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Vehicle Number 110

2015 University of Texas at San Antonio Baja SAE Design Report

Chase Jaffray Project Manager/Lead Engineer

Michael Didion, Geronimo Robles Contributing Team Members

Copyright © 2007 SAE International

ABSTRACT

The Roadrunner Racing Baja SAE team of the University of Texas at San Antonio has designed, analyzed, built, and tested a vehicle for the 2015 Baja SAE® Competition to be held in Portland, OR. This vehicle adheres to the Baja SAE® Rules and has been designed with sound engineering practice. This document describes the major design aspects of the 2015 model. All engineering decisions were made with a focus on safety, manufacturability, durability, and performance.

INTRODUCTION

SAE International® hosts annual collegiate design competitions for students around the world. The Baja SAE® competition is a part of this series and challenges engineering students to design and build a single-seater off-road vehicle to survive the most severe and rough terrain. Roadrunner Racing has approached this challenge with a focus on safety, manufacturability, durability, and performance. Economic and manufacturing constraints were large factors in the design process, but ultimately sound engineering practice was used. All computer aided design was done within SolidWorks®, and analysis software such as ANSYS® and Lotus SHARK® were used to validate these designs.

Figure 1: 2015 UTSA Baja SAE Vehicle

FRONT SUSPENSION

OBJECTIVE – The front suspension, Figure 2, was designed to succeed in rock crawling and high speed maneuverability scenarios. This was accomplished by minimizing bump steer and utilizing roll steer to improve high speed steering. The front suspension components were engineered to reduce weight while maintaining structural rigidity. This assisted in the reduction of the vehicles un-sprung weight, and therefore decreased lateral forces induced by turning, i.e. improved handling.

Figure 2: Front Suspension

DESIGN – The double wishbone suspension system was chosen due to the adjustability of the kinematic parameters. Spherical bearings were used for the wishbone outer joints and polyurethane bushings with a bronze-graphite dry-lubrication sleeve were used for the inner joints. These rigidly connected components promote robust force paths and low compliance. The upper wishbone is made of 5/8” OD AISI 4130 steel tubes, and the lower 3/4” OD AISI 4130 steel tubes. The lower wishbones outer joint is positioned at the same height as the dead spindle to allow for greater ground clearance. The largest forces that this joint sees are due to road

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force, therefore the lower wishbones spherical bearing was orientated vertically. The largest forces seen by the upper wishbones spherical bearing come from braking, therefore it was orientated tangent to the rotation about the spindle. Fox coil-over 2.00” shocks were selected due to their long travel and the adjustable dual-spring setup that controls roll and bottom out parameters. These long travel shocks improve articulation and wheel travel, which helps the tires maintain contact with the ground on uneven terrain. Custom uprights were designed of water jet steel plate welded together in a structurally rigid box. These uprights were engineered with a king pin inclination of 13.50°, and a scrub radius of 1.50” that promotes tension in the steering components. Front hubs from a Yamaha Raptor were repurposed due to cost savings and manufacturing limitations. Ride height was set at approx. 13.25” with 13.80” of total wheel travel; 6.55” of bump travel and 7.25” of rebound travel. The front roll center lies 7.00” above the ground, which is slightly lower than the rear roll center of 9.10”. This down sloping roll axis, from rear to front, will act as a mechanical advantage for the center of gravity to load the front tires in a turn. Static toe-out of -0.30° was chosen to keep the suspension components in their strongest modes, and static camber of -1.50° to counteract tire compliance in turning, i.e. keeping the tires normal to the ground. Finally, 10.00° of caster and zero caster-change were chosen to assist the transmission of forces when impacting obstacles and preventing false driver feedback.

ANALYSIS – Lotus Suspension Analysis Software, SHARK®, was used to analyze the front suspension in bump, roll, and steering applications. Data for toe change was exported to Excel, Figure 3, and shows the front tires toe-in when rebounding. This was designed to keep the tie rods in compression in the case of a front nose dive landing off a jump.

