- 1 E-Board and Design Leads
- 2 Design Report 2010
- 2.1 FRAME AND BODY
- 2.2 ACCELERATION
- 2.3 BRAKES
- 2.4 DRIVETRAIN
- 2.5 ELECTRICAL
- 2.6 ENGINE
- 2.7 SAFETY FEATURES
- 2.8 STEERING
- 2.9 SUSPENSION
- 2.10 CONCLUSION
- 2.11 ACKNOWLEDGMENTS
- 2.12 REFERENCES
- 2.13 ADDITIONAL SOURCES
- 2.14 DEFINITIONS, ACRONYMS, ABBREVIATIONS
- 2.15 FIGURES
E-Board and Design Leads
University of Rochester Baja SAE Team Members:
- Susana Acosta
- Rachel Bierasinski
- Dustin Canzonieri
- Brent Gordon
- Max Greenberg
- Kohei Hamano
- Phil Katz
- Harold Lander
- Brandon Martindale
- Sam Sadtler
Advisor: Professor Sheryl Gracewski
Design Report 2010
This year the University of Rochester Yellowjackets has built a new car to compete in the Carolina and Rochester competitions. This year’s design features significant improvements in weight, suspension handling, and drive-train durability. The University of Rochester Baja SAE team built an entirely new car this year, focusing on several goals for improvement. The main goals are to reduce the overall weight of the frame, improve suspension tuning, build a more robust drive train, and improve the general aesthetics of the vehicle. The overall goal for this project is to improve these elements of the design and to produce a prototype vehicle which will appeal to off- road enthusiasts.
FRAME AND BODY
Designed for minimum weight while maintaining driver comfort and safety. In order to design the cockpit, we obtained measurements from all potential drivers while being sure to accommodate the 95th percentile male driver. We then measured each driver's legs in order to determine the optimum length of the frame and the ideal width of the SIMs at the location of the seated driver. The greatest sitting height obtained from measuring each driver wearing a helmet in the seat of the car was used to determine the height of the RRH and RHO members. The 4.5in radius bends of the RRH are spaced 31 inches apart at 31 inches above the bottom bar or 27 inches above the surface of the seat to accommodate rule 31.2.2 . This year we decided to minimize weight and firewall area by centering the beginning of the bends on a line located ￼￼￼27 inches from the seat bottom and closing the top of the RRH at a shorter distance than in year’s past.
The length of the cockpit was designed so that the shortest drivers are able to reach the pedals and the tallest drivers are not too cramped to drive the vehicle. This provides flexibility in available drivers. In addition, we made sure to allow for three inches at all points around the driver in accordance with rule 31.1 . The location of the seat mounting bars is so that no area is wasted behind the seat. Three 1.0” OD 0.035” wall bars are used to support the seat in the plane of the LFS. The first bar runs between the LFS members and two bars spaced 5 inches apart to accommodate 31.2.6 run from the first bar to the bottom of the RRH.
We determined the roll cage geometry and tube size based on knowledge of engineering concepts and an iterative design process utilizing FEA to allow for the strongest possible frame using the least material. Results of the analysis can be found in Table 1. Five load cases were used; full frontal impact (A), side impact on the extreme bend of one SIM member (B), side impact on both extreme bends of the SIM and LFS (C), a top downward load distributed on the four highest points of the RHO (D) and two downward loads acting at a 45 degree angle on the bends of the RRH and FBM members (E). In all load cases, the frame was constrained in six degrees of freedom at four points located at the front and rear of the vehicle on the LFS. A unit load of one pound, distributed among the number of forces in each load case, was scaled to determine both yielding and failure loads in addition to the g- forces necessary to produce yielding or failure based on an estimated car and driver weight of 600 pounds. Any section of the frame was considered unacceptable if it yielded with a g-force of 1.5 or smaller and fails with a g-force of 2.0 or smaller. The frame was acceptable under yield in load cases A, B, C and D with yield g- forces of 11.2, 1.8, 3.4, and 6.4 respectively. Load cases A, B, C and D are also acceptable under failure criteria with failure g-forces of 17.3, 2.7, 5.3 and 9.9 respectively. Plots of the FEA results for load case A, B, C, and D with yielding stresses shown are included in the appendix (Figures 1-4, Table 1). Load case E, however, proved unacceptable when only rear bracing is present with a calculated yield g force of 0.9 and failure g-force of 1.4 (Figure 5, Table 1). Analysis of this load case was rerun with the addition of front bracing made of 1” OD 0.035” wall bars producing a g-force of 1.5 in yield and 2.3 in failure (Figure 6, Table 1), prompting us to add front bracing in addition to rear bracing. Additional bracing extending from the beginning of the bends on the RHO to the outside members of the RRH was also examined in order to add strength while reducing weight. It was determined that this configuration under load case E would feel a g-force of 0.8 at yield of the side bracing (Figure 7, Table 1), meaning yielding in the frame was seen at lower loads with only side bracing present under load case E than with rear bracing only. Adding front bracing to this configuration, we determined that the frame was still weaker with side bracing as these members yield under a g-force of 1.3 under load case E (Figure 8, Table 1), which indicated that in order to strengthen the frame to our desired 1.5 yielding g-force and 2.0 failure g-force we would have to increase the size and weight of the side bracing. Therefore, the configuration of front and rear bracing was determined to be best under our yield and failure criteria while reducing the weight of the frame. Isometric and side views of the frame as modeled are seen in Figures 9 and 10.
