Baja Design Report 2010

Delhi Technological University SAE Mini-Baja Design Report

Ashish Sharma, Ayush Goyal, Aditya Krishna, Ankit Goila, Aadityeshwar Singhdeo, Rishabh Bhargava Mechanical Engineering

Vijay Gautam Faculty Advisor

ABSTRACT The objectives of the mini-baja competition are to design and manufacture a “fun to drive”, versatile, safe, durable, and high performance off road vehicle. Team members must ensure that the vehicle satisfies the limits of set rules, while also to generating financial support for the project, and managing their educational responsibilities. This vehicle must be capable of negotiating the most extreme terrain with confidence and ease. The 2010 Mini-Baja Team worked towards these objectives by dividing the vehicle into its major component subsystems. Each team member was responsible for a specific system that was designed according to the objectives and given rules. By examining the 2009 entry, the team was able improve on many design features to better meet the stated requirements.

intended for production and eventual sale to the nonprofessional weekend racer. The main objective of the competition is to subject students to real-world engineering design projects and their associated challenges. The team consists of eleven undergraduate students in Mechanical Engineering. For 2010, the team’s design objectives were structured around not only designing a competitive vehicle for the competition, but also producing a practical vehicle targeting diverse customer markets. The car should appeal to the off-road weekend racer, outdoor sports enthusiasts, adventurers, cottagers, and drivers of every skill level. The team employed engineering design and judgment to refine many of the successful design features of the 2004 entry, as well as develop several new subsystems that further improve the vehicle and meet the objectives. These core systems include:      

Continuously variable transmission (CVT), in series with a two stage chain sprocket reduction. Two front and a central rear disc brakes Rack and pinion steering Rigid and lightweight chassis Four wheel independent double wishbone suspension with rising rate spring setup and variable damping rates Vacuum molded body panels and mud guards

The design of each of these components is further detailed in the following report.

CHASSIS INTRODUCTION Mini-Baja is an international collegiate design competition sponsored by the Society of Automotive Engineers (SAE) that attracts engineering student teams from all over the world. Each team’s goal is to design, build, test, promote, and compete with a prototype of a single seat off-road vehicle

The following is list of design requirements the Mini-Baja chassis must satisfy while meeting or exceeding SAE specifications:  Provide full protection of the driver, by obtaining required strength and

torsion rigidity, while reducing weight through diligent tubing selection.  Utilization of nodal geometry to minimize internal bending moments in frame members.  Design for manufacturability, as well as cost reduction, to ensure both material and manufacturing costs are competitive with other SAE vehicles  Improve driver comfort by providing more lateral space and leg room in the driver compartment.  Maintain ease of serviceability by ensuring that frame members do not interfere with other subsystems.  Modifications in front & rear chassis acc. to the suspension geometry.  Increased legroom as compared to last year.  Modification in rear chassis as we are changing our drive train this time. The new frame design began by using our previous vehicle frames as references and the restrictions in the rule book. A 3-D frame was created using Solidworks modeling software. This technique allowed fast design modifications to be made as required. We did the analysis and the optimization using the Ansys9.0 and Cosmosworks . Consideration taken during the frame design was the placement of the engine, gearbox and driver to assure a low center of gravity. This not only helps with the handling of the vehicle, but also gives much more stability and fewer anti-features.

assumed a front impact for 0.28s on our vehicle with 60kmph and the mass of the vehicle and driver system is 300kg).During the analysis we applied a force at the front nodes joining the FBM and the SIM and constrained the RRH with all Degree of freedom ,we performed the optimization by taking the element type as pipe16 and setting the outside diameter and wall thickness as the design variable(independent variables that directly affect the design variable) ,maximum stress is the state variable (dependent variable that changes as a result of changing DVs), and summation of element volumes is our objective ,after 30 iterations the variables converges to the optimum values which gives the outside dia. as 32mm, wall thickness as 1.7091mm and maximum stress as 293.0975 N/mm2, with the element volume equal to 152872.332mm3. A graph is plotted with x axis as the maximum stress and the y axis as the value of iteration, green line shows the outside dia. and the purple shows the wall thickness. This is converging when the outside diameter is 32 mm. this again satisfies the material specification of our roll cage material whose OD is 31.75 mm (1.25 inch) & thickness is 1.7 mm.