Figure 3: Toe Change Graph

Figure 4 shows the vehicle with the maximum amount of roll before the tires lift off the ground in a turn. Roll steer was engineered such that the inner tire toes-out more than the outer tire in a turn. This allows the vehicle to turn about a single point, improving handling at higher speeds.

Figure 4: SHARK Model Roll Analysis

Finite Element Analysis (FEA) was completed on the front suspension, Figure 5, and showed high stress concentrations in the upright around the spindle carrier. This analysis was calculated at an 8’ drop onto the fully extended front suspension. To neutralize these stresses, the spindle carrier was fully boxed-in with water jet steel.

Figure 5: Front Suspension Drop Analysis

REAR SUSPENSION

OBJECTIVE – The rear suspension, Figure 6, was designed to limit toe change, with respect to wheel travel, to ± 0.05°, and to function best on uneven terrain.

Figure 6: Rear Suspension

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DESIGN – A five-link independent suspension system was the best solution for this year’s vehicle. This was chosen over a trailing-arm system because the two trailing links in a five-link system could be altered to change the anti-squat characteristics of the vehicle. The links are made of AISI 4130 steel tube with opposite threaded rod ends at either end of the tube. The two trailing links and the toe link are 5/8” OD tubing, and the two lateral links are 3/4” OD tubing. The links were attached to the bearing carrier by water jet steel tabs, thus avoiding expensive machining services. The tabs holding the links to the frame were designed to hold two links per tab, maintaining the proper distance between the mounting positions. This removes one degree of freedom when manufacturing, improving quality control of the vehicle. The same Fox coil-over 2.00” shocks were used as the front, but wheel travel is limited to 11.20” due to the CV drive shaft joints operating angle. The rear suspension has 6.25” of bump travel, and 4.95” of rebound. Static toe was set to 0.0° with limited toe change, Figure 3. This will keep the vehicle from behaving unpredictably when the driver accelerates after a large bump or jump. Static camber was set to -1.50° to work in conjunction with the front suspension kinematics. Polaris RZR hubs were repurposed due to the matching spline pattern of the CV drive shafts and the high cost of machining female splines. These hubs were post machined to decrease weight and maintain a factor of safety of 1.2 in severe drop conditions.

ANALYSIS – The rear suspension was also analyzed in SHARK to determine the toe link placement, Figure 7. This figure shows the roll axis and theoretical center of gravity location. The camber change was designed to mimic the front suspensions kinematic trail.

Figure 7: SHARK Model

Spring rates were calculated to critically damp the vehicle hitting a bump at 20 mph with a weight bias of 45/55 (front/rear) and the weight of a 95th percentile male driver. A static structural analysis was completed on the rear components, Figure 8. This FEA was similar to the one completed on the front suspension. Results showed the weakest point in the rear suspension system was the lateral link outer rod ends. To counter this, larger rod ends were implemented.

Figure 8: Rear Suspension Drop Analysis

DRIVETRAIN

OBJECTIVE – The overall goal for the drivetrain was to design a light weight and durable system in an economically friendly manner. This was accomplished by avoiding advance manufacturing processes and purchasing commercial products to complement the provided Briggs and Stratton engine such as Continuous Variable Transmission (CVT), gear reduction/differential unit, and CV drive shafts (Figure 9).