We used 4130 steel instead of 1018 steel for the roll cage material. While 4130 steel has a greater cost, we benefit from an increased strength-to-weight ratio. To fulfill the requirements of rule 31.5 , we used 1.25” OD 0.065” wall for roll cage material. Our analysis showed that 1.25” OD 0.049” wall bars were adequate for our strength needs in the SIMs and LFS members. 1.00” OD 0.035” wall bars are shown in our analysis to be acceptably strong for the cross members as well as the seat belt and seat mount bars.
The engine compartment serves multiple purposes. The main purpose is to provide protection and a secure mounting point for the engine. The second purpose is to provide an additional safety factor of rear roll hoop bracing. The tertiary purpose is a mounting point for the rear suspension.
The compartment was designed to be as small as possible while allowing for the engine to be easily installed and serviced as well as sufficiently supported. To increase the strength of the over all frame we used the compartment members as rear roll hoop bracing. The shape was also designed so that the force from the suspension would be evenly distributed among the members.
The members where the least stress would occur used thinner steel pipe in order to minimize the weight added by the compartment and keep the center of mass over the wheelbase. Similar to the rest of the frame 4130 alloy steel was used, both to reduce the amount of steel used and to keep all frame materials homogenous.
The body panels will be aluminum sheets cut to fit over each plane created by the cross members between the SIMs and LFSs. They will be bolted on to the frame with small tabs welded on to each cross member. Aluminum was chosen over fiberglass, plastic and carbon fiber using a Pugh Matrix (Table 2). Skid Plate – The skid plate is made of 0.125 inch thick HDPE, which has a high strength-to-weight ratio, exhibits high impact and wear resistance. Additionally, HDPE has a low surface coefficient of friction making it less likely that our vehicle will get caught on obstructions during the dynamic events.
For the accelerator cable, we used a bicycle brake cable. The cable housing was constrained on either end by a small, machined aluminum block.
The accelerator pedal is a purchased component featuring a pivot at the bottom of the pedal and a rough surface for better traction. There is also an adjustable stop so the pedal throw cannot exceed cable travel.
The front brakes each utilize a standard Honda caliper with after market drilled rotors that are drilled to increase the venting of hot gasses generated while braking.
The rear brake assembly of our car uses a single Wilwood caliper. A single brake was used because of our car's swing arm and solid axle design. This one larger brake is powerful enough to lock both rear wheels. This is advantageous because it is lighter, cheaper, and uses less brake line and fewer fittings than two separate brakes would. The use of a single ten-inch diameter steel drilled rotor in the rear also optimizes vehicle weight, reduces temperature gain in the rotors during braking, and reduces brake pad wear.
The brake pedal was custom machined from aluminum stock and is mounted below the master cylinders.
Wilwood aluminum master cylinders were chosen due to their known reliability and compact lightweight design. They allow for the necessary pressure a 95th percentile driver can provide to the pedal. These two master cylinders were mounted in such a way that both are actuated simultaneously; one delivering hydraulic pressure to the front calipers and one to the rear caliper in a redundant manner as a safety consideration.
A standard Briggs and Stratton provided engine was coupled with a CVTech continuously variable transmission to provide power to the wheels.
