Design Analysis And Optimization Of All-terrain Vehicle(ATV) CHAPTER N0. 1 INTRODUCTION We approached our design by considering all possible alternatives for a system & modeling them in CAD software like CATIA, AutoCAD etc. to obtain a model with maximum geometric details. The models were then subjected to analysis using Analysis Work Bench 14 software. Based on analysis results, the model was modified and retested and a final design was frozen. Dynamics analysis was done in Lotus suspension analysis software. The aim was to optimize suspension variables to improve maneuverability. Theoretical calculations of performance characteristics were also done. Extensive weight reduction techniques were followed at every stage of the design to improve performance without sacrificing structural integrity. Page 1 Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. 2 DESIGN CRITERIA FOR THE VEHICLE & METHODOLOGY: NO. CRITERION PRIORITY REQUIRMENTS :- Low Weight Vehicle. 1 Reliability Essential Better Economy. Better Comfort And Durability. 2 Ease of Design Essential DESIGN AND CAD WORK :- 3 Performance High Collection Of Data And Calculation. CAD And CEA Work of the 4 Serviceability High Subsystems. REVIEW AND IMPLEMENTATION :- 5 Manufacturability High Design Review And Project Plan. Maintaining Quality in Fabrication. 6 Health and Safety High Follow up And Project Plan. 7 Lightweight High DFMEA AND VALIDATION :- Maintain DFMEA And DVP. 8 Economic/Low Desired Validate of The Vehicle For Designed Aspect. Cost TESTING :- 9 Easy Operation Desired Testing The Vehicle For All the Terrains. 10 Aesthetically Desired Expecting Failures And Correcting Pleasing Them. TABLE NO.- 2.1 As shown in above table, special considerations were given to safety of the occupants, ease of manufacturing, cost, quality, weight, and overall attractiveness. Other design factors included durability and maintainability of the frame. Page 2 Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. 3 ROLL CAGE : The purpose of the Roll cage is to provide a safe environment for the occupant while supporting other vehicle systems. Several steps were taken to ensure this objective was met. For the frame design, we focused on a lightweight and safe frame that still meets all of the requirements set forth by SAE. Special considerations were given to safety of the occupants, ease of manufacturing, cost, quality, weight, and overall attractiveness. Other design factors included durability and maintainability of the frame. The frame design incorporated bends instead of miters in many of the structural members, believing that this allowed for faster construction, and increased material strength from cold working resulting in an overall increase in product quality. Although there was added cost associated with out-sourcing tube bending, this cost was offset by a reduction in fabrication man hours through decreasing the amount of mitered and welded joints and eliminating man hours and material needed to fabricate fixtures for fit-up ,The Roll cage consists of two main criterions as follows: 3.1 MATEARIL SELECTION: The materials used in the cage must meet certain requirements of geometry and minimum strength requirements found in SAE. Since the frame is being used in a racing vehicle rather than a recreation vehicle, weight and cost is a very large factor in the shape and size of the frame. The proper balance of strength, weight and cost is crucial for the team’s overall success. Page 3 After running all five analyses it was found that there is a need of additional member. durability and weight of Chassis Analysis Work Bench 14 was used to analyze the chassis for all six loading condition. After having added these members.1 mm for the O.5 210 4130 2 AISI 394. condition The six analysis tests conditions are Front Impact.1 GRAPH NO NO. proper analysis was done in the ANSYS Workbench which is tabulated as follows: follows Page 4 . For satisfying the bending stiffness criteria and bending strength the thickness of the pipe was decided to be 2. From the above tables. Rollover Impact. a second analysis using identical loading constraints was completed and results of these tests are shown in table table.2 FINE ELEMENT ANALYSIS: In order to optimized the strength. I and Torsional ansys heave and the loading on the frame from the front and rear shocks.. selection depended mainly on the cost and availability of the material.3.4 mm with the thickness of 2.7 36.7 294. we concluded that AISI 1018 was best suitable for the roll cage with economical cost and easier availability.3. O.D. was selected as 25.D. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) Materia UTS UYS Elongation Young’s l modulus Mpa Mpa % Gpa 1 AISI 560 450 21.1 In addition to the above table.1 mm 3. Side Impact. of 28 mm for the primary members of the chassis and for the secondary members. for confirming the safety of the roll cage.5 200 1020 3 AISI 440 385 15 205 1018 TABLE NO NO. . 1.88 3 FRONT - analysis 2.74 1.2 4 Torsional 1.64 4.2 3 Roll Over Impact 6.84(F) 3.09 1.49 3.3 RESULTS: DETAILS MAX STRESS MAX FOS DEFORMATION (Mpa) (mm) 1 Front impact 385.3.26 TABLE NO.2 3 Roll Over Impact 272.67 1 2 Side Impact 303..3 4 Torsional ansys .4 2 0.2 DETAILS MAX MAX TIME OF FORCE FORCE IMPACT (kN) (in terms of g’s) (s) 1 Front impact 30 10 0.2 2 Side Impact 9 3 0. Design Analysis And Optimization Of All-terrain Vehicle (ATV) (Assuming the total weight of the vehicle is 320 kg) Setting up the analysis: Ultimate Yield Modulus of Percentage Tensile