin (power input), GND (ground, power return signal), Reset, VREF (noise
reduction), VSSA (analog ground). This testing will be accomplished by applying the
necessary power across the board and first checking pins PA0-2 (initially LED pins) for
correct power flow.
We will then test the serial input by constructing an assembly
command in NMI Terminal on a laptop and sending it across the XBEE Doggle to the
onboard radio controller (Sin on the module).
controller, PLC) in the same method.
We will similarly be testing Sout (to
3
3..5
5:: D
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Errggoonnoom
miiccss
3.5.1: Acceleration Pedal
3.5.1.1: Engineering Specifications
The main design objective is to design an acceleration pedal around a given potentiometer
(pot). The design has to incorporate the pot and protect it from exceeding its operational
limits. Also for the ergonomics of the driver, it can not interfere with the driver’s foot
movement and imitate a conventional acceleration pedal with regards to resistance to foot
and travel distance. It also has to fit within the dimensions of the vehicle and be
mountable within the vehicle chassis.
3.5.1.2: Design History
During the design process of designing the mount and foot pedal the design took drastic
turns within the process. First pot used was a right-handed one but due to the fact that
the mounting plate was on the right hand side and would interfere with the movement of
the driver’s foot from the acceleration pedal to the brake pedal.
After switching to a left handed pot the spring that moves the foot pedal into its initial
position had to be chosen. Initially the choice was a torsional spring applied directly at the
pivot point of the foot pedal. Instead the choice was made to use a extension spring
attached to the foot pedal and the base plate. This system is easier and a larger variety of
springs with the same extension and spring constant.
Lastly the attachment that comes in direct contact with the foot was changed from a bar
sticking out of the right hand side of the pedal to a plate attached on the top of the pedal.
This way the force exerted on the pedal is more central and doesn’t cause as much
torsion as before. The final product can be seen in figure 45.
Figure 45 Acceleration Pedal
3.5.1.3: Engineering Analysis
The pot has given dimensions such as mounting holes and maximum travel distance of the
lever. The lever moves a max of 45 degrees in total, 22.5 in each direction starting from
the vertical upright.
Figure 46. Potentiometer with all given dimensions [8]
Due to the limitations in dimensions given most of the pot’s dimensions are estimated. The
link connecting the lever and the foot pedal therefore has to be positioned at a certain
location along the foot pedal and to determine that location a function approach has to be
taken.
t = distance pivot point of foot pedal – attachment point of link on foot pedal
s = distance pivot point of lever – attachment point of link on lever
α = max travel angle of lever (here 45 deg)
γ = max travel angle of foot pedal (here 30 deg)
x = travel distance of both lever and pedal at given distances s and t
A simple relationship calculation will give us the wanted function:
x 2 * * t
 * t * 30

x
30
360
180
x 2 * * s
 * s * 45

x
45
360
180
 * t * 30  * s * 45

180
180
s * 45  t * 30
45
3
t
s s
30
2
t
s
r
d
Figure 47. Pedal assembly from side with dimensions
To calculate the dimensions needed for the spring the dimensions of the pot are not
needed. The only thing that might be of importance in this case is if the pot has its own
spring and is pushing back on the pedal. The force needed to push back the pedal is
estimated at 5 lbf. The max force that was estimated will be applied to the foot pedal is
50 lbf. All these forces are estimated based upon experimental trials on previous cars of
the same size and classification.
The spring force is given by
k = Spring constant
x = length the spring extends by
Fs  k * (Vex  V
The distance the spring travels is given by
r = distance pivot point of pedal – attachment point of spring on pedal
d = distance pivot point of pedal – attachment point of spring on base plate
Assuming we want to keep the attachment point of the spring on the pedal a variable to
give us a larger range in springs we can use the following describes the position using
cosine law. The angle the pedal makes with regards to the base plate is independent to
the position of the spring attachment along the pedal and can be found by using
measuring tools within Pro/E. Using the law of cosine we can project the spring constant
as a function of the distance of the spring attachment.