Figure 9: Drivetrain

DESIGN – The drivetrain system was designed around the provided Briggs and Stratton 10hp engine. A CV-Tech CVT belt driven transmission was purchased because it complemented the engine and required less post-processing than its competitors. This unit outputs a maximum ratio of 3.00:1 and a minimum ratio of 0.43:1, which was tuned for the engines operating range. A CVT is also a safer choice. In the case of a locked up powertrain, the belt will slip and prevent excess damage to the engine or gearbox. The CVT is then connected to a Dana Spicer H-12 FNR gearbox. This unit provides a gear reduction of 12.58:1, a limited-slip differential, and Forward-Neutral-Reverse helical gearing for greater maneuverability. While this is a heavier alternative, the gearing and differential components provide more benefit to the overall vehicle. A coupler was then designed to attach the internally splined output shafts of the H-12

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gearbox to the CV drive shafts. This coupler, Figure 10, consists of a splined shaft attached to the CV drive shaft via a rubber giubo, or flex disc. The giubo is bolted on either side using alternative hole positions, so the splined output shaft and the CV drive shaft are not directly connected. With this design, the giubo acts as a torsional damper to absorb impulses from situations like landing the vehicle from a jump while the accelerator is engaged. The giubo also provides ~3° of angular deflection to the CV drive shafts 32° maximum operating angle, allowing more travel of the rear suspension.

Figure 10: Drivetrain Coupler

The CV joints from a Polaris RZR, inboard and outboard, were repurposed and attached to gun-bored and balanced shafts. These CV joints were chosen due to the mating of the outboard splines to the rear hubs, and the large amount of plunge the inboard CV joint provides.

ANALYSIS – Drivetrain calculations, Figure 11, are based on data from an engine dynamometer graph and factory specifications of the other drivetrain components.

Figure 11: Drivetrain Calculations

These calculation show the expected dynamic output of the vehicle. The largest variable in these calculations is the coefficient of friction of the tires and the ground, and the efficiency of the CVT and gearbox unit. Figure 12 shows the static analysis of the subframe with impulse forces from the engine, gearbox, and braking components in a situation where the vehicle is dropped from 8 ft and the wheels are suddenly stopped from max RPM. This component purposely has a higher factor of safety relative to other components on the vehicle. This is to ensure a rigid connection between the drivetrain components for a higher efficiency of torque transmission.

Figure 12: Subframe Impulse Analysis

CONTROLS

OBJECTIVE – The objective of the controls system was to provide a durable and responsive vehicle that was capable of being driven by a 95th percentile male for 4+ hours consecutively. This system includes both steering and braking subsystems.

DESIGN Steering – The steering system was designed around a 10” OD steering wheel located ~18” from the drivers chest. At this location, an average driver was found to output 48 ft*lbs [4], and still have room to egress the vehicle in 5 sec in case of an emergency. The steering wheel was rigidly connected to a dual-link column with a single sealed u-joint in the center. A steering rack with a ratio of 12:1 was used to generate full steering motion with 0.75 turns of the steering wheel. This will prevent the need for hand-over-hand driving, giving the driver more control of the vehicle. Lastly, custom steering rack spacers were designed to locate the inner tie rods accurately for the proper roll steer characteristics. Braking – The braking system was designed to lock up all four tires at our theoretical top speed. Wilwood PS1 1.12” bore calipers were used outboard on the front wheels and inboard in the rear. This was done to help reduce the un-sprung weight of the vehicle. Two 5/8” Wilwood brake masters were used to keep the front and rear brake

