Car 28

New Mexico State University 2016 Baja SAE Design Report

Reese Myers, Justin Squire, Gregory Taylor New Mexico State University

Arthur Cox, Glen Throneberry New Mexico State University

Copyright © 2007 SAE International

ABSTRACT

The report details the design of the 2016 NMSU Baja vehicle. Motivating the design this year was the overarching goal of a light weight vehicle with excellent performance and durability in all dynamic and endurance events. To achieve these goals, analysis with Matlab, Unigraphics, SolidWorks, and MSC ADAMS was used to produce and validate design parameters. With these parameters established, a top-down design strategy was employed to produce an optimal integration between the various subsystems. Manufacturing of the vehicle was simplified and expedited through reduction of part complexity and the selection of fabrication methods best suited to the design.

INTRODUCTION

Primary goals for the 2016 NMSU Baja vehicle included minimizing weight, tailoring the design to the requirements of dynamic events, and operating within budgetary constraints. Since the power output from the engine is limited, an emphasis was placed on minimizing weight throughout the system. Each component of the car was designed with consideration given to the incremental weight that it would add to the system. Because the project was self-funded, careful thought was put into the marginal benefit of design options relative to the marginal costs. This allowed for investment of resources in design options with the greatest determined value, while making compromises where secondary options resulted in minimal performance sacrifices. In keeping with the overall goals of the vehicle, a top-down design process was employed and utilized consistently across all subsystems. Goals were established for each subsystem in order to identify the parameters of highest importance. The analysis then provided a definition of specific requirements for the system and allowed for the maximization of desirable traits in parallel with a minimization of weight and monetary costs. Following the establishment of subsystem requirements, component design was initiated by the subgroups, with special consideration being given to intersystem dependencies and interfaces. The top-down design approach was

exemplified by the use of a master CAD model and management scheme. Utilizing free, cloud-based services, an extensive CAD management system was developed. Rules were established for part check-out and check-in, to eliminate the risk of overwritten or lost work. Additionally, a consistent part numbering procedure was created, along with a protocol for file versioning and access to revision history. Coordinate systems were placed in the master model as a tool to define the locations of the subsystems. With this structure, subsystem assemblies can be constrained to the master model by using only one coordinate system, which allows for modification or replacement without issue. Only after integration was complete did manufacturing begin, precluding changes to the design so as to accommodate unforeseen issues. Finally, the testing phase of product development allowed for final tuning to the system prior to competition.

DRIVELINE

GOALS - The driveline system was designed around the primary goals of minimizing weight and reducing the amount of time required for the vehicle to travel 150 feet. While keeping these primary goals in mind, consideration was given to designing a system capable of withstanding cyclic loading, with excellent fatigue resistance. Finally, a large part of driveline design was driven by ergonomic considerations such as comfortable shifting and throttle controls, in addition to providing easy access to engine and transmission components on the vehicle.

ANALYSIS - Top speed, peak torque and the ideal gear ratio for the vehicle were the first factors considered for analysis. An ODE was determined that would describe the motion of the car (Eqn 1). The ODE is not analytically solvable due to several nonlinear variables including the RPM dependency of the CVT, the dependency of the RPM on gear ratio and velocity, variable engine torque output, engine governing, and aerodynamic drag. A Matlab script was written and employed to simulate the vehicle traveling over 10 seconds at variable transmission ratios. A range of 9.85 to 10 was found to be the range of ideal gear ratios, in order to travel 150 feet in the minimum

amount of time (Fig 1). The higher end of this range yields a top speed of approximately 28 mph. Based on transaxle selection, the actual gear ratio of 10.15 was used in calculating the maximum incline of 36 degrees at which the vehicle would be capable of maintaining a constant speed of 10 mph. With steeper inclines, the vehicle would begin to decelerate until coming to a stop.