Performance This year a space frame has been used and constructed of alloy steel seamless tubing with plain end to create a strong, light, and durable frame. Also the material should be within the rules.

Material

Outside dia.(mm)

Wall thickness (mm)

Bending Strength

3.05

Required Yield strength (Sy) 365

M.S. 1018 ASME SA179(2 001)

25.4 31.75

1.7

341.96

6363328.6

6363328.6

Thus ASME SA 179(2001) was chosen as compared to

M.S. 1018 Design Optimization The main objective behind the new design was weight reduction, while maintaining safety standards. One method of weight reduction was to keep the vehicle as small as possible, which inherently decreases the size of the cockpit. The second and more effective method is to optimize the frame for maximum load condition (which is during the front impact) on the frame. We calculated that during the front impact value of force acting should be 35000N or 10G (as we

Design optimization curve For further weight reduction, thinner firewall and skid plate materials were used. The firewall material was chosen to be 0.025” utility grade aluminum. The skid plate was optimized for weight reduction by using a combination of sheet metals. On the front nose 0.06” stainless steel was used where additional strength is required to protect the driver in the event of a front impact. The rest of the skid plate is made of much lighter 0.032” utility grade aluminum where strength is not such a necessity. These materials resulted in weight savings of 20% over the 2009 vehicle’s shielding materials. Following are the nodal analysis curve for the front impact with the loads and constraints as discussed earlier.

  

Control lateral load transfer distribution to influence both steady state and limit of adhesion oversteer/understeer handling characteristics The suspension must provide enough wheel travel to dampen the impacts imposed on the vehicle. Since the differential has been eliminated the rear track has bee reduced to 60 “. Also the wheelbase has been shortened by 4”.

The non-professional weekend off road enthusiast requires a vehicle which exhibits both safe, stable, responsive handling; and a soft, comfortable ride.

Von-Mises Stress analysis (analysis shows maximum stress = 582.917 N/mm2)

CHANGES IN DESIGN OVER PREVIOUS YEAR   

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

Deformation Curve

 

Distance between the ball joints was increased to reduce reaction forces Scrub was minimized at both front and rear to improve straight line stability Negative castor of 7.5o and a kingpin of 7.6o was given to improve straight line stability, good turn in response and lower lift The neutral roll axis is brought closer to the CG height by raising the roll centre heights to minimize roll and lateral load transfers Adjustable a arm mountings were provided to give variable roll centre heights to add to the dynamism of the suspension The ground clearance was optimized to about 12” Use of custom shockers to improve wheel travel

SUSPENSION & VEHICLE DYNAMICS DESIGN CONSIDERATIONS (FRONT AND REAR) A Mini-Baja suspension system must satisfy the following design requirements: 

Control movement at the wheels during vertical suspension travel and steering, both of which influence handling and stability

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Provide sufficient sprung mass vibration isolation to maintain satisfactory ride quality, while maintaining high tire-ground contact rate and low tire vertical load fluctuation rate to improve road holding and handling

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Limit chassis roll during cornering to prevent roll-over, decrease roll camber, and therefore, decrease steering reaction time

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Prevent excessively high jacking forces by managing static roll center location and roll center migration Limit lateral tire scrub to maintain straight line stability and minimize horsepower losses at the rear suspension