Vex= Extended length of entire spring
V= free length of spring
β1,2 =angles between pedal and base plate
V 2  r 2  d 2  2rd * cos(  1 )
Vex  r 2  d 2  2rd * cos(  2 )
2
r ?
d  1.467
Vex  r 2  2r * 1.467 * cos(101.303)  1.467 2  r 2  0.5r  2.15
V  r 2  2r * 1.467 * cos(67.1884)  1.467 2  r 2  1.137r  2.15
Fs  k * (Vex  V )
Using moment arms we can establish the force the spring has to exert on the pedal to
counter the 5 lbf of the foot.
The distance of the pedal to the pivot point is given by 6.10 in
F foot  5lbf
F foot * 6.1in  Fs * r
30.5
 Fs
r
Fs
k
(Vex  V )
For reasons of simplifications if the distance r is chosen to be 1.5in temporarily then the
entire calculations break down to the spring constant being k= 31.1574lbf*in, although
other combinations are possible. We could also determine if a given spring constant will
evolve in a suitable solution.
3.5.1.4: Material Study
The materials chosen for this project are to accompany the design in its simplicity. The
materials must have easily manufacturability and be readily available on the market and
also inexpensive. The main material that will suit this need is aluminium. It is light
weighted, strong, easy to machine, cheap and available within short time and distance.
The bolts and the pin part used are made of steel for easier machining and availability
from local hardware stores.
3.5.2: Steering Wheel
3.5.2.1: Engineering Specifications
The steering wheel for this vehicle needs to meet with several specifications set by the
group. The steering wheel needs to be easily removable so that the driver is able to enter
and exit the car quickly. Also, the steering wheel needs to be designed to hold the touch
screen that will be installed to monitor the car. This touch pad is going to be mounted
below the handles but above the steering column. The steering wheel needs to be able to
turn a full rotation without hitting the driver, because otherwise the driver will not be able
to complete all of the maneuvers that the vehicle will be capable of.
(a
(b
3.5.2.2: Design History
)
)
Steering wheel underwent a couple iterations of design review to best determine driver
ergonomics and positioning. Initially, a standard 10” round steering wheel was determined
along with an instrument panel for the electronics. As the research and development of the
vehicle continued, a different approach to the steering wheel was considered. Figure 48
shows how the touch screen of the PLC is attached directly to the steering wheel. On the
back, a quick release switch, as indicated by Formula SAE rules, is seen in figure 41b.
Figure 48. (a) Steering Wheel with touch screen, (b) rear view showing quick release
The latest revision incorporates the touch screen with a barrier around the edges to the
wheel. The touch screen will not be polarized; therefore sunlight will need to be blocked
from the screen as best as possible for the driver to clearly view. This rendition is shown
in figure 49.
Figure 49. (a) Steering Wheel with touch screen, (b) rear view showing quick release
3.5.2.3: Material Study
The material selected for the steering wheel is aluminum 6064. This was selected
because the material needs to be light weight and durable. This is especially necessary for
the steering wheel because it will be removed from the car most of the time, which means
that a heavy piece will not only be detrimental to the weight of the car but also will be
difficult to handle for the people carrying it.
Also, while the steering wheel is removed
from the car it stands a greater chance of being abused.
These reasons along with
material availability lead to our choice of aluminum.
(a
3.5.3: Driver’s Seat
)
3.5.3.1: Engineering Specifications
(b
)
A driver seat has to accommodate for the driver comfort and ergonomics. It also needs to
lightweight to keep down weight of the vehicle.
3.5.3.4: Material Study
We will purchase a lightweight seat to integrate into the vehicle. The Tillet T11 seat shown
in figure 50, only weighs 3.5 pounds and can fit the dimensions of the driver’s cockpit.
