1:1: IntroductionIntroductionREV, 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 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.

1.1: Purpose1.1: PurposeThe 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

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they seek to further school spirit, Florida Tech’s relations with the community, and public interest in electric vehicle technology.

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1.2: Team Goals1.2: Team GoalsThe goals of the REV project are:

To design and build an electric vehicle for Autocross style racing that would be capable of being competitive with other electric vehicles in NEDRA (National Electric Drag Race Association) 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.

1.3: Background1.3: BackgroundIn 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 vehicles. During the 1990’s there was another push for electric vehicles because of new regulations in pollution 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

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Figure 1. Historical EV

Figure 2. EV Pick-Up Truck

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.

2: Design 2: Design ObjectivesObjectives 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 secondsTop speed of 85 mphMaximum power available between 20 and 40 mph. Lightweight (under 650lb with driver)15 minute battery life running at high performance speeds

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3: Design3: DesignWe 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 with the same motor. 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.

3.1: Chassis and Body3.1: Chassis and Body3.1.1: Frame3.1.1.1: Engineering Specifications

Table 3.1: Engineering SpecificationsLabel ValueLength Minimum (front-to-rear wheels) 60 inch

Main Hoop Angle 10 º from vertical

Driver Head Clearance 2.00 inchVehicle 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 psiMaximum Stress on Weld Points 26,600 psiMaximum 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

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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 reference [19]. If the frame is too flexible (i.e. below 1600 ft-lbs/deg), the vehicle will not handle well.

Chassis also needs to follow some basic specifications. Because we will reuse body to the 2001-2002 Formula SAE car, the chassis will need to fit within that body envelope. The chassis will need to accommodate for a 95 th

percentile person, as defined by Formula SAE rules [15]. The chassis will also need to be designed and sized to support a motor and differential in the rear. The side pods need to follow certain envelope restrictions too. Side pods need to be designed as to not interfere with driver and reasonable in size. The width should not extend past the mid-plane of the front wheels. The side pods should also be capable of supporting the size and weight of the battery packs.

3.1.1.2: Design HistoryThe 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. The first design concept is shown in Figure 3. This concept represented what size we had for the vehicle, but would not adequately support the different weights. This design would not be support the weight of the heavy components and the rear is not sized to support a motor, differential and a rear suspension setup.

Figure 3. First Conceptual Chassis Design

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The 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. This idea was eliminated because the complexity of a dual-motor setup. A single motor inline with a geared differential became our final setup. For this idea, we attempted to modify the rear section of the frame shown in Figure 4b for a differential mounting box. This design was changed due to manufacturability.

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. Also this design will be easier to manufacture because it only involves welding steel tubing.

Figure 5. (a) Side view, (b) Top view

Now that the chassis has been developed, the side pods have to attach following the set specifications. The first concept of the side pod

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(a)

(a)

(b)

(b)

Open to air flow

Front

Release air

(Figure 6) was a general layout of columns of batteries and side pod layout. The idea was to have the side pod open to the outside to allow air to pass through the pod over staggered columns of batteries to cool the batteries. This design was modified because the battery cooling will involve freezing packs of batteries rather than air cooling in the side pods. This general design was enhanced in later concepts.

Figure 6. Top view of side pod (First concept)

The second side pod conceptual design is shown in figure 7. 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 side pod shown in Figure 7 is mirrored to the other side. It fits with in the envelope and adequately supports the battery packs. The side pod is built to be oversized form the battery packs to accommodate any changes or additions to the battery packs.

Figure 7: Top view of side pod

The final conceptual design of the chassis includes the side pods and some small changes from the previous layout. The variation of placement of the square tubing for the differential, bending of roll hoops for ease of manufacturability, and measurement changes to support the components are some changes made to this final chassis layout. Also, the chassis was evaluated and modified to incorporate the rules set by Formula SAE [15]. This latest version can be seen in Figure 8.

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Figure 8. Final Chassis Concept

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 and lightening the frame. Figure 9 shows a layout of the loads and constraints applied to the chassis for analysis. Table 3.2 lists the optimization iterations the frame underwent from finite element analysis.

Figure 9. Constraints and Loads applied for finite element analysis

Table 3.2: Optimization IterationsIteratio

nDeflecti

on(inch)

Stresses

(psi)Cause of Change

1 .0068 8493 No change, conceptual chassis2 .0054 4341 Add cross member in chassis, see Figure 93 .0070 5706 Some member wall thicknesses changed

to .049”4 .0020 8934 Driver side support changed angle, see Figure

105 .0055 4348 Suspension perches wall thickness changed to

.095”6 .0054 4342 Suspension perches wall thickness changed to

.049”

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(a) Isometric

(b) Side (c) Front

For the side pod, the cross member was added to decrease the overall deflection (see figure 10). This added support also stiffened the frame, increasing the torsional rigidity.

Figure 10. Side Pod cross members, iteration 2

To optimize deflections and stresses, the driver side support angle was changed as shown in figure 11. The increase in stresses was too great to justify the decrease in deflection.

Figure 11. (a) Driver side support, iteration 4, (b) Final design choice

3.1.1.3: Engineering AnalysisAll 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. These components include the motor, differential, driver, controller, and batteries. The frame is constrained at the attachment points of the suspension to accurately represent the actual model. As shown in Figure 12, the frame holds and maximum load is placed at the back of the frame near the motor.

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(a) (b)

(a)

(b)Figure 12. ANSYS Static Loading, (a) Loads and Constraints, (b) Stress Results

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Figure 13. ADAMS Model of Vehicle

Figure 13 shows the ADAMS model that was used for dynamic analysis. In ADAMS we used braking, cornering and lane change analysis. In doing so the forces at each A-Arm attachment point were acquired. These forces, seen in figures 14 and 15, were then applied to the ANSYS model at the attachment point locations to see the stress that occurred in the frame. Because we are evaluating the stresses applied at the A-Arms we need to re-constrain the frame. For this we will constrain the center of the frame at the front and rear of the chassis. With this constraint, the static loading is still taken into account in the dynamic loading. For braking, the analysis was done for a 50mph to 0mph 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 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.