Common in low horsepower, low torque, off-road utility vehicles, CVTs cycle through an infinite number of gear ratios providing smooth acceleration and do not require driver controlled shifting of any sort. Chain drive from the CVT to the solid axle was chosen over belt drive as chain provides less loss, is more robust, and does not require large pulleys. The minimum allowable distance between the upper and lower CVT pulleys was utilized to keep the engine as low as possible, thus keeping the center of gravity of the vehicle as low as possible. ￼￼￼￼￼￼￼￼￼
The gear ratio of the chain drive was designed to work with the high end set ratio (3:1) of the CVT and the engines max torque of approximately 13.8 ft-lbs, to give an average short distance acceleration of 11 m/s^2. The reduction of 7:1 from the CVT output shaft to the axle. This provided our vehicle with good short distance acceleration and an acceptable top speed for standard Mini Baja courses of around 40 mph. We also looked at using a 9:1 gear ratio. However, based on our experience with last year’s car, this would reduce the max torque output of the car. Chain – ANSI 50 chain was chosen based upon the horsepower and peak torque output of the motor . The center-to-center distance of the chain sprockets were chosen as to have an even number of links, thus avoiding the need to use half links, which are 35% weaker than full chain links . Chain tensioners are not required, as it was possible to calculate exact distances between chain sprockets, and because the articulation of the rear suspension does not affect center-to-center sprocket distance.
We chose not to include a manually adjustable gear reduction box or transmission due to the added weight involved in doing so. We do recognize that in a production off road utility vehicle the addition of separate high and low gears as well as reverse would be desirable; however, for the requirements of the Baja SAE competition itself, we feel a small and light car is more competitive than one with many different heavy components such as a reverse gear. In addition, the high and low gears would have been too similar to offset the weight gain required to have a reliable transmission.
The CVT, sprockets, and hubs were secured to the intermediate and drive shafts with the use of keyways. Hardened steel key stock was calculated strong enough to withstand the maximum torque the regulated Briggs & Stratton engine can produce. The sprockets and hubs were secured against lateral movement using set screws. To ensure that the drive shafts themselves would not be able to slide laterally, snap rings were installed immediately next to the shaft bearings.
For our hub design we looked at both aluminum and steel hubs. Both were strong enough in some simple initial calculations. Aluminum seemed liked the best option because it is a lighter then the steel hub would be. However previous experience with aluminum hubs showed that, while they may survive a race, once they are removed they show significant warping. These hubs would be risky because after extended use the continual warping would eventually case a fatigue failure. Steel is the best way to remedy this problem due to its greater stiffness and fracture toughness.
In order to reduce the weight of the entire swing- arm, an aluminum axle was chosen for the drive axis. A maximum stress value for the surface of the axle was calculated based on the maximum engine torque output and driveline gear ratio using the following formula: ￼￼￼￼τ= T ⋅ GR ⋅ r πr4 2
Cast aluminum alloy wheels (ITP SS112) were chosen because of their decreased weight over steel wheels and increased aesthetic value. The front and rear wheel sizes are both 10” diameter by 5” wide. They also feature the same bolt pattern (4x144mm). Using the same wheels in the front and the rear allows our team to keep half the number of spares as opposed to using different sized (height, width, or bolt pattern) front and rear wheels. These wheels from ITP also have a lifetime structural warranty, so if they bend or break they are replaced free of charge. This is a very enticing benefit for a potential customer. We also considered using bead-lock wheels. The advantage of bead-lock wheels is that they prevent the tires from slipping off the rim and thus deflating (unseating), usually at low tire pressures. The disadvantage is their increased cost over traditional steel or aluminum wheels. Since our vehicle will not be running with extremely low tire pressures (we do not compete in West events with a rock climbing (1) component where low tire pressures are desirable), the benefits do not outweigh the increased cost.
For the front, we decided to use Maxxis RAZR2 22x7-10” tires. In the past, we used the original RAZR tires and were very pleased with their traction performance in both turning and braking. The new RAZR2 series boasts increased depth and angled tread for even better traction and 6-ply construction for better durability and puncture resistance. They are also designed for the Grand National Cross Country Circuit, which has similar packed dirt terrain to what is usually seen on Baja SAE tracks.
In the rear, we chose to use ITP Mud Lite 22x8-10” tires. As with the front, through previous experience we have been satisfied with their performance. The radial tread design results in a great amount of traction from the drive wheels. They are also one of the lightest and most cost effective 6 ply mud tires on the market. We have, from experience, had very few durability issues, and they performed very well in an off-road environment.
The car incorporated three separate electrical systems; the brake light, two kill switches, and a digital display. The goal was to make all of them robust enough to stand up to the intense movement of the car during operation. They must also had to be insulated against moisture to avoid a short.
For brake lights, we used a LED strip light. It is not SAE rated as required by older editions of the rules, but is significantly bright enough to conform to the updated rules section. The light will mount to a bracket on a frame member in the rear of the car.