DESIGN AND INTEGRATION - The first step in the design process was to choose a transmission in the range of the optimal gear ratio. Additional measures included weight, availability, and cost. Based on this rationale, the Dana H12 FNR gearbox was eventually selected. With a 10.15 ratio and an approximate weight of 35 pounds, the transaxle narrowly falls outside of the optimal range, but fits well with the other criteria. Forethought was placed into its integration with the driveline, suspension, and chassis. By mounting the transaxle lower in the chassis, benefits are noticed in the rear suspension, where CV shaft angles are significantly reduced. This allows for a prolonged lifespan and reduces the risk of a failure. Additionally, the center of gravity is lowered, increasing the overall stability of the vehicle. In order to reduce the spinning mass on the vehicle, the design was based on the specifications of a Gaged CVT with a 10” center-to-center distance. Following transaxle placement, this constraint allowed for the determination of engine placement and mounting. Since the CVT belt is known to stretch by approximately 0.1”, a mounting system was devised to accommodate belt tensioning through the use of shims. With the short wheelbase designed to improve handling, engine mounting was limited to locations directly above the transaxle. Although beneficial in some aspects, the engine mounting also comes with the drawback of raising the center of gravity. The next step for system integration involved selecting a CV shaft that could mate to the spline pattern on the output shaft of the transaxle. Although the transaxle offers compatibility with a Volkswagen Type 1 CV joint, these CV joints only provide an angular displacement of 15 to 17 degrees, much less than the minimum requirement of 20 degrees determined from the suspension geometry. Although type 4 CV joints offer a much higher angular displacement and plunge capability, they also have a wider diameter flange, which does not readily mate with a Type 1 flange. A special adapter was designed to accommodate the two dissimilar flanges and provide a link from the transaxle to the wheels. This flanged CV end also improves the serviceability of the vehicle by allowing for quick installation and removal of the CV shafts and transaxle, without disassembling the suspension components. As a final example of system integration, the CVT housing was designed with the intent to interface with the firewall of the vehicle. Manufacturing data indicates that the CVT belt performs best at around 180 degrees Fahrenheit, with a possibility of breakdown or degradation occurring when the belt reaches temperatures higher than 200 degrees Fahrenheit. Since the CVT belt has such a narrow range of operation, a cooling mechanism was designed to gather the compressed air present at the firewall surface when the vehicle is moving. A fully enclosed CVT case

with one inlet and one outlet was selected, in order to increase the efficiency of the convective cooling method.

MANUFACTURING AND TESTING – Several components of the driveline system necessitated noteworthy manufacturing processes. Firstly, because the selected CV joints are intended for use with a larger Type 4 flange, a specialized adapter was required to allow for mating with a smaller Type 1 flange. 6061 Aluminum was machined to accept the two different step sizes of the CV flanges. Although both Type 1 and Type 4 CV joints accept the same bolts, the bolt circles are different, which required provisions for all twelve bolts to be secured in the flange adapter. The two different bolt patterns were offset from each other by fifteen degrees and threaded for helical inserts. Due to the use of modified Polaris rear hubs, it was necessary to splice the Volkswagen CV shaft with the Polaris CV shaft. Ideal CV shaft length was determined from the CAD model, examining extension and compression throughout the range of suspension travel. The two different axle shafts were cut to length and turned down to a common diameter on a lathe. An interference-fit sleeve consisting of 4130 steel was prepared with holes pre-drilled for rosette welds. Final assembly was facilitated by cooling the shafts in a freezer and heating the sleeve before both ends of the sleeve were fully welded, along with the rosette welds.

SUSPENSION

GOALS - The primary goals of the suspension system included a low stiffness at ride height, in order to isolate the vehicle from bumps and other small obstacles, and the ability to shift the chassis out of the way of larger obstacles. Additionally, it was desired that the entire system be lightweight but robust, which coincides with the overall goal of minimizing vehicle weight.