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Figure 14. Loads applied on front A-Arm geometry

Figure 15. 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, the frame will be able to handle loads taken from a lane change.

Table 3.3: Forces of Dynamic Loads Cornering Braking Lane change

point: Magnitude(lbf) Magnitude(lbf) Magnitude(lbf)1 39.31908413 -220.762382 44.065311242 34.84538617 272.4684389 47.822223523 -20.45761381 -64.2953577 -38.737001124 66.09382924 245.9409836 81.595055545 -70.81481705 216.7158211 -102.81439836 89.9235772 -275.3909552 -60.852191057 -100.7144065 -62.7216951 52.193805688 184.3433333 -243.2432763 152.52797329 -41.58965446 -135.7846016 -67.56172658

10 -38.89194714 98.01669915 83.1005885311 -37.99271137 -87.00106094 22.2182330412 -10.56602032 80.48160159 19.7424495113 -51.03163006 137.8078821 -19.3988104614 -59.5743699 -98.69112598 46.41746715 56.42704469 85.65220728 -31.8499612216 66.99306501 -80.93121948 25.4444498

stress(psi) 15120 15506 NA

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The expected max Von mises stress that we experienced is 15506 psi with a factor of safety of 3, which is below our engineering specification of 21300 psi.

To prove that ADAMS analysis is correct, hand calculations for the lateral acceleration and lateral tire forces are calculated and then compared to the ADAMS output.

Where a is angular acceleration, v is velocity, r is the radius of turn, F is the force, and m is the mass of car.

The data output from ADAMS:a= F= 1555 N

Percentage error between ADAMS analysis and hand calculations:

The error between ADAMS and the hand calculations is due to the damping and the springs added in ADAMS analysis. Therefore, the error from ADAMS analysis can be considered negligible.

Figure 16. Torsional Rigidity, (a) Loads and Constraints, (b) Stress Results

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(b)(a)

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 16). 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 each node encountered, the torsional rigidity is calculated as:

Where k is the torsional rigidity, L is half the length of the hoop, y1 and y2 are the displacements of the node, and F is the force applied. Based on other studies, 100 pounds of force is a reasonable amount of force to determine a rigid structure.

The welds at each intersection are another point that must be analyzed. We considered the weld as 1/8” fillet at the point of intersection. The weld material, type S70-2, will be placed as 1/8” bead around the attachment points of the chromoly. This material has yield strength of 70,000 psi. Dynamic analysis showed the forces that each beam will hold. These forces were then applied to a small section of the chassis. Then the stresses were observed just at the area of the weld material. Figure 17 shows the analysis results of the weld analysis. According to Finite Element Analysis, the highest stress the weld material will see is 14,608 psi. From this data, the welds will be able to endure the stresses that the vehicle encounters.

Figure 17. Weld Analysis Stress Results

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3.1.1.4: Material StudyAccording 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.

Based on the rules, we chosen to study the material properties of carbon steel, alloy steel, aluminum, and carbon fiber. For the materials under consideration, we compared strengths, corrosion resistances, machinability, weldability, availability, and costs of each of the materials.

Material Strength and Corrosion Resistance:For material strength we looked at the yield strength of the material and the modulus of elasticity. Table 3.4 shows a list of strengths of the various materials.

Table 3.4: Material PropertiesMaterial Tensile Yield

StrengthModulus of Elasticity

Carbon Steel, ie AISI 1010 [3] 44200 psi 29700 ksiAlloy Steel, ie AISI 4130 [3] 64000 psi 29700 ksiAluminum, ie 6061-T6 [3] 40000 psi 10000 ksiCarbon Fiber [23] 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 normal 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].

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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. Carbon Fiber is costly and not easily accessible. We also do not have the resources to handle this type of material for space frame tubing. 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/ftAISI 4130 $3.24/ft $2.52/ft $4.68/ftAluminum 6061-T6

- $3.24/ft $3.72/ft

Although carbon fiber material is very lightweight, it is expensive and not widely available. Aluminum becomes a much stronger alloy after heat-treating but a heat treatment process 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. AISI 4130 chromoly is a readily available material that, with our resources, may be donated or discounted by local vendors. 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: Body3.1.2.1: Engineering SpecificationsThe 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

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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 HistoryThe shell of the car is designed around the frame and any externally mounted components. Figure 18 shows several artists’ renderings of original concepts of the shell of the body. 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.

Figure 18. (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 19.

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(b) (d

)

(c)

(a)

Figure 19. Final body with paint design

3.1.2.3: Material StudyFor the shell of the vehicle there were two types of material that were considered – sheet metal and fiberglass with resin. The sheet metal was considered because it would be easy to manufacture. The sheet metal is bent to the shape of the frame and then painted. Fiberglass involves more time and steps. Fiberglass first needs a form, or mold, and then several layers of fiberglass would be applied. We have the facilities and resources to work with both materials, but only one method is necessary. Fiberglass is 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.

3.2: Vehicle Dynamics3.2: Vehicle Dynamics3.2.1: Suspension and Steering Geometry3.2.1.1: IntroductionWhen designing a suspension for a high performance vehicle such as this many different parameters must be taken into consideration: 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 points of a triangle at the same time, the closer you get to one point the further you get from another. With this in mind, the design of a suspension system is made into an iterative process in which decisions are made with certain parameters in mind and check to ensure other parameters are not greatly compromised. Finally the

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(a) Single color

(b) Lightning design

designer must choose what is most important and make compromises accordingly.

3.2.1.2: Geometry AnalysisThe definition of track width is the distance of the centerlines of the tires when viewed from the front, see figure 20. 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 years’ cars, as well as other highly competitive schools vehicles, the track width is set to 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).