FRONT SUSPENSION DESIGN Independent double wishbone suspension linkage configuration was used at the front end of the vehicle. Independent suspensions are preferable in the case of rough terrain because they provide better resistance to steering vibrations and reduce unsprung mass. Further advantages of

the double wishbone setup include easy control of the roll centers by choice of the geometry of the control arms, the ability to control track and camber change with jounce and rebound, larger suspension deflections, and greater roll stiffness for a given suspension vertical rate. Hence the design was chosen as independent short long control arm arrangement. The suspension was designed using Wingeo3 Demo Version & SolidWorks. A higher kingpin angle results in lower scrub and higher steering returnability at low speeds. But a higher kingpin also increases the wheel lift which tends to destabilize the car. Hence a moderate kingpin angle of 7.6 degrees was chosen and the scrub radius so obtained was accepted.

Due to inboard disc brakes, the rear scrub radius is minimized to zero which considerably improves straight line stability and also minimizes power losses. A front toe-out of 1 degree was chosen as a trade-off between straight line instability and quicker steering response. Negative caster in front provides low steering into the corner, easy steering out of the corner and a more responsive steering. The loss in straight line stability due to toe out was partially compensated by providing a negative caster of 7.5 degrees n the front. Suspension should provide a slight camber angle in the direction of rotation for optimized performance. Hence a negative camber of 3 degrees was set in the rear.

SUSPENSION COMPONENT ANALYSIS

FEA OF LOWER A ARM

SLA wishbones were used to design the 2010 vehicle. 1”OD and 14 gauge MS 1018 pipes were used for the UCA whereas 1”OD and 12 gauge pipes were used to design the LCA of the suspension. This resulted in 14% weight reduction from last year’s where 12 gauge pipes were used for both LCA & UCA. FEA of the a arm resulted in a FOS of 6 for the design. Gusset plates are used to reduce bending stress and increase torsional rigidity. The front suspension UCA inboard mounted rod end joints used in 2009 were replaced with a single, bushed pivot joint. Two inboard mounted pivot joints were used to mount each LCA to the chassis. Pivot joints are stronger than rod end bearings, and thus, are more reliable in high impact bending load situations, as experienced by the front suspension of a mini-baja vehicle. Threaded ball joints were used at the uprights to facilitate static camber adjustment and increase strength. Also, Greater the effective distance between transverse links, smaller will be the forces in the suspension control arms and their mountings. Hence, it has been increased from last year’s value of 5.5” to 7.5” this year. But the distance is restricted because of the constraints provided by the rim size and the increasing scrub radius As the body tilts it produces a change in its ground height between inside and outside wheels. By careful design the suspension geometry was made to alter the tracking direction to provide oversteering effect. This system is not adopted on the front suspension to avoid the interference with steering geometry. The rear suspension is designed based on the semi trailing arm to provide anti-dive and squat features .this is obtained by converging the swing axis of the double wishbone suspension at some point along the wheel base.

FEA OF UPRIGHT The front uprights were redesigned to incorporate the changes in suspension geometry namely kingpin and caster angle, which also resulted in 7 % weight reduction as compared to last year’s model. The new upright, made of 6061 Al alloy has improved strength, serviceability; factor of safety being above 2.5 in all cases.

FEA OF HUB The front hub was redesigned to meet the demands a more compact wheel assembly. It is also made of 6061 Al alloy as compared to that of mild steel last year, resulting in weight reduction of 63%. The strength was not sacrificed since the analysis showed FOS of 3.

to be the best but also small. With these restrictions in mind we have decided on a rack and pinion style steering system. The rack and pinion style has many benefits. First, the rack and pinion isn’t sloppy at the center point and gives the drive a large range of motion. Second, the rack and pinion provides a large degree of feedback and allows the driver to feel the ground. Third, the rack and pinion places the pivot points of the steering system near the pivot points of the suspension system which greatly reducing bump steer. Finally, the rack and pinion unit is very compact and fits more easily into the front frame. Studying the past year’s designs we have notice a lack of space in the cockpit down by the driver’s feet. This lack of space causes this area to be widened to allow the pedals, driver’s feet and steering column which in turn makes the A arm shorter to keep within the width restrictions.