ANALYSIS – Modal analysis was used to verify that any existing modes between the suspension system and the engine or chassis were not being reinforced. This allows energy to be distributed more evenly among the chassis members and to prevent possible reinforcing interactions in cyclic impact. Additionally, a half car model was used to determine the linear natural frequencies of the car. With the creation and use of a Matlab script, the linear and nonlinear forms were compared. It was determined that for angles less than 45 degrees, the linear form maintains accuracy. The frequencies of the undamped modes are 1.2 Hz and 2.0 Hz (Fig 2) with the former consisting mainly of vertical motion and the latter consisting mainly of rotational motion. In order to maximize driver comfort and reduce vehicle fatigue, minimization of these frequencies were sought by reducing the spring rates of the shocks. Taking advantage of the progressive rates offered by the shocks, the natural frequencies at ride height were reduced, while still allowing for higher spring rates at larger displacements.

DESIGN AND INTEGRATION – To satisfy the suspension goals, a double wishbone setup was chosen for the front and a three-link trailing arm setup was chosen

for the rear. Front suspension design began with a sequence of 2D sketches to establish the constraints and design parameters chosen for the vehicle as a whole (Fig 3). These parameters included a front track width of 52”, a ride height of 12”, and 2” of droop. An 8° camber change throughout the travel was also desired. This leads to a 0.52 degree of camber change per degree body roll. With this amount of body roll, the tire patch maintains some contact through turns, while still preventing too much camber change from impact. A static Ackerman percentage of 100 was selected to prevent bump steer and help prevent understeer while at higher speeds. Both the front and rear suspension systems were designed to minimize toe change throughout the travel which helps maintain power efficiency. From the 2D sketch, the kinematics of the front suspension were evaluated and adjusted iteratively until the design goals were achieved. Rear suspension design was carried out similarly, with the use of simplified 3D geometry constrained to the ride height, droop, and rear track width (48”) selections. The rear link components of the trailing arm setup were designed to be parallel and rotationally in line with the CV shafts. With this arrangement, relative length changes of the CV shaft are minimized as the suspension moves through its travel. The front knuckles feature a unique, lightweight design, consisting primarily of a sheet metal weldment. Based on the original kinematic sketches, a sheet metal part was designed around the critical locations needed to preserve the geometry and allow for connection with the control arms, steering link, hubs, and brake calipers. Particular care was given to ensuring the compatibility of the knuckle and axle design with the front hub that was selected. The structure of the knuckle provides for a high moment of inertia per unit mass, allowing for high resistance to bending forces applied from impacts on the tire. FEA-driven design iterations resulted in the final product that was integrated into the front suspension system. Mounts for brake calipers were also designed for placement on the front knuckle and the rear carrier portion of the trailing arm. In order to meet the goals for the suspension system, progressive shock rates were required. With both budgetary and dimensional constraints in mind, dual-rate shocks were selected for the front suspension and air shocks were selected for the rear. Both of these options allow for the vehicle to absorb small impacts without much shock to the system while also preventing the car from bottoming out under large impacts.

MANUFACTURING AND TESTING – Unique to this vehicle are sheet metal weldments for both the trailing arm and the front knuckle. With this arrangement, a complex geometry was achieved without highly technical or time-intensive manufacturing processes. The CNC water-jetting operation used to form the variety of sheet metal parts on the vehicle prevents the need for complex tooling and expensive machining costs. Although bending the pieces for the trailing arms was straightforward on a break, a two-piece jig was created to form the radius on the front knuckles using an arbor press. Front axle shafts were made out of normalized 4130 steel with the use of a manual lathe. Operations included the machining of

bearing surfaces, threads, and complex chamfers. Another manufacturing process included the creation of a jig for use in fabricating the control arms, in order to maintain the co-axial constraint between the bushings. The jig fixes the distance of the bushing contact surfaces relative to each other, in addition to fixing them along the same axis.

BRAKES

GOALS - The braking system is critical to the performance and safety of the vehicle. With this in mind, the brakes must be able to easily stop the car with an appropriate amount of input force by the driver. The two goals of the braking system were to allow for locking of the brakes with 60 to 80 pounds of force applied at the pedal, and to reduce the mass which the brakes are responsible for slowing to a stop.