Figure 20. Track width

Once the track width is 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, shown in Figure 21. 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. The initial baseline value of caster in the front suspension is a positive degrees.

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Figure 21. Caster Angle

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 22). This distance affects the amount of steering force required by the driver. Small amounts of KPO are beneficial for steering feedback, but need to be kept minimal to reduce any excessive steering forces. Kingpin inclination (defined as the angle between the steering axis and the wheel centerline seen in figure 23) 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. However, KPI has the negative drawback of adding positive camber to the outside wheel in a corner. We are utilizing the wheel centers and uprights from the 2001-2002 Formula SAE car which will provide us with values of KPO and KPI. The wheel centers 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. This value of KPO seems reasonable despite the lack of any KPI.

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Figure 22. Scrub Radius [26]

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 (see Figure 23). It 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 with 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 a maximum at all times and thereby supplies the maximum amount of grip in corners.

Locating the Roll Center of the geometry can be argued by some designers to be the corner stone of suspension design. 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 24). 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 choice. 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. However, the roll center is an instant center and will move as the suspension travels. So, our goal is to keep this roll center in as fixed a position as possible. A final layout of the front suspension geometry is shown in Figure 25.

Figure 24. Roll Center [26]

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Figure 23. Camber and Kingpin Inclination

Kingpin Inclination

Camber

Figure 25. Front View showing Roll Center

In figure 26, 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 27 shows the amount of caster built into the suspension. The first iteration contains 5° of negative caster.

Figure 27. Side View with caster

Figure 26. Front View with Kingpin Offset and Camber

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Roll Center

Kingpin Offset

Caster

Camber

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.

Camber gain is also an important value in suspension design. Because the A-arms are not equal in length, non-parallel, the amount of camber seen at the wheel will change as the suspension travels. The A-Arm geometry setup can be designed and analyzed as a simple four-bar mechanism. As the tire travels upwards relative to the chassis (known as Jounce) it experiences negative camber in the order of 1.035˚ for every inch of travel. Similarly, as the tire moves down relative to the chassis (known as Rebound) it experiences positive camber in the order of 0.954˚ for every inch of travel.

Table 3.6: Front Suspension Geometry

Static Camber -1.5˚Camber Gain in Jounce -1.0353˚/

1"Camber Gain in

Rebound .954˚/1"Caster 5˚

Kingpin Offset .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"

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 or need to be modified. The amount of migration the roll center undergoes and crossing the ground plane during migration is a concern. The first modification of the front suspension geometry analysis was to raise the ground clearance height to 2” (which ensures the chassis never bottoms out in full rebound). The rise in ground clearance will alter the location of all three listed Roll Centers (RC) and change the equivalent four-bar location. Thus this alters the amount of camber gain and loss through suspension travel.

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Figure 28 shows 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 28 more clearly presents the equivalent four-bars that represent the front suspension. Figure 29 shows the suspension at equilibrium (in blue), at 1” of Jounce (in red), and at 1” of rebound (in green).

Figure 29. Suspension at Equilibrium, Jounce, and Rebound

Before a complete geometric analysis of the rear suspension, several iterations of the static Roll Center (RC) 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 to slightly 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˚Camber Gain in Jounce -.95˚/1"

Camber Gain in Rebound .89˚/1"

Caster 5˚Kingpin Offset .915"

Kingpin Inclination 0˚Toe In 0˚

Ground Clearance 2"Static Roll Center 1.23"

Roll Center @ 1" Jounce -.18"

25

Figure 28. Suspension Geometric Layout

ReboundJounce

Equilibrium

Roll Center @ 1" Rebound 2.65"

Front Track Width 50"

Table 3.8: Rear Suspension Geometry

Static Camber 0˚Camber Gain in Jounce -1.05˚/

1"Camber Gain in

Rebound 1.02˚/1"

Caster 3˚Kingpin Offset 0.02"

Kingpin Inclination 0˚Toe In 1˚

Ground Clearance 2"Static Roll Center 1.34"

Roll Center @ 1" Jounce 0.4"Roll Center @ 1"

Rebound 2.75"Rear Track Width 48"

Using ADAMS the vertical migration of the roll center during suspension travel can be determined for the full range of motion. Figure 30 shows 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 maximum vertical 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 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 the vehicle is rolling while turning a corner. Minimizing the migration of the roll center to maintain predictable and near constant handling is important and it becomes clear that a small value of roll center movement is ideal.

Figure 30. ADAMS Analysis of Suspension

The Rear Suspension geometry layout can be seen in figure 31 and figure 32. 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

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Figure 24. ADAMS Analysis of Suspension

therefore the ability to ideally locate the Roll Center. Figure 32 shows the amount of Toe In visible in red. When designing a suspension, the analysis is based on theoretical values and the actual layout will differ from these values. 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, seen in figure 33, that are bolted to allow shimming between the clevis and the chassis which adjusts the amount of camber built into the suspension and the static roll center. Furthermore, in the rear the clevises are placed on the vertical rails (which will be made of square tubing for ease of adjustability). In this way the clevises will be shimmed, and move up or down along the rails to adjust the roll center, camber gain, and static geometry (see figure 33).

By analyzing the lateral weight transfer of the vehicle based on the parameters listed below the amount of load on each wheel can be calculated. Because the most weight transfer occurs during hard cornering this will be the time in which there is the most load on the tires and therefore the time of maximum suspension travel. As mandated by the SAE rules the car must have 1 inch of travel in both directions. Therefore if we find the maximum load during cornering and want the suspension to travel one inch during that load we have found our spring rate. The following calculations

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Suspension Rail

Figure 33. Shimmed Clevises

Figure 32. Top View of Rear Suspension

Clevis

Bearing

Figure 31. Rear View of Rear Suspension

assume a 1g corner. This assumption is reasonable in light of researching high performance vehicles with maximum cornering capabilities of around 1g.