SHOCKS In order to determine the correct spring rate for the shocks, availability and constraining parameters were considered. The weight of the vehicle and driver was estimated to be 600 lb. The weight distribution for the car was Intron 4402 used to test the spring stiffness

estimated to be approximately 60/40 from the rear to the front. Using the total weight of the car and the weight distribution, the weight on each of the front tires was determined to be 115 lb.The spring stiffness of last year front shockers spring was 15-20 N/mm(calculated using UTM instron 4482), thus this time lower spring stiffness will use for larger travel . The weight on each of the rear tires was evaluated to be 190 lb. The desired static ride height is 13 in. This allows for a wheel travel of at least 8” for the front tires and 6” for the rear tires. The shocks used in last year’s car were analyzed after the event and it was found that the dampers had given in and also the car did not provide the desired travel. This year the shock are being custom made with requisite spring stiffness and matching dampers. For increasing the travel the total length of the shocks was increased to 20”(between mountings). The spring is mounted perpendicular to the plane of the aarms so as to eliminate any lateral forces on the damper so that it functions with minimum friction. The lower mounting will be kept nearer to the upright end of a-arm as compared to the previous year’s mounting. This should increase the maximum spring deflection.

STEERING SYSTEM The steering system is a vital and crucial component in any vehicle. It was decided early that the steering system needed

STEERING SYSTEM Rack specifications Length of rack(in.) Rack diameter (in.) Pinion specifications No. of teeth of pinion Pressure angle Module taken Pitch Diameter Addendum diameter(in.) Dedendum diameter(in.) Geometry Steering Ratio C-factor of rack and pinion(in) Steering arm length(in.) Steering arm initial angle

12 0.875 16 20 1.75 1.102362 1.240157 0.942933 5:01 4 3.18 66.86 deg

The overall steering ratio is the ratio of the steering wheel angle to the average tire angle. An overall steering ratio of approximately 4:1 might be desired which would provide a 45° steer angle of the tires with a 180° input to the steering wheel, eliminating the need for hand-over-hand maneuvers. We have decided to set a steering ratio or 5:1 which would provide a 36° steer angle of the tires with a 180° input to the steering wheel. This allows us to increase the steering arm length to decrease the effort.

BRAKES DESIGN CONSIDERATIONS

The purpose of the braking system is to increase the safety and maneuverability of the vehicle by locking all four wheels in a time of less than 1/3 of a second.

The first option the team explored was the use of drum brakes. A drum brake system consists of a rotating cylinder which is stopped through two brake shoes. The brake shoes surround the outer surface of the rotating cylinder. The braking system consists of a pedal mechanism which applies a force to the brake shoes through the use of a cable. Through research of previous braking designs, it was determined the use of a braking cable arrangement is not feasible. This is due to the lack of area on the vehicle’s chassis. Another disadvantage of the drum brakes is encountered in an offroad application. Since the vehicle will be traveling through mud, there is a high possibility of mud and debris to gather in the space between the shoe and the drum. The other consideration is the disc braking system. Braking with this system can be obtained both mechanically and hydraulically. However, the same problems of the drum brake occur with a mechanical disc brake system. A hydraulic disc braking system uses fluid displacement to engage the brake calipers on the rotor. This is a more ideal system because there is no need for a mechanical system. Therefore, available area along the vehicle’s chassis is not required. Another advantage is mud build up is no longer a limiting factor. ANALYSIS From analysis of the different systems, the team finalized hydraulic disc braking system, which locks all four wheels by using a brake for each wheel on the front and a single brake for the rear axle. The calipers are powered by dual master cylinders. Both master cylinders are standard 3/4 in. bore direct mounted Master Cylinders with 4 oz. reservoirs. Two master cylinders are used to increase safety by incorporating dual redundancy as well as for SAE rules compliance. Another reason for choosing this design was its component arrangement. The master cylinders mounts directly to the custom made brake pedal and are located above the driver’s feet, allowing the driver to easily enter and exit the car.