ANALYSIS - It was determined through testing that a force ranging from 60 to 80 pounds is a reasonable expectation for the driver to produce in attempting to lock the brakes and stop the vehicle. Maximizing rotor diameter is the only design parameter where the system can gain mechanical advantage without increasing pedal travel required to lock the brakes. Additional mechanical advantage can be obtained through extending pedal length and increasing hydraulic forces in the brake system but at the cost of pedal displacement. The NMSU team has developed a Matlab GUI which can assist in the design of correct brake bias in addition to normal forces and net negative acceleration. Utilizing the braking system GUI with the parameters of the vehicle (see Fig 5), it can be seen that the rear brakes should lock up at 67.5 lbs and the total system locks up at 76.6 lbs. This was a pre-determined feature in the design of the braking system, allowing the driver to intentionally produce oversteer by applying an intermediate pedal force.

DESIGN AND INTEGRATION – The final system design converged on the usage of outboard brakes for several reasons. First, the distance between the CVT housing and the output shaft of the transaxle was not adequate to accommodate a rotor of reasonable size. Second, it was desired to eliminate alternating loads on the CV shafts that would be induced by inboard brakes. Additionally, because the driveline system features a limited slip differential, rotors would be required on both sides of the transaxle, increasing the complexity which inboard brakes are generally intended to eliminate. Caliper mounting was seamlessly integrated into the front knuckles and the rear bearing carrier with the routing for brake lines also incorporated into the suspension components. This reduces exposure of these components to course hazards since brake components are able to travel with the suspension components. Both master cylinders and the bias bar are mounted compactly under the kick plate in the foot box area, reducing obstructions and providing for a more ergonomic cockpit. Due to space constraints under the kick plate, external reservoirs for the master cylinder are mounted higher in the cabin, providing the added benefit of residual pressure in the lines.

CHASSIS

GOALS - Integral to the capabilities and functionality of the vehicle was a purposefully engineered chassis. It was desired to design a chassis that represents an ideal integration of the required structure for subsystem functionality with the physical constraints mandated by the rules. Top level goals for the chassis consisted of several facets. First and foremost, a high stiffness to weight ratio was desired. Accessibility and ergonomic considerations were also a driving factor in the design, in order to accommodate comfortably a wide range of drivers. Finally, a high torsional and axial bending strength were needed so as to safely handle the loads delivered by the suspension and driveline subsystems.

ANALYSIS - Since the tubing choice for the chassis most directly impacts the goal of a high stiffness to weight ratio, the first step in design was to identify the optimum tubing material, thickness, and outside diameter, within the constraints of the rules. Due to its high Young’s modulus per unit mass, 4130 chromoly steel was chosen. Following this decision, wall thickness (0.065”) and outside diameter (1.25”) were chosen as the best parameters that also complied with the Baja SAE rule requirements. Using ADAMS, the modes of vibration were analyzed to provide insight into optimum placement, orientation, and sizes of chassis members. With the suspension mounting points fixed the lowest mode was at 48.5 Hz (Fig 6). This was much higher than any mode of the suspension system so resonance based fatigue interactions are minimized.

DESIGN AND INTEGRATION - Based directly on the modal analysis that was carried out, chassis members relating to low stiffness were reinforced in critical areas, through triangulation and increasing tube stiffness. Particularly important to the optimized structure of this chassis was the fact that it was designed following the completion of finalized designs for the subsystems of the vehicle, thus being driven by their constraints. In this manner, the chassis design provides the necessary structure for the designed functionality of the subsystems while offering ample reinforcement for mounting and attachment points. Through this method, the design employs the most efficient use of tubing, while eliminating the unnecessary dead spaces and sub-optimum subsystem placement that typically prevail when the chassis is designed first. Examples of this design intent can be seen in the built-in rake angle of the front suspension mounting points, the placement of a chassis node at the ideal upper mount location for the front shocks, and the single crossmember in the footbox supporting both the compactly placed master cylinders and the steering rack. Aside from structural and rigidity considerations, forethought was also placed into accessibility for assembly, service, and maintenance tasks. Easy installation of the engine from above, installation of the transmission through the convenient interface mounts, and easy removal of the CVT and its

cover are all notable examples. Finally, chassis design was also heavily influenced by ergonomics. In addition to satisfying the overall goal of weight reduction, a suspension seat was selected in order to comfortably accommodate a wide range of drivers.