Other models of the T11 seat are shown but cost raise with the padding or flexibility of the
seat material. This data is shown in Table 3.9.
Table 3.9: Tillet Seat Variations [13]
Model
Weight
Cost
T11 1/4 pad
4.0 lb
$239.00
T11 no pad
3.5 lb
$138.00
T11VG
2.5 lb
$170.00
Figure 50. Tillet Racing Seat no pad, Item #T11 [13]
With the prices and weights of the various lightweight racing seats, the standard Tillet T11
with no pad is the best option for our application.
3.5.4: Safety Equipment
3.5.4.1: Engineering Specifications
For safety requirements, the driver must comply with the safety guidelines of Formula
Hybrid [16] and Formula SAE rules [15]. The driver is required to have a helmet, fire
suit, gloves, goggles or face shields, and shoes as required by these rules. Specifications
for each are as follows:
Helmet
- Snell M2000, SA2000, M2005, K2005, SA2005
- SFI 31.2A, SFI 31.1/2005
- FIA 8860-2204
- British Standards Institution BS 6658-85 types A or A/FR rating
Fire Suit
- SFI 3-2A/1 (or higher)
- FIA Standard 8856-1986
- FIA Standard 8856-2000
Fire resistant gloves with no holes, no leather gloves
Goggles or face shields made of impact resistant materials
Shoes of durable fire resistant material which have no holes
Also required by safety rules is the safety harness for the driver. This harness is a 5-point
harness made of Nylon or Dacron polyester. These harnesses are typically found at stores
selling racing components.
3.5.4.4: Material Study
Figure 51 shows a layout of the driver apparel that is required to wear. Prices for these
required safety items are accounted for in our budget as miscellaneous expenses. For a
complete set of safety apparel, the cost is approximately $500.
Figure 51. Driver Racing Apparel [14]
To meet safety requirements from Formula Hybrid and NEDRA, a
5-point harness is implemented into the design. 5-Point harnesses
made of Nylon or Dacron polyester are widely available. They
generally range in price from $100 to $200. Figure 52 shows a
typical harness used for this application.
Figure 52. RJS 5-Point Harness [14]
4
4:: B
Bu
ud
dggeett
4
4..11:: I
Inniittiiaall B
Buuddggeett
This budget is an initial estimation to construct a running electric vehicle. If time and money permits, the budget will turn to the final budget.
This breakdown allows us to develop an electric vehicle, and then improve the design and performance of the vehicle. This budget also
includes a donation of controller and tires and a donation of half of the batteries (250 cells).
Item # Part
1
2
3
4
5
6
7
8
9
10
11
12
13
14
15
16
17
18
19
20
21
22
23
24
25
26
27
28
29
30
31
32
33
34
35
36
37
38
39
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Welding Filler Rod
Differential
Motor Mounts
Differential Mounts
Controller Mounts
Shielding
Seat
Fiberglass Body
Fiberglass Resin
Brake Lines
Brake Fluid
Tires
DC Motor
Controller
Throttle Control
Speed Sensor
Fuse
PLC
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
Touch Screen Display
Contactor
Wire
Wireless Device
Li-ion Batteries (1 set)
Battery Management
Misc Electrical Comp
Misc Hardware
Misc Expenses
Manufacturer
Budget for Racing Electric Vehicle:
Description
Part Number
41-1-049
Chassis Shop
41-1-065
Chassis Shop
41-1-095
Chassis Shop
41-1-1-065
Chassis Shop
41-58-058
Chassis Shop
C73-002
Chassis Shop
Kawasaki
ALRO
ALRO
ALRO
ALRO
T11
Tillet
Fiberlay, Inc.