One driver-accessible and one bystander-accessible kill switch is wired to the vehicle’s engine for shut-off in the event of an emergency. The switches are a push style kill switch grounded to the frame and the 40-millimeter diameter red exterior can be easily seen. They are wired in parallel to the engine so that either will short the sparkplug, effectively stalling the engine.
Digital Display Console
A digital display console featuring individual tachometer, speedometer, and voltage indicators allows us to monitor the engine’s rpm, vehicle speed, and battery system voltage in real time while operating the car. This gives the driver information on the current status of the vehicle. The tachometer reading for example is helpful in the dynamic acceleration event, where the driver can increase engine speed to the CVT’s optimum torque engagement RPM before starting forward.
The tachometer sender uses an antenna wire to pick up the firing times of the spark plug which is then used to calculate engine RPM. The speedometer sender uses a magnetic pickup sensor (or Hall Effect Sensor) to measure the time between axle revolutions. We considered using a potentiometer to make this reading, but it requires an actual electrical contact to record revolutions and is less accurate. The hall sensor senses an increased magnetic field as the magnet passes by the pickup to record revolutions. Since this is mounted in a place which could get potentially muddy, an electrical contact would be undesirable.
The engine used is the required Briggs & Stratton 10 horsepower OHV Intek Model 205432 – Type 0036-el engine. The only modification to the engine was to remove the gas tank and place it above the engine compartment to better protect the engine, allow space for a spill guard, and to improve accessibility to the engine itself. This modification also allows for the engine to be mounted lower on the car and for the gas tank to be more easily refilled. Gas tank – The gas tank was mounted onto the rear section of the car in the optimum position for gas flow from the tank to the engine and a splash guard was mounted around it.
Seat – Driver safety and comfort is assured by the high back Summit racing Baja seat and cover. This seat provides the driver with a comfortable ride and holds him or her in place in tight turns. The seat is bolted ￼￼￼￼￼￼ directly to the frame to ensure that it will not come loose from vibrations. Seatbelts- For seatbelts, we chose a G-force five point harness. This belt cost less than those from Simpson and satisfies the required mounting conditions. It also has a quick release lever to allow the driver to make a quick escape from the vehicle. The lap and submarine belts are mounted using a steel tab which is bolted to tabs on the frame in double shear. The shoulder belts are both mounted around a 1 inch, .035 wall tube in the RRH and are constrained from sliding on the bar. Guards – Protection from moving parts are critical on vehicles, so the CVT and chains are made in a variety of different ways, all of which are safe and conform to the rules. The CVT guard will be made of vacuum formed HDPE. The other chain guards will be made out of steel because, in addition to containing the chain in case of an incident, will also protect the chain and sprockets from being destroyed by obstacles the car may encounter.
Our goal for the steering design this year was primarily to eliminate excessive bump steer and camber gain and to create Ackerman steering characteristics. Both were accomplishedusingUnigraphics® .Thisyearwe purchased a rack and pinion style steering mechanism with custom extensions and nylon sliders mounted to either side of the rack. Because of the foreword placement of the rack and pinion, it was necessary to attach a bumper to the car to shield the tie rods from impact. The tie rods have both left and right hand threads to allow for quick adjustments and jam nuts keep them from moving during operation. There is a 10- degree castor angle resulting from the rake in the front of the car, a 10-degree kingpin angle on the uprights, and the camber has limited adjustment in the ball joints. The steering column utilizes no U-joints and has been placed in a location that allows for comfortable operation for a range of drivers.
Front- The primary objectives for this year’s front suspension were to maintain a static ride height of 12 inches and to address a previous issue with approach angle. Other goals include serviceability, ease of manufacturability, and resilience. To meet these goals, we chose a duel wishbone style suspension coupled with a pair of reservoir air shocks. The Suspension assembly was modeled in Unigraphics in order to perform FEA and dynamic analysis (Figure 11).
An unconventional geometry for the a-arms was chosen to increase the approach angle for the car (Figure 11). This choice is in response to an inadequate approach angle on the previous car. In order to drastically reduce construction time, an alignment jig was constructed for each of the a-arms. The top a-arm is constructed from one inch, 0.035 4130 steel tube and the bottom is constructed from on inch, 0.065 4130 tube to account for the greater load from the shock. The bottom a-arm is modeled using the Unigraphics® design package, to simulate extreme loading on the suspension (Figure 12).
The front uprights are assembled from 4130 steel plate and a spindle was fabricated to accommodate an OEM Honda hub. The hubs have a 10 degree kingpin angle in order to improve steering stability. We chose to fabricate the front hubs this year in order to accommodate a foreword rack and pinion placement and to allow for easier customizability of steering geometry.