The idea of calculating the weight transfer in suspension is that the spring rates must be assumed in order to calculate the lateral roll resistance of the vehicle. Therefore an excel is produced to make easier iterative calculations. Because the spring rates are given by the maximum load divided by one inch, the numerical value of the maximum load becomes the numerical value of the spring rate. Table 3.9 lays out the iterative process of the spring rate calculations. The terminology and calculations can be found in the Appendix, section 7.1.

Table 3.9: Total Spring Rate ValuesSW WDR TM CM GM LM St Wt

550 0.545 48.909 1.29 14.63

613.34

6150.08

5187.06

8550 0.545 48.90

9 1.29 14.636

13.346

150.085

187.068

550 0.545 48.909 1.29 14.63

613.34

6150.08

5187.06

8

Many of the values for these initial iterations were known while others were assumed from previous data and estimations of our own vehicle. With that said these values can be easily updated in excel once further analysis of the vehicle is performed and the vehicle begins production so true values can be determined. Furthermore, these hand calculations can be used as benchmark numbers when using ADAMS to analyze the vehicle and optimize the spring rates.

3.2.1.3: Material StudyFollowing the logic behind our material use for the chassis, the suspension components will primarily be made of chromoly tubing. The tube diameter is 5/8 inch based on data from previous years’ vehicles. This diameter and wall thickness of the tubing is verified using FEA analysis on the suspension components. Before FEA can take place, the suspension is placed in an ADAMS simulation to determine the forces acting on the suspension components and where they are acting.

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The manufacturing of the a-arms will include flattening them at the ends to join to the bearing rings where the bearings mount shown in figure 34. The spherical bearings are mounted

to the A-Arms by press fitting them into the rings welded to the ends of the tubes. Aurora Bearing [20] recommends using this process. The bearings will allow 24˚ of misalignment which is sufficient for our design.

3.2.2: Braking3.2.2.1: Engineering SpecificationsAccording to the SAE rules [15], the brake system must act on all four wheels and operate by a single control. It must have two independent hydraulic circuits in case of a leak or failure the braking power is maintained. We will dynamically test the brake system to show the brakes lock all four wheels and stop 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 the brake pedal extending beyond its range, this switch will be activated and will stop the vehicle. This switch will cut the power to all electrical devices. The switch is executed with analog components, and not through the programmable logic controller. The car will have a red brake light 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.

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Figure 34. (a) A-arm with flattened ends at bearing rings (above), (b) pressed bearing welded to the a-arm (below).

(b)

(a)

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 HistoryThe 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 AnalysisCalculations 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, section 7.1.

3.2.2.4: Material StudyThe rotors are made from hardened steel and meet the specifications for handling the forces applied during braking.

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

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Figure 35. Brake and upright Assembly

Figure 37. Brake Pedal and Master Cylinder

Figure 36. Brake and upright Assembly

3.2.3.1: Engineering SpecificationsAccording 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 38 and 39.

Figure 38. Front view, Wheel shell and center

Figure 39. Rear view, Wheel shell and center

The uprights were also taken from the 2001-2002 Formula SAE car. The uprights, in figure 40, are made from aluminum 6061-T6 and were manufactured to fit inside the shells, as shown in figure 41. The suspension was designed to accommodate these uprights so that they provide correct amounts of inclination. The uprights’ unique design includes the calipers and rotors for the braking system.

Figure 40. Front Upright brake assembly

Figure 41. Rear Upright, brake, wheel, tire assembly

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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 42). This tire has an operating temperature of approximately 60-70 degrees Celsius.

Figure 42. Goodyear D1385, R065 - 20.0x6.5-13 [7]

3.2.3.2: Design HistoryOriginally two tire widths had been chosen – 7.5” and 6.5”. The 7.5” tire is a possibility because these are readily available. The 6.5” tire is an option because it weighs less and can reach its operating temperature faster than the 7.5” tire. Each tire also needs to be compatible with our chosen 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.

3.2.3.3: Engineering AnalysisThe following calculation is a simple design calculation to determine the distance it takes a tire to reach its operating temperature. Some assumptions are made based on values for tires and vehicle. The following data and assumptions are used in our 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: 25 C = 298.15 KelvinDesired Final Tire Temperature: 70 C = 343.15 Kelvin

Assumptions:Full sliding contactOverall Car weight plus driver: 165 lb = 74.842 7 kgConstant Specific heat is equal to rubber (Cv): 1600 J/kg-K Assuming a coefficient of friction: 1.7

Where mtire is the mass of the tire, Cv is the specific heat of the tire, Tdesired is the desired tire temperature, Tamb is the ambient tire temperature, m is the coefficient of friction, mcar is the mass of the entire vehicle, and g is gravity.

It was found that the 7.5” tire will take 1115.98 feet and the 6.5” tire only 772.54 feet. These calculations prove that the 6.5” width tires can reach full potential at a shorter distance than the 7.5” width tires. Based on these conditions, the 6.5” width tires are preferred.

3.2.3.4: Material Study Tires selection is based on weather, dry or wet, compound, and size. For our application, tires will be dry slicks. The tire material is made on only one type of compound – R065 compound. Sizes of the tires vary, but for our application a best fit tire is a 13” rim with a 6.5” tire width as proven in the engineering analysis. 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]. Since these types of tires are the same in material and close fit in size, the performances between Goodyear and Hoosier tires are comparable. With good contacts and the possibility of tire donations, the Goodyear tires have been chosen.

3.3: Drive System3.3: Drive System3.3.1: Motor and Power Train3.3.1.1: Engineering SpecificationsBased on race records of autocross events [24], the vehicle will most often race 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 a 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 REV. On courses with longer straight-aways, speeds of as much as 85mph may be achieved. To take advantage of these courses, the car also needs to be able to reach 85mph. Our requirement for a 0-60mph time was set at 5 seconds; however, as a race car, the faster it is capable of, the better.