Fiberglass Florida
SUM-220136
Summitt
950-290-0632
Jegs
20.0x6.5-13
Goodyear
00-08219
NetGain Technologies, LLC
Zilla Z1K-LV
Café Electric
PMC #PB6
Curtis
480-2015-ND
Digi-Key
A30QS800-4
FERRAZ/SHAWMUT
EZPLC-D-96E
EZ Automation
EZIO-4THI
EZ Automation
EZIO-4DCIP4RLO
EZ Automation
EZIO-8ANIV
EZ Automation
EZIO-8ANIC
EZ Automation
EZIO-8HSDCI
EZ Automation
EZIO-8DCOP
EZ Automation
EZIO-HSCM2
EZ Automation
EZC-T6C-E
EZ Automation
SW200
Albright
#2/0
Prestoflex
PIC
ANR26650M1
A123 Systems
F877 & 1287
PIC
Radio Shack
Ace Hardware
-
Round Chromoly Tubing, 1" OD, .049" THK, per FT
Round Chromoly Tubing, 1" OD, .065" THK, per FT
Round Chromoly Tubing, 1" OD, .095" THK, per FT
Square Chromoly Tubing, 1" OD, .065" THK, per FT
Round Chromoly Tubing, 5/8" OD, .058" THK, per FT
#65 Filler Rod, 1/16"x36", per LB
Kawasaki Bruteforce Front Diff, 4.375:1 ratio
Aluminum, 1/4" THK per SHT
Aluminum, 1/2" THK per SHT
Aluminum, 1/8" THK per SHT
Aluminum, 1/16" THK per SHT
Seat, Large
Fiberglass matting 3.2 oz.
Epoxy Resin Kit (3 Gallon Size)
3/16" Steel Hard Lines, 25 ft
570-Brake Fluid, 12-ounce Can
Tires, Slicks, for 13" rims, 6.5 wide, D1385, R065
32.3HP continuous series wound DC motor
72-156VDC series wound controller, 1000A max. w/ HardBall
Swinging arm throttle input, 5k ohms
Hall Effect Sensor
Up to 800A systems
12 Slot EZPLC Base (96I/O Max)
4 Thermocouple Input Module
4 DC In, 4 DC Out Relay Module
8 Analog Input (voltage) Module
8 Analog Input (current) Module
8 DC High Speed Input Module
8 DC Output (source) Module
High Speed Counter Module
5.7 viewable Touch Screen LCD display
400A continuous 12V contactor
00 gauge (Black) 33 feet
Serial Wireless Adapter
Depends on Differential Ratio. 44S10P with 1 set
Voltage, Current, and Temperature Measurements
Electrical Stuff (wire, fuses, etc)
Hardware (nuts, bolts, etc)
Misc Expenses (Registration fees, shirts, cards, etc)
QTY
Retail Price
$3.24
$2.52
$4.68
$6.96
$2.16
$4.99
$500.00
$60.00
$60.00
$60.00
$60.00
$179.00
$6.98
$45.81
$19.95
$7.49
Donated
$1,450.00
Donated
$75.00
$7.40
$42.00
$289.00
$139.00
$39.00
$99.00
$99.00
$24.00
$19.00
$99.00
$719.00
$119.99
$99.00
$161.00
$18.00
$100.00
$200.00
$200.00
$500.00
15
15
20
8
25
5
1
1
1
1
1
1
7
1
1
2
4
1
1
1
1
2
1
4
1
1
1
2
2
1
1
1
1
2
250
20
1
1
1
Total Cost:
Total Price
$48.60
$37.80
$93.60
$55.68
$54.00
$24.95
$500.00
$60.00
$60.00
$60.00
$60.00
$179.00
$48.86
$45.81
$19.95
$14.98
$0.00
$1,450.00
$0.00
$75.00
$7.40
$84.00
$289.00
$556.00
$39.00
$99.00
$99.00
$48.00
$38.00
$99.00
$719.00
$119.99
$99.00
$322.00
$4,500.00
$2,000.00
$200.00
$200.00
$500.00
$12,906.62
4
4..2
2:: F
Fiinnaall B
Buuddggeett
This is the overall budget. If time and money allow, all components will be implemented in the design to construct a better performing, more
enhanced electric vehicle.
Item # Part
1
2
3
4
5
6
7
8
9
10
11
15
17
18
19
20
21
22
23
24
25
26
27
28
29
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
44
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Chromoly Tubing
Welding Filler Rod
Differential
Motor Mounts
Differential Mounts
Controller Mounts
Shielding
Seat
Fiberglass Body
Fiberglass Resin
Brake Lines
Brake Fluid
Tires
DC Motor
Controller
Throttle Control
Speed Sensor
Fuse
Fuse
PLC
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
PLC I/O
Touch Screen Display
Contactor
Wire
Wireless Device
Wireless Device
Li-ion Batteries (2 sets)
Charger
Misc Electrical Comp
Misc Hardware
Misc Expenses
Manufacturer
Chassis Shop
Chassis Shop
Chassis Shop
Chassis Shop
Chassis Shop
Chassis Shop
Kawasaki
ALRO
ALRO
ALRO
ALRO
Tillet
Fiberlay, Inc.
Fiberglass Florida
Summitt
Jegs
Goodyear
NetGain Technologies
Café Electric
Curtis
Café Electric llc
FERRAZ/SHAWMUT
FERRAZ/SHAWMUT
EZ Automation
EZ Automation
EZ Automation
EZ Automation
EZ Automation
EZ Automation
EZ Automation
EZ Automation
EZ Automation
Tyco Electronics
Prestoflex
newmicros
newmicros
A123 Systems
A123 Systems
Radio Shack
Ace Hardware
-
Budget for Racing Electric Vehicle:
Part Number
Description
41-1-049
41-1-065
41-1-095
41-1-1-065
41-58-058
C73-002
T11
SUM-220136
950-290-0632
20.0x6.5-13
00-08219
Zilla Z1K-LV
PMC #PB6
2171S
A30QS600-4
A30QS800-4
EZPLC-D-96E
EZIO-4THI
EZIO-4DCIP4RLO
EZIO-8ANIV
EZIO-8ANIC
EZIO-8HSDCI
EZIO-8DCOP
EZIO-HSCM2
EV500
#2/0
Xbee PlugaPodS
Xbee Dongle
ANR26650M1
-
Round Chromoly Tubing, 1" OD, .049" THK, per FT
Round Chromoly Tubing, 1" OD, .065" THK, per FT
Round Chromoly Tubing, 1" OD, .095" THK, per FT
Square Chromoly Tubing, 1" OD, .065" THK, per FT
Round Chromoly Tubing, 5/8" OD, .058" THK, per FT
#65 Filler Rod, 1/16"x36", per LB
Kawasaki Bruteforce Front Diff, 4.375:1 ratio
Aluminum, 1/4" THK per SHT
Aluminum, 1/2" THK per SHT
Aluminum, 1/8" THK per SHT
Aluminum, 1/16" THK per SHT
Seat, Large
Fiberglass matting 3.2 oz.
Epoxy Resin Kit (3 Gallon Size)
3/16" Steel Hard Lines, 25 ft
570-Brake Fluid, 12-ounce Can
Tires, Slicks, for 13" rims, 6.5 wide, D1385, R065
32.3 HP continuous series wound DC motor
72-156VDC series wound controller, 1000A max. w/ HardBall
Swinging arm throttle input, 5k ohms
Advanced DC Motor Speed Sensor
Up to 600A systems
Up to 800A systems
12 Slot EZPLC Base (96I/O Max)
4 Thermocouple Input Module
4 DC In, 4 DC Out Relay Module
8 Analog Input (voltage) Module
8 Analog Input (current) Module
8 DC High Speed Input Module
8 DC Output (source) Module
High Speed Counter Module
5.7 viewable Touch Screen LCD display (outdoors)
Kilovac 600A continuous 12V contactor
00 gauge (Black) 33 feet
Connects 2 serial ports
Connest to the user interface
Depends on Differential Ratio. 44S10P with 2 sets
110 or 220 system charger
Electrical Stuff (wire, fuses, etc)
Hardware (nuts, bolts, etc)
Misc Expenses (Registration fees, shirts, business cards, etc)
Retail Price
QTY
$3.24
$2.52
$4.68
$6.96
$2.16
$4.99
$500.00
$60.00
$60.00
$60.00
$60.00
$179.00
$6.98
$45.81
$19.95
$7.49
$119.00
$1,600.00
$2,950.00
$75.00
$42.50
$54.50
$42.00
$289.00
$139.00
$39.00
$99.00
$99.00
$24.00
$19.00
$99.00
$2,500.00
$931.00
$99.00
$161.00
$95.00
$18.00
$2,800.00
$300.00
$250.00
$800.00
15
15
20
8
25
5
1
1
1
1
1
1
7
1
1
2
4
1
1
1
1
2
2
1
4
1
1
1
2
2
1
1
1
1
2
2
1000
1
1
1
1
Total Cost:
Total Price
$48.60
$37.80
$93.60
$55.68
$54.00
$24.95
$500.00
$60.00
$60.00
$60.00
$60.00
$179.00
$48.86
$45.81
$19.95
$14.98
$476.00
1,450.00
2,950.00
75.00
42.50
109.00
84.00
289.00
556.00
39.00
99.00
99.00
48.00
38.00
99.00
2,500.00
931.00
99.00
322.00
190.00
18,000.00
2,800.00
300.00
250.00
800.00
$34,008.73
5
5:: O
Orrggaan
niizzaattiioon
n aan
nd
dC
Caap
paabbiilliittiieess
Team Member
Discipline
Title
Elizabeth Diaz
Mechanical Engineering
Team Lead
Valerie Bastien
Electrical Engineering
Sensors Lead
Jared Doescher
Mechanical Engineering
Thermal Effects Analyst
Kristine Harrell
Electrical Engineering
Power Systems Lead
Jason McSwain
Computer Engineering
Communications Lead
Jason Miner
Mechanical Engineering
Mechanical Lead
Audrey Moyers
Electrical Engineering
Kathleen Murray
Aerospace/Mechanical Engineering
Aerodynamics Specialist
AJ Nick
Mechanical Engineering
Mechanical Designer
Matthew Reedy
Electrical/Computer Engineering
Electrical Lead
Joshua Wales
Mechanical Engineering
David Wickers
Mechanical Engineering
Frame Analyst
Oliver Zimmerman
Mechanical Engineering
Mechanical Designer
Programmable Logic
Controller Lead
Systems Integration,
Drive System Lead
2007 R.E.V. TEAM STRUCTURE
Team Lead
Development Group
Procurement Group
Manufacturing Group
Integration Team
Design Teams
Chassis & Body
Chassis Redesign
Body Redesign
Mounting Points
Aeros/Ground
Effects
Vehicle Dynamics
Suspension
System
Steering System
Braking System
Driver Interface & Ergonomics
Cockpit Design
Safety Equipment
Driver Interface
Drive System
Motor
Drivetrain
Control System
Battery System
Cooling System
Shielding System
Electrical
Battery Management
Instrumentation
Data Transfer System
Power Management
6
6:: S
Scch
heed
du
ulliin
ngg
6
6..11:: G
Gaanntttt C
Ch
haarrtt
6.1.1: Mechanical Task Schedule
6.1.2: Electrical Task Schedule
6
6..2
2:: M
Miilleessttoonneess aanndd D
Deeaaddlliinneess
March 30, 2006 –
Sponsorship Package complete
Motor/Controller determined
April 28, 2006 –
Team Initial Proposal Complete
May 15, 2006 –
Finish Research (include pricing)
Finish Frame Design Concept
July 15, 2006 –
Finish Suspension Layout
Organize Electrical COTS Parts
September 14, 2006 –
Finalize Preliminary Vehicle Design
Begin ordering Major Components
October 23, 2006 –
PDR
November 1, 2006 –
Finish Analysis
November 11, 2006 –
Finish Written PDR
January 19, 2007 –
Complete Mechanical Build
January 20-21, 2006 –
Battery Beach Burnout, NEDRA Event
March 1, 2007 –
Complete Vehicle Build
March 30, 2007 –
Finish Optimization and Testing
April 2, 2007 –
Present Completed Car
May 1-3, 2007 –
Formula Hybrid Competition
77:: A
Ap
pp
peen
nd
diixx
77..11:: C
Caallccuullaattiioonnss
To calculate the top speed, we factor in the top rpm the motor can handle and the gear
ratio.
TopSpeed  6500
rev 60 min
1 foot 1mile
1
*
* 20.6in *  *
*
*
 84.Mph
min
hr
12in 5280 ft 4.375
The following calculations are used to find the distance the tires will take to reach its
optimum driving temperature assuming full slippage.
Using the Energy Equation:
E 1 Q2 1 W2
where :
E  U  mCv T
Assuming 1Q2  0 due to neglible external heat flow
Rearrangin g :
mCv T  1W2
where :
 J 
Cv  Constant Volme, Specific Heat  K 
 kg 
ΔT  (Tfinal tire temp  Ttire amb ) ( K )
m  mass of the tire (kg)
W  F*d
where :
d  the distance traveled (m)
F  force of friction
Rearrangin g and solving for d :
m Cv(Tdesired  Tamb )
d  tire
mcar g
Sample Calculation:
Tires:
Mass of 7.5” wide tire: 13 lb = 5.896 7 kg
Mass of 6.5” wide tire: 9 lb = 4.082 3 kg
Temperature:
Ambient Tire Temperature: 25ْ Celsius = 298.15 Kelvin
Desired Final Tire Temperature: 70ْ
Celsius = 343.15 Kelvin
Assumptions:
Full sliding contact
Overall Car weight plus driver: 165 lb = 74.842 7 kg
Constant Specific heat is equal to rubber (Cv): 1600 J/kg-K
Assuming a coefficient of friction: 1.7
Final Calculation:
7.5" Tire :
(5.896 7 kg)(1600 J/kg - K )(343.15 K  298.15 Kelvin )
d
 340m  0.21miles  1115.98 ft
1.7(74.842 7 kg)(9.81 m/sec 2 )
6.5" Tire :
(4.082 3 kg)(1600 J/kg - K )(343.15 K  298.15 Kelvin )
d
 240m  0.15miles  772.54 ft
1.7(74.842 7 kg)(9.81 m/sec 2 )
The previous calculations show the 6.5” wide tires are prefect for our application.
Brake Force Calculations
Brake Pedal:
Assume that the driver input force is 90 lb.
90 lb
4 in
360 lb
Moment output from pedal:
Moment  (InputForce )( Distance)
(90)  (4)  360 lb  in
The master cylinder:
You can adjust pressure output of each master cylinder by increasing or decreasing length
of the piston push rod in the master cylinder. This is allows for an adjustable rear and
front braking force. To account for this difference in the front and rear braking a percent
is applied to the pressure calculation.
F
P   ( PercentBra king )
 A
Where:
A
D 2
4
D: the master cylinder diameter
F: the force from the brake pedal
P: the pressure from the mater cylinder
2
3
4
A     0.442 in 2
4
Front :
 
 360 
P
  0.60  487.58 psi
 0.442 
Rear :
 360 
P
  0.40  325.79 psi
 0.442 
The caliper:
The calipers have two pistons that actuate the brake pads so the force is multiplied by 2.
F  2( P)( A)
Where:
A
D 2
4
P: the pressure from the mater cylinder
D: the diameter of the caliper
FCaliper Force: the clamp load
A: area of the caliper
Front Calipers:
A
 (.84) 2
4
 .554 in 2
FCaliperForce  2  487 .58  0.554  540 .24 lb
Rear Calipers:
A
 (1.15) 2
4
 1.04 in 2
FCaliperForce  2  325.79  1.04  677.64 lb
The brake pads:
There are two brake pads so the force is multiplied by a factor of two.
Rotor Force  2  (Caliper Force)  (  )
Where:
 = coefficient of friction = 0.45 (good assumption for most race cars)
Front:
F  2  540.24  0.45  486.26 lb
Rear:
F  2  677.64  0.45  606.88 lb
The rotor:
The torque applied on the rotor acts on both side so the torque is multiplied by 2.
Torque  (2)(Rotor Force)( d )
Where:
d: The distance between the center of the rotation and the force to act at a point midway
across the rotor face.
Front:
T  2  486.22  5  4862.2 lb  in
Rear:
T  2  606.88  3.5  4248.16 lb  in
The wheels and tires:
F
Torque
r
Where:
F: Force generated between the tires and road
r: Rolling radius of tire
Front:
F
4862.2
 486.22 lb
10
Rear:
4248.16
F
 424.82 lb
10
Acceleration calculation:
2( FFront wheel )  2( FRear Wheel )
a
W
Where:
a: Lateral deceleration
F: Force generated between the tires and the road for the front and rear tires. Force is
multiplied by a factor of 2 because there are 2 front and 2 rear tires.
W= Total estimated weight of the car, which includes car and driver.
a
2(486.22)  2(424.82)
 2.80 g
650
Stopping distance:
2
V
D i
2a
Where:
Si: the initial speed
a: Lateral deceleration
Vi 
80 mile
1 hr
5280 ft
X
X
 117.3 ft / s
hr
3600 s
1 mile
a  2.80 g X
D
32.14
1g
ft
s 2  89.99 ft / s 2
117.32
 76.45 ft
2(89.99)
Calculations based on Equations from Wilwood Engineering – ( www.wilwood.com)
77..2
2;; R
Reeffeerreenncceess
(1) “The History of Electric Vehicles.” 2006. New York Times Company. 20 Apr 2006.
http://inventors.about.com/library/weekly/aacarselectrica.htm
(2) General Information of Metals, http://www.suppliersonline.com/
(3) MatWeb, http://www.matweb.com/index.asp?ckck=1
(4) Chassis Shop, http://www.chassisshop.com
(5) American Society of Nondestructive Testing, www.asnt.org/ndt/primer3.htm
(6) Hoosier Tires, http://www.hoosiertire.com/Fsaeinfo.htm
(7) Goodyear Tires, http://www.racegoodyear.com/sae.html
(8) Café Electric, http://www.cafeelectric.com/
(9) New Micros, Inc., http://www.newmicros.com
(10) Omega Thermocouples, http://www.omega.com/prodinfo/thermocouples.html
(11) NetGain Technologies, LLC, http://www.go-ev.com/motors-warp.html
(12) EZAutomation PLC, http://www.ezautomation.net
(13) Tillet Race Seats, http://www.tillett.co.uk/estore/shop/kartSeats.asp?seat=T5
(14) Thunder Racing Apparel, http://thunderracing.com/
(15) Formula SAE Competition, http://students.sae.org/competitions/formulaseries/
(16) Formula Hybrid Competition, http://www.formula-hybrid.org
(17) Fundamentals of Heat and Mass Transfer, 5th Edition, By Incropera and DeWitt
(18) A123 Systems, http://www.a123systems.com
(19) Cornell Stress Analysis Paper,
(20)Aurora Bearing Company, http://www.aurorabearing.com/
(21) Keizer Aluminum Wheels Inc., http://www.keizerwheels.com/
(22)Kawasaki Motorcycles, http://www.kawasaki.com