MANUFACTURING AND TESTING - Manufacturing a chassis to strict tolerances can be a challenging task without proper preparation and protocols in place to ensure accuracy and consistency of lengths, bends, and notch profiles of the tubing. To ensure the chassis was manufactured according to design, full sized templates of each chassis member were printed and used for direct comparison during the fabrication process. Additionally, a method was developed to create paper templates for notching the tubes accurately. The process involved manipulation of the individual tubes in the CAD software to create paper-thin representations, allowing for the ends containing notches to be cut and unrolled as a sheet metal body. Flat patterns could be printed at a 1:1 scale, cut out, and taped to the tube at the correct location. Since all tubes had only in-plane bends, a view that was normal to the plane of the bend(s) was used as a reference for aligning the split lines of the paper templates, so as to provide the correct phasing of the notches. Following this procedure, the chassis was fabricated quickly and accurately, providing exceptional, gapless joints which greatly facilitated the welding process and contributed to the overall goal of high torsional and impact strength. As a final note on manufacturing associated with the chassis, the suspension seat features an interwoven nylon strap frame with attached side, back, and base bolstering, in addition to a cover comprised of high strength nylon rip-stop fabric.

CONCLUSION

As detailed in this report, the design of the 2016 NMSU Baja vehicle was highly goal and analysis driven, with a careful consideration for subsystem integration and overall manufacturability. The top-down design approach follows a logical progression, ultimately leading to a high performance solution, given the constraints and parameters that were evaluated.

ACKNOWLEDGMENTS

We would like to thank our advisor Ken Ruble for his assistance and guidance over the course of the project.

REFERENCES

1. Juvinall, R. C. and Marshek, K. M., 2012,

Fundamentals of Machine Component Design, John

Wiley & Sons, NY

CONTACT

Please contact Gregory Taylor with any questions or comments at gtaylor3@nmsu.edu, Thank you.

APPENDIX A: EQUATIONS

Equation 1

(𝑀𝑐𝑎𝑟 + ∑𝐼𝑖

𝑟𝑖2

𝑛𝑖=1 )

𝑑𝑣

𝑑𝑡=

{

𝑇(𝜔𝑒𝑛𝑔)𝑟𝑐𝑣𝑡(𝜔𝑒𝑛𝑔)𝑟𝑡𝑟𝑎𝑛𝑠𝜂𝑡𝑜𝑡𝑎𝑙

�̅�𝑤ℎ𝑒𝑒𝑙− 𝐷𝑟𝑟 −

1

2𝜌𝑎𝑖𝑟𝐴𝑓𝑖𝑟𝑒𝑤𝑎𝑙𝑙𝑐𝑑𝑣

2 −𝑀𝑐𝑎𝑟𝑔𝑠𝑖𝑛(𝜃),𝑣𝑟𝑐𝑣𝑡(𝜔𝑒𝑛𝑔)𝑟𝑡𝑟𝑎𝑛𝑠

𝑟𝑤ℎ𝑒𝑒𝑙< 𝜔𝑔𝑜𝑣

0,𝑣𝑟𝑐𝑣𝑡(𝜔𝑒𝑛𝑔)𝑟𝑡𝑟𝑎𝑛𝑠

𝑟𝑤ℎ𝑒𝑒𝑙= 𝜔𝑔𝑜𝑣

−𝐷𝑟𝑟 −1

2𝜌𝑎𝑖𝑟𝐴𝑓𝑖𝑟𝑒𝑤𝑎𝑙𝑙𝑐𝑑𝑣

2 −𝑀𝑐𝑎𝑟𝑔𝑠𝑖𝑛(𝜃),𝑣𝑟𝑐𝑣𝑡(𝜔𝑒𝑛𝑔)𝑟𝑡𝑟𝑎𝑛𝑠

𝑟𝑤ℎ𝑒𝑒𝑙> 𝜔𝑔𝑜𝑣

APPENDIX B: FIGURES

Figure 1 Figure 2 Figure 3

Figure 4 Figure 5

Figure 6