Max Torque 19.90 ft*lb 2340 RPM

Max Power 10.60 hp 3740 RPM

RatioMIN 3.00 1 1100 RPM

RatioMAX TORQUE 1.82 1 2340 RPM

RatioMAX POWER 0.49 1 3740 RPM

RatioMAX 0.43 1 3800 RPM

Ratio

Efficiency

WeightVEHICLE

DiameterWHEEL

μRUBBER-DIRT

TorqueWHEEL @ MAX TORQUE

ωWHEEL @ MAX TORQUE

Velocity@MAX TORQUE

ωWHEEL @ MAX Power

Velocity@MAX POWER

Engine Output

CVT Transmission

Gearbox

Calculations

0.95

112.58

mph

RPM

mph

550.00

23.00

0.70

432.77

107.60

5.15

642.45

30.77

lb

in

-

ft*lb

RPM

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systems independent. Hard brake line connects the masters and calipers to reduce the pressure loss due to expansion. A cutting brake was implemented between the rear brake master and the calipers. This will work in conjunction with the gearbox differential and allow the driver to lock one rear wheel independent of the other in the case of high centering or taking a smaller radius turn. The brake rotors are made of stainless steel due to its high coefficient of friction and corrosive resistance properties. The brake pedal provides a pedal ratio of 8:1 and can be repositioned 1” forward or back to accommodate drivers of varying leg lengths. ANALYSIS Steering – The tie rods were constructed of 3/4" OD 4130 steel tube to be sacrificial parts. Analysis was done to ensure that this tube would buckle before the upright or the steering rack yield, as this is the quickest and most inexpensive part to replace in the steering system. Braking – Thermal analysis was conducted on the brake rotors to compare geometry and the effects on cooling rate. This analysis concluded that by increasing the surface area of the rotors outer ring, the time to cool would decrease, thus reducing the risk of brake fade. The brake pedal was designed to endure 330 lbf for a minimum of 100,000 cycles. This is the 95th percentile male peak foot output force with respect to the angle of thigh and calf in a seated position [5].

FRAME

OBJECTIVE – The objective of the frame was to maintain the minimum amount of space around a 95th percentile male driver while still providing safety. The frame was also designed to be within a torsional rigidity range of 800-1200 lb/deg. to allow the frame to flex with the suspension, Figure 13.

Figure 13: Frame

DESIGN – The suspension points and the drivetrain were the driving factors for the basic frame design. AISI 4130 steel tubing was chosen due to its superior strength

properties. Tube sizing and node locations were based on iterative analysis.

ANALYSIS – The frame was analyzed for several cases, Figure 14. These cases were determined to simulate loads developed in off-road driving conditions. The forces were calculated using a composite spring rate to compensate for the front/rear springs, tires, and suspension compliance. A damping rate of 1.0 and an impulse time of 0.8-1.2 seconds were used. These assumptions led to force magnitudes of approximately 1800 lbf per tire in Case 1. Transfer functions were then used to transmit the tire loads to the various suspension nodes.

Figure 14: Frame Analysis Loading Cases

ANSYS® was used to complete static linear FEA of the resulting nodal forces. Figure 15 shows the results of Case 1, scaled to yield. As indicated, the frame does not yield in this case. Cases 2-3 showed some yielding, but it was determined that suspension components would yield before the frame in these cases. Later analysis revealed the vehicle would survive a 4ft drop at 35 mph landing on one tire. And to validate driver safety, Cases 4-6 showed no signs of yielding.

Figure 15: Frame Analysis Results, Case 1

CONCLUSION

Roadrunner Racing has designed and analyzed a vehicle for the 2015 Baja SAE Competition. With a focus on safety, manufacturability, durability, and performance, this vehicle has been engineered and validated to overcome the harshest of terrains.

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ACKNOWLEDGMENTS

Roadrunner Racing would like to acknowledge the support from our sponsors: Weebz Welding and Water Jetting, Woods Cycle Country, Colbath Transmissions, Rae Acuna, UTSA College of Engineering, Zachry Holdings, Boeing, Intertek, and O’Rielly Auto Parts.

Roadrunner Racing would also like to acknowledge our professional mentors: Prof. Jim Johnson, Dr. John Simonis, Allen Weible, Paul Krueger, and David Kuenstler.

REFERENCES

1. SAE International®, “Baja SAE® Rules”. 2015. Web.

http://www.sae.org/students/mbrules.pdf

2. Gillespie, Thomas D. “Fundamentals of Vehicle

Dynamics”. Print.

3. W.F. Milliken and D.L. Milliken, “Race Car Vehicle

Dynamics”. 1995. Print.

4. Steven Fox. “Cockpit Control Forces”. 2010. Web.

5. National Aeronautics and Space Administration. Man-

Systems Integration Standards. Volume I, Section 4.

“Human Performance Capabilities”. Web.

APPENDIX

See attached.

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