The motor is the heaviest component of the vehicle. Therefore we must carefully analyze the motor, its placement, and mounting. There are several standards that the motor mount must conform to. The first is that the mount must be designed to fit within the frame and fit the specifications of the motor (i.e. set attachment points). Secondly, the mount must be designed to hold the engine weight so if the rear motor mount were to fail, the front motor would hold as shown in figure 43. Next, the mount must be able to withstand the maximum torque that the engine outputs. Finally, the rear motor 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.

Figure 43. Motor Diagram without Rear Motor Mount

Motor

Weight of Motor

Front Motor Mount

Attachment Points

Another important mounting bracket is the bracket for the differential. This bracket is required to hold the differential and its unusual shape. It also needs to mount the differential and align the input shaft directly with the motor shaft. This is an important tolerance design to ensure the direct drive of the motor.

3.3.1.2: Design HistoryThe 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. These performance calculations are found in the Engineering Analysis, section 3.3.1.3.

The design of the motor mount originated from a combination of several different electric motor mounts that were found during research on the subject. A more complex idea was first presented to create a thick plate to mount the motor. This design was simplified to vertically mounted motor and a damper system to reduce vibrations of the motor. The vertical mounting would allow us to adjust the height of the motor thereby moving the center of gravity of the car. The motor mounting, shown in figure 44, attaches directly to the skeleton of the car. Calculations were applied to this mount and it was found have a factor of safety of over 10. So, the forces that would be applied to these bars were determined (shown in Engineering Analysis, Section 3.3.1.3). This configuration was found to support the loading. So, the two bars were added as the front motor mount, and two straps were placed as the rear mounting (as shown in figures 44 and 45).

Figure 44. Front Motor Mount Shown With Partial Frame

Figure 45. Rear Motor Mount Shown With Partial Frame

The differential mount has been developed for the Kawasaki Prairie 700 front differential. The differential will be mounted to the rear of the car directly facing the motor. Mounting the differential to ensure that the differential is aligned to the axle of the motor will be a manufacturability concern. The differential has three mounting holes which will easily help oriented itself to the proper position. The mounts will be made to align with the three holes and attach to the frame with mount brackets as seen in figure 46. This design will allow for easy adjustability while being manufactured by first tacking the mounting brackets

to the chassis and checking the alignment until the differential lies inline with motor. Space will be left between the brackets and the differential mounts to allow for finer adjustments with washers. The mounts will be manufactured using the CNC mill machine.

3.3.1.3: Engineering AnalysisPerformance CalculationsBelow 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. 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

MountingBracket

DifferentialMount

Differential

Figure 46. Differential and Mount

lines connecting the points to visualize the area over which the torque is averaged.

FB01 motor, 0-60mph Speed Torque curve

0

20

40

60

80

100

120

140

0 500 1000 1500 2000 2500 3000 3500 4000 4500 5000

RPM

Torq

ue (f

t-lbs

)

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, or the average height.

Ti (i=1-6) represents the torque (in ftlbs) at each rpm starting at 0 rpm and going to 4300 rpm (4300rpm is approximately the rpm that coincides with 60mph). Tavg is the average torque at the motor shaft.

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.

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,.

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.

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 be easily 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.

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

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 which is 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.

For the analysis of the front motor mount, several hand calculations and a weight calculation are performed in Pro-Engineer Wildfire 2.0. First, a free body diagram is created to show the forces that the motor will apply to the bars of the mounting. This is shown below in figure 47, and after these forces were found they were resolved into X and Y components. These forces were used to find the shear stress in the 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 47. Left hand image shows the forces the motor causes, torsional and weight. Right hand image is the free body diagram with the reaction forces.

3.3.1.4: Material StudyMotor SelectionThe 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

Figure 48. Warp 9 motor

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 (in figure 48) did not have data available for it at high currents and voltages, but conservative extrapolations showed that it would provide better performance than 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 [11] 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 can be expected in heavy acceleration and cornering. A limited slip differential causes more of the torque to go to the wheel with traction, thus increases acceleration and control.

Figure 49. Kawasaki Prairie 700 Front Differential

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 as shown in figure 49 [22]. It is also lockable which allows full torque to be given to both wheels regardless of whether one is slipping 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. Since the motor mount design is welded to the frame, this material will be the same material as the frame for ease of weldability. 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 Source3.3.2.1: Engineering SpecificationsBatteries are the biggest component of the car. With today’s technology, there are many different batteries we can choose for this application. However, we need to compromise and chose the best battery for this vehicle by comparing efficiency, power, cost, and weight. Optimally, we want these batteries to include the following specifications:

- Battery Temperature Rangeo -30°C to +60°C

- Battery Temperature Rangeo Total battery pack voltage: 144 Voltso Total battery pack amperage: 550 Ampso Total battery pack amp-hours: 25 Ah

- Number of Batterieso Based on voltage and amperage of each

- Battery Pack Container

o completely isolate batteries and connections from the chassis and body

o hold half of the complete set of batterieso batteries easily accessibleo reduce movement and vibration of batteries during usage of

vehicle to keep batteries in contacto Container easily removable from side-pods

- Battery Connectiono Highly safe, no chance of electrocutiono connection between battery pack container and controllero easy to connect and disconnect battery packs for faster battery

pack switching

3.3.2.2: Design History

Figure 50. Current Side pod and battery layout

The figure above shows the battery packs which is 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) battery cells. Between the batteries, there are disc connectors made from a copper disc surrounded by a non-conducting (plastic) annulus. The annuluses hold the copper connector in position. The columns of batteries are held in place by a tube of PVC with a slot cut out to allow for thermocouples to be placed on the batteries. This setup is shown in figure 51.