The shocks are purchased from F.O.A (First Over All). These shocks retail for significantly less than comparable shocks used in the past. They have also been custom valved to match the suspension geometry and the spring rate is easily adjustable by altering nitrogen pressure. Rear- The intended purpose of the rear suspension of our vehicle was to efficiently transmit the power generated from the engine to the wheels efficiently. This made wheel to ground contact, or traction, the main consideration in our design. Safety, cost, weight, ￼￼￼￼ and ease of maintenance were the other main factors taken into consideration. Two different suspension designs were discussed: one solid axle based and one independent design. The basic swing-arm design was also investigated as its weight savings, stability, and simplicity make it a viable option for this application. An independent rear suspension was also considered. At larger speeds and on terrain much more severe than seen by Mini Baja vehicles, where there is a significant amount of normal to the ground forces on jumps and such, an independent suspension would be a better choice. With speeds limited to around 45 mph and the lack of large jumps, a robust and simple solid axle swing arm design was chosen.
This swing arm design is an improvement on the previous year. We have chosen to build a tubular 4130 steel swing arm, which is a much lighter choice than the 1018 steel rectangular box used last year. Also, as opposed to previous years, bushings were chosen over heim joints for the connection to the rotational axis. An FEM model was designed and analyzed for a 1000 lbf load on the side of the wheel. We then performed iterative design to find the optimum bracing tube size and thickness. See figure 13 for the stress analysis results from our optimal design.
Dual 18-inch nitrogen filled reservoir shocks without coils were used to support the swing-arm. The pressure in the shocks is adjustable and was set to allow an approximate wheel travel of 8 inches upward and 4 inches of droop. This in combination with the swing arm’s uni-axial rotation allows for sufficient articulation on uneven terrain.
This year’s design has made many improvements, including a reduction in overall vehicle size and weight, a more serviceable and robust drive train, a polished appearance, and better suspension dynamics. Besides accomplishing these goals in design and fabrication, underclassmen and new members have been given an invaluable opportunity to get engineering experience outside the classroom.
We would like to acknowledge the staff of the Mechanical Engineering Department, the Student Association, and the River Campus Machine Shop at the University of Rochester for their assistance with funding and organizing.
- 2009 Baja SAE Series Competition Rules
- Bedford, A. and Liechti, Kenneth M. Mechanics of Materials. Prentice Hall. Upper Saddle River, NJ: 2000.
- Online Materials Information Resource. http://www.matweb.com. 2009.
- Marks, Lionel S. Marks’ Standard Handbook for Mechanical Engineers, 9th ed. McGraw-Hill. New York, NY: 1987.
- Brady, Scott, Tire Selection for Expedition Travel: The impact of tire width on traction. http://www.expeditionswest.com/research/white_papers/tire_selection_rev1.html, 2005
- Milliken, William F. Race Car Vehicle Dynamics. SAE International: Warrendale, PA; 1995.
DEFINITIONS, ACRONYMS, ABBREVIATIONS
- CVT: Continuously variable transmission HDPE: High density polyethylene
- LC: Lateral Cross Member
- LFS: Lower Frame Side
- NiMH: Nickel-metal hydride
- OD: Outer diameter
- rpm: Revolutions per minute
- RC: Remote control
- RRH: Rear roll hoop
- SAE: Society of Automotive Engineers SIM: Side impact member
￼￼Figure 1 - Front Impact (A)
Figure 4 - Top Load Distributed On Top Of The Roll Cage (D)
￼￼Figure 2 - Side Impact Directly On SIM Bend (B)
￼Figure 5 - 45 degrees Downward Load at Both RHO Bend – Rear Bracing (E)
￼Figure 3- Side Impact on SIM and LFS Bends (C)
Figure 6 - 45 degrees Downward Load at Both RHO Bend – Rear and Front Bracing (E)
￼￼Figure 7 - 45 degrees Downward Load at Both RHO Bend – Rear and Side Bracing (E)
Figure 10 - Side View
￼￼Figure 11 – Front Suspension and Steering Model
￼Figure 8 - 45 degrees Downward Load at Both RHO Bend – Rear, Side, and Front Bracing (E)
￼Figure 9 - Isometric View
Figure 12 – Lower A-arm with 1000 lb load
￼Figure 13 – Swing arm side load stress analysis
￼Table 1 – Summary of Frame Analysis Results
￼Table 2 - Body Panel Selection Pugh Matrix (Fiberglass is the baseline in this matrix)