Braking Analysis Parameters Parameter Front disk O.D. Pedal ratio Laden weight Height of C.G. Coefficient of friction Front weight bias percentage Front master cylinder bore

Value 140mm 5:1 260 kg 508mm .55 33% 19.05mm

The vehicle was required to stop within a span of 5m moving with initial speed of 48 kmph. Ui=42 kmph or 11.5 m/s, Vi=0 m/s, S=5m,a=-1.34 g After calculations it was found out that to have deceleration Torque to be generated by front brakes =384.806 N/m Torque to be generated by rear brakes = 95.522 N/m Force to be generated by front caliper = 3498.24 N Force to be generated by rear caliper = 796.26 N

We did FEA of the discs using ANSYS 9.0. The maximum stress for the front discs came out to be 1678 N/mm^2 which was well within the yield strength of material chosen for the discs And has a factor of safety above 2.

DESIGN MODIFICATIONS FROM LAST YEAR The manufacturer requires a system that is easy to install and easier to bleed. The team considered using 2 front wheel disk brakes and a simple system of using a brake on the final drive. The reason for using an onboard rear disc was that it reduced the unsprung mass of the vehicle and is easy to install and easier to bleed. This system was chosen because of its reliability and safety. Since we had a leak in the brake line last year so this time the brake lines utilized all steel construction with a combination of hard line and flexible steel line. To make efficient use of limited cockpit space, and increase leg room, a reverse actuated master cylinder pedal assembly with custom aluminum pedals and compact remote reservoirs was chosen for this year’s vehicle. This assembly allows the pedals to be mounted as close to the front of the frame as possible, while protecting the master cylinders, and utilizing virtually the entire length of the cockpit. The compact master cylinders with remote reservoirs keep the pedals as low as possible for improved foot space.

DRIVETRAIN OBJECTIVES The purpose of the drive train is to transmit shaft power and torque of the Lombardini engine to the rear wheels of the car. The 10 HP engine produces 23.6 Nm of torque at 2600 rpm. The objective for the Mini Baja competition is to optimize the power delivered to the wheels regardless of the vehicle speed for the various competition conditions. High speed was desired for the acceleration and speed trials while high torque was preferred for towing and hill climbing events. These characteristics are achieved without continuous shifting by coupling a continuously variable transmission (CVT) to the engine. Unlike last year’s manual

transmission , the CVT allows any driver, of any skill level, to focus on the obstacles ahead without concentrating on proper gear selection. The CVT eliminates the burden of a manual clutch and is far more simplistic than an automatic transmission.

Graph comparing vehicle velocity behavior in case of a CVT and a manual transmission

DESIGN The drive train design consists of a 10 HP engine, CVT and a 2 step chain sprocket system providing the final gear reduction. The Comet 790 Series CVT was selected because of its large range of gear ratios.

Theoretical speed to be considered(Vc) =60*1.25 = 75 kmph =20.83 m/s Wheel radius (Rw) =11 inches = 0.2794 m Therefore, Theoretical angular velocity of wheels (Wm) = (Vc/ Rw) = 20.83/0.279= 74.56 rad/s Angular velocity at which jack shaft rotates (Ws) = 2*3.14*Rs/0.54*60 = 6.28*3600/0.54*60 = 697.7 rad/s Final gearing ratio = Ws/Wm = 697.7/74.56 = 9(approx) TRANSMISSION CHAIN SELECTION Chain category…………. Precision roller type No. of strands in the chain ………………104 Pitch of the chain ………………… 12.7 mm Width between inner plates…………7.85 mm Roller diameter ………………….8.51 mm Transverse pitch …………………….13.92 mm Average weight …………………….. 0.70 kg/m Tensile strength ……………………….1835 kgf LIMITING FACTORS Max engine power = 10 hp Angular speed of drive shaft (Wd) =(3600* 2* π) /60 rad/s= 376.99 radian/s Final gearing ratio (r) = 9 CVT high gear ratio = 0.54:1 CVT low gear ratio = 3.38:1 Max angular speed of the rear sprocket (Ws) = Wd/Rt = (376.99/9*0.54)=77.57 rad/s (where Rt is the total gearing ratio=r*CVT ratio)