The concept for the battery layout has changed because we need to ensure continuous connectivity. Since these cells were originally taken from DeWalt battery packs, we have decided to use the same concept from the DeWalt setup to ensure connectivity. The configuration used by the Dewalt batteries is shown in figure 52.

Side Pods (Frame support)

Battery packsFront

of car

Batteries (3)

Connector

PVC

The following is the third design concept for the battery packs.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

Figure 51. Single Battery Layout

Figure 52. Battery pack layout

Figure 52 shows two columns of five batteries each. These packs will be connected together to make up the battery packs in each side pod. The number of batteries needed has decreased according to our electrical engineers and the new number is 400. The current layout includes space for 480 batteries to allow for flexibility in the number of batteries in the future; batteries will be able to be removed from the current design if necessary.

There is a box that the battery blocks will be placed into make up one battery pack (see figure 53). The box will be used for two purposes, one to hold the batteries in place and the other is to safely isolate the batteries as described in the design objectives. The box will be completely enclosed with the exception of two power leads connected to the batteries, a small opening for sensor wiring exiting the box. As of right now there is no air flow vents in the box. If, during testing, the batteries are found to heat up excessively, over 35˚C, then air vents will be designed into the box.

Figure 53. Battery Pack in side pod Figure 54. Batteries and Box layout

3.3.2.3: Engineering Analysis

SL

SLSd

V, Ti

Figure 55. Staggered Tube Arrangement

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 layout of the batteries, shown in figure 55, are staggered cells to help improve the flow of air through the set of batteries.

The governing equation for rate of heat transfer from a liquid flowing through the bank of tubes is as following:

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 using the maximum fluid velocity. It is used in calculating the convection coefficient, h.

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 in Fundamentals of Heat and Mass Transfer [17].

Now all that is needed to find is the log mean temperature, T lm, to find the rate of heat transfer. To find log mean temperature the outlet temperature must be estimated by the following equation:

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.

These calculations are only rough estimates. Due to minimal information on the thermal characteristics of the batteries, further testing will need to be done to find the surface temperatures of the batteries under a full load. This will be done using the motor that will be used on the project. With this testing and limited to the battery layout, the above calculations will be used to space the individual battery packs and determine if air vents will be helpful.

3.3.2.4: Material StudyThe 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 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 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 (in figure 56) 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

Figure 56. A123Systems Lithium-Ion ANR26650M1 cell

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.

For the battery pack box, plastics were the first choice due to the fact that they do not conduct electricity and can be easily found. The approximate square inches of plastic for both battery pack boxes is 1,700 in2. LEXAN® fits both of these criteria. The specific type of LEXAN® would be 9034 Clear Polycarbonate sheet with a thickness of .220in. The cost of the sheet of LEXAN® would be in the ball park of $100 for a 36x48 in .220in.

3.3.3: Cooling3.3.3.1: Conceptual DesignIn 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 intermittence use and low amperage usage. Since our usage is beyond those specifications 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 StudyThe 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 pump system that we are using is a Alphacool AP700 12V Water pump capable of pumping up to 190 GPH. With this pump we are assured to be able to move the water from any location with in the vicinity between the radiator, controller, and water supply. With this extra power the water supply can be mounted lower for safety incase of a leak.

The Radiator we are using is either a Swifttech Quiet power 3x120mm Radiator or a Black Ice extreme II 240mm Radiator. The choice is pending data for the heat transfer rate of one of the radiators. The smaller of the radiator has a heat reduction power of 6270 BTU/h (1837 watt). Given this

value the controller produces around 2 watts of heat energy per amp. Assuming a 700 amp current draw the produced heat power is around 1400 watt.

The water overflow and storage system is still undetermined because it can be any coolant holding device which shape is unimportant. The container will fit in a location where it causes the least amount of problems, can be refilled easily, and causes the least amount of damage if it does leak, break or damaged in any other way.

Figure 57. Alphacool AP700 Water pump

3.4: Electrical3.4: Electrical3.4.1: Electrical Specifications and Interface RequirementsHigh Voltage (HV) RequirementsThere 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 ConnectionsNo 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 ConduitAll 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 non-conductive 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.

FusingAll 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 EquipmentThe 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 SwitchesThe 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 SwitchThe 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 SwitchThe 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 DisconnectThe 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.

SensorsThe PLC shall have the capacity for the input of the following sensors: Thermal, Voltage, Current, and Encoder.

Interface System Requirements:MenuThe 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. The menu is discussed in greater detail under software design details.

Warning LightsProgrammable 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 DesignThe 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.

Major hardware components and their technical specifications are listed below:

DC MotorNetGain Technologies, LLC [11]

Figure 58. Electrical Layout

Part Number: WarP 9

Length: 15.70 inDiameter: 9.25 inWeight: 156.0 lbsInput Voltage: 96-144V

Data at 144V inputTime On Volts Amp

sRPM HP KW

5 min. 134.0 320 4200 48.80 36.801 hr. 138.0 185 5700 30.40 22.90Continuous 139.0 170 6000 28.50 21.50

Peak Horsepower 100.00

Zilla ControllerCafé Electric [8]Part Number: Z1K-LV

Length: 9.00 inWidth: 7.00 inHeight: 4.63 inWeight: 15.5 lbsMaximum Motor Amps: 1000 ANominal Battery Voltage: 72 – 156 VPeak Power: 320,000 Watts

Dimensions of the Hairball interface:Length: 10.00 inWidth: 3.5 inHeight: 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 inWidth: 4 inHeight: 3.75 inWeight: 0.625 lbs

Figure 59. WarP 9 Motor

Figure 60. Zilla Z1K-LV Controller

Figure 61. Zilla Hairball Interface

Resistance: 0 – 5 kΩ

Speed SensorCafé Electric [8]Part Number: 2171S

Nominal Voltage: 12V

Figure 62. Throttle Control

Figure 63. Speed Sensor

FuseFERRAZ/SHAWMUTPart Number: A30QS600-4

Ampere Rating: 600 ADimensions (in): A: 3.13B: 1.22C: 1.63D: 2.44E: 2.31F: 0.31G: 1.00H: 0.19