As the engine reaches its governed rpm limit, 3600 rpm, the gear reduction across the CVT has been determined to be 0.54:1 thus serving as an “overdrive” for the car. At low engine speeds the CVT produces a reduction of 3.38:1 providing necessary torque. The final gear reduction for the drive train comes from the sprocket sizes on the CVT output shaft and the driven sprocket located on the axle. CALCULATION OF FINAL GEARING RATIO Maximum Power of engine =10 hp Rpm delivered at this power by the engine (Rs) =3600 rpm Actual Speed = 60 kmph = 16.67 m/s If we consider 75% efficiency of the drivetrain , then

Max torque on the rear sprocket (T) = Power / Wh (Wh is the angular speed in case of CVT low gear ratio , corresponding to max torque 0.75*(10*746) Watt/ (118.34*6.28/60) = 451.5 Nm Radius of the rear sprocket (R) = 97.1 mm Force acting on the chain mounted on sprocket 4 = T/R (451.5 / 97.1) *1000 N = 4649.8 N Tensile strength of the chain = 1835 kgf = 18350 N Factor of safety = tensile strength/force acting on chain sprocket = 4 Thus, considerable factor of safety has been taken so that even in worst conditions the system remains reliable. After much calculations , it was decided to use an initial reduction of 3 with smaller sprocket having 12 teeth and the larger sprocket 36 . The final reduction would be 3 with the smaller sprocket having 16 teeth and the final output sprocket having 48 thus giving a combined final reduction of 9

SPROCKET

PITCH DIA (MM)

MAX RPM

MAX TORQ (NM)

FORCE (N)

1

49.07

6666.67

50.16

2044.4

2

145.72

2222.2

150.5

2065.6

3

65.1

2222.2

150.5

4623.6

4

194.2

740.74

451.5

4649.8

ANALYSIS FEA of the largest sprocket was done by applying the maximum torque on a side of a single tooth of the sprocket and keeping the two surfaces of the key fixed , one to the shaft and the other to the sprocket . The material chosen was cast stainless steel , giving a calculated stress value of 5.056e+008 . Thus the factor of safety for sprocket was calculated to be 2.5 .

FEA of sprocket 4

DRIVER ERGONOMICS The DTU Mini Baja Team 2010 designed its vehicle keeping 6 categories in mind with respect to the occasional weekend off road enthusiast: safety, performance, durability, comfort, cost and serviceability. Changes in chassis over last year  LFS has been inclined at an angle of 5 degrees so as to decrease the component of force along the vehicle and hence the driver during front impact  The width of the SIM has been increased so as to prevent any part of the driver from protruding out  The distance between the FBM has been increased significantly from last year’s design to improve visibility  Heat shield along with the firewall is provided to reduce conduction of heat and also absorb vibrations  The RRH is at an angle of 5 degrees with the vertical to ensure comfort of the driver and also enhance visibility. Enough overhead clearance is provided to ensure swift movement of driver in and out of the vehicle without comprising on safety.

CONCLUSION Designing a “fun to drive”, versatile, practical and attractive vehicle that targets diverse customer markets is the goal of every design team. The Delhi Technological University’s 2010 SAE BAJA ASIA entry has been successfully designed according to the stated objectives. Since the design process is never ending, the design and modification will continue well beyond the competition. Experience gained from the competition and testing will highlight areas that require design improvements.

VEHICLE SPECIFICATION