FuseFERRAZ/SHAWMUTPart Number: A30ZS800-4

Ampere Rating: 800 ADimensions (in): A: 3.13B: 1.22C: 1.63D: 2.44E: 2.31F: 0.31G: 1.00H: 0.19

ThermocoupleOmega [10]Part Number: SA1

Temperature Range: -60˚C to 175˚CInsulation: TeflonSize: 25 x 19 mmLength: standard 1 m

Wireless Interface AdapterNew Micros, Inc [9]Part Number: XBEE PlugaPod-S

Size: 1.3" x 1.5"Weight: 0.4 oz Small C – freeware, includes limited assembler

Figure 64. Fuse

Figure 65. Fuse

Figure 66. Wireless Interface

512 words Program Ram24 General Purpose Digital I/O lines share functions with 4 wire SPI InterfacePower requirement for the PlugaPod is 6-9VDC @ 300mA or higher.

EZPLCEZAutomation [12]Part Number: EZPLC-D-96E

12 module slots96 I/OCommunication: Ethernet and serialNominal Voltage: 24VDC

Figure 67. Programmable Logic Controller

Lithium Ion BatteriesA123 Systems [18]Part Number: ANR26650M1

Nominal capacity and voltage: 2.3 Ah, 3.3 VMax continuous discharge: 70 AOperating temperature range: -30˚C to +60˚CWeight: 0.154 lbs (70g)

Figure 68. A123 Systems Lithium-Ion Specifications [18]

3.4.3: Technical Software DesignThe 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 InitializationAt 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 InputsThe 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 ProgramThis 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 OutputsAfter 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.

The following figures show the design of the output screen for the driver. This design features a touch screen in which the driver can observe the Main Screen and then switch to the Battery Status to evaluate any battery packs.

Figure 69. Monitor Display (Main Screen)

Figure 70. 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

Figure 71. Complete Vehicle Schematic

3.4.5: Reliability and Maintainability AssessmentsDue 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. The Zilla 1K calls for a liquid cooling which will be implemented prior to testing. 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/InterfacesDielectric Withstand TestThe 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 TestFor 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 TestingWe 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 TestingTo 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.

PLCTesting 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 TestsThe 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). We will similarly be testing Sout (to controller, PLC) in the same method.

Testing EquipmentA cart has been designed to facilitate the transportation and testing of the electrical hardware. To keep on schedule the electrical components need to be ready to install in the vehicle as soon as the mechanical build phase of the vehicle is completed. So to be ready the electrical components need to be tested together. Unfortunately all the equipment for testing of the electrical equipment is not in the same place and thus all the electrical components (weighing well over 250lbs) need to be transported from place to place while still being connected (the setup will be very complex and disassembly would not be practical). Also the motor needs to be constrained so that when it is tested it doesn’t roll all over the place (the high power of the motor combined with the large rotational inertia would spin the motor all over if it were not held in place).

For these reasons a cart has been designed that will hold the motor in place and provide both transportation and storage of the controller, batteries, and motor all while keeping them connected. The cart needed to have enough space, be strong enough to hold the weight, be able to fit in the school elevators, be able to be pushed over various surfaces, and require few resources (time and money) to make.

The cart is simply rectangular with two levels, with the top level a common height for work spaces. 1.5” tubing is to be used for the base frame and 1” tubing for the verticals, motor mount (a simple square with vertical bars to bolt to as in the final car), and top level frame. The tubing layout, seen in figure 72, will be welded together. Plywood or High density fiberboard will be used for the top and bottom surfaces. Large wheeled casters will be attached to the four bottom corners.

Figure 72. Electrical Testing Rig

3.5: Driver Interface/Ergonomics3.5: Driver Interface/Ergonomics3.5.1: Acceleration Pedal3.5.1.1: Engineering SpecificationsThe 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 73.

Figure 73. 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. The potentiometer layout is shown in figure 74.

Figure 74. Potentiometer with all given dimensions [8] Due to the limitations in dimensions given most of the potentiometer’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. The figure below shows the variables we will use to calculate a relation to determine these lengths.

Figure 75. Pedal assembly from side with dimensions

A simple relationship calculation will give us the wanted function:

s t

rd

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 leverα = 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

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 constantx = length the spring extends by

The distance the spring travels is given by

r = distance pivot point of pedal – attachment point of spring on pedald = 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 springβ1,2 =angles between pedal and base plate

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

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 StudyThe 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 Wheel3.5.2.1: Engineering SpecificationsThe 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.

3.5.2.2: Design HistorySteering 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 76 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 76b.

In addition to the previous specifications the touch screen needs to have some form of protection so that glare from outside light sources will not obscure the driver’s ability to read the screen. Also, this screen must attach in such a way as to not interfere with the driver’s hands.

Figure 76. (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 77.

(a) (b

)

Figure 77. (a) Steering Wheel with touch screen, (b) rear view showing quick release

3.5.2.3: Material StudyThe 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.

For the screen shielding there are two options. First option was to shield the screen with aluminum sheet metal. This metal would attach to the outside of the screen casing and would extend towards the driver for approximately four inches. This option does provide difficulties in manufacturing due to the shape of the shield, and to the restricts the space for a driver’s hand. The metal chosen was aluminum for its light weight and resistance to corrosion. The second option was to find an LCD screen protector that also provides an anti-glare protection. These screen protectors are commonly used on palm pilots or on laptop screens. These screen protectors will be an optimal choice for our vehicle.

3.5.3: Driver’s Seat3.5.3.1: Engineering SpecificationsA 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 StudyWe 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 (a

)(b)

but cost raise with the padding or flexibility of the seat material. This data is shown in Table 3.10.

Table 3.10: Tillet Seat Variations [13]Model Weight CostT11 1/4 pad 4.0 lb $239.00T11 no pad 3.5 lb $138.00T11VG 2.5 lb $170.00

Figure 78. 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 Equipment3.5.4.1: Engineering SpecificationsFor 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 StudyFigure 79 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.

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 80 shows a typical harness used for this application.

Figure 80. RJS 5-Point Harness [14]

Figure 79. Driver Racing Apparel [14]

4:4: BudgetBudget4.1: Initial Budget4.1: Initial BudgetThis 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).

4.2: Final Budget4.2: Final BudgetThis 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.

5: Organization and Capabilities5: Organization and Capabilities

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 Programmable Logic Controller Lead

Kathleen Murray Aerospace/Mechanical Engineering Aerodynamics Specialist

AJ Nick Mechanical Engineering Mechanical Designer

Matthew Reedy Electrical/Computer Engineering Electrical Lead

Joshua Wales Mechanical Engineering Systems Integration, Drive System Lead

David Wickers Mechanical Engineering Frame AnalystOliver Zimmerman Mechanical Engineering Mechanical Designer

Team Lead

Design Teams

Development Group Procurement Group Manufacturing Group

Integration Team

Chassis & Body

Chassis RedesignBody RedesignMounting PointsAeros/Ground Effects

Drive System

MotorDrivetrainControl SystemBattery SystemCooling SystemShielding System

Electrical

Battery Management InstrumentationData Transfer SystemPower Management

Vehicle Dynamics

Suspension SystemSteering SystemBraking System

Driver Interface & Ergonomics

Cockpit DesignSafety EquipmentDriver Interface

2007 R.E.V. TEAM STRUCTURE

6: Scheduling6: Scheduling6.1: Gantt Chart6.1: Gantt Chart6.1.1: Mechanical Task Schedule

6.1.2: Electrical Task Schedule

6.2: Milestones and Deadlines6.2: Milestones and DeadlinesMarch 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

7: Appendix7: Appendix7.1: Calculations7.1: CalculationsTo calculate the top speed, we factor in the top rpm the motor can handle and the gear ratio.

Suspension:The following is a set of variables and equations used in an iterative process to term the spring rates used on the suspension.

Key of TermsW = Total weight of car and driverWF = Total weight, frontWR = Total weight, rearUWF = Total unsprung weight, frontUWR = Total unsprung weight, rearUGF = Unsprung CoG height, frontUGR = Unsprung CoG height, rearTF = Track width, frontTR = Track width, rearCF = Height of front roll centerCR = Height of rear roll centerSGF = Sprung CoG height, frontSGR = Sprung CoG height, rearSF = Front spring rateSR = Rear spring rateWmF = Front wheel movementWmR = Rear wheel movementSmF = Relative front spring movementSmR = Relative rear spring movementSWF = Total sprung weight, frontSWR = Total sprung weight, rearSW = Total sprung weight

UtF = Unsprung weight transfer, frontUtR = Unsprung weight transfer, rearCtF = Weight transferred via front roll centerCtR = Weight transferred via rear roll centerTM = Mean track of sprung weightCM = Mean roll center of sprung weightGM = Mean CoG of sprung weightLM = Mean roll moment of unsprung weightSt = Weight transferred due to the sprung massWt = Total weight transferArF = Front roll resistance due to springsArR = Real roll resistance due to springsDrF = Front roll resistanceDrR = Real roll resistanceWtF = Total weight transfer, frontWtR = Total weight transfer, rear

Weight TransferSWF = WF- UWFSWR = WR – UWRSW = SWF + SWR

Unsprung weight transfer

Weight TranferSWF = WF- UWFSWR = WR – UWRSW = SWF + SWR

Unsprung weight transfer

Weight transfer via roll centers

Weight transfer via sprung mass

Total weight transferWt = UtF + UtR + CtF + CtR + St

Roll Resistance

Tire Load

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:

90 lb

360 lb

Brake Force CalculationsBrake Pedal:Assume that the driver input force is 90 lb.

Moment output from pedal:

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.

Where:

D: the master cylinder diameterF: the force from the brake pedalP: the pressure from the mater cylinder

4 in

The caliper:The calipers have two pistons that actuate the brake pads so the force is multiplied by 2.

Where:

P: the pressure from the mater cylinderD: the diameter of the caliperFCaliper Force: the clamp loadA: area of the caliper

Front Calipers:

Rear Calipers:

The brake pads:There are two brake pads so the force is multiplied by a factor of two.

Where: = coefficient of friction = 0.45 (good assumption for most race cars)

Front:

Rear:

The rotor:The torque applied on the rotor acts on both side so the torque is multiplied by 2.

Where:d: The distance between the center of the rotation and the force to act at a point midway across the rotor face.

Front:

Rear:

The wheels and tires:

Where:F: Force generated between the tires and roadr: Rolling radius of tire

Front:

Rear:

Acceleration calculation:

Where:a: Lateral decelerationF: 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.

Stopping distance:

Where:Si: the initial speeda: Lateral deceleration

Calculations based on Equations from Wilwood Engineering [25]

Motor mounting calculations prove that these mounts will handle the amount of load from the motor. A free body diagram shows the forces and loads applied to the front motor brackets.

RxTop

RyBottom

RyTop

W 2

W 2

F2y

F1y

F1x

F2x

RxBottom

Motor Mount Free Body Diagram

7.2; References7.2; References(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

(23) Carbon Fiber Tubing, http://www.carbonfibertubeshop.com/

(24) SCCA Autocross, http://www.scca.com/

(25) Wilwood Brakes, http://www.wilwood.com/

(26) Introduction to Formula SAE Suspension and Frame Design, http://campus.umr.edu/fsae/library/sae_paper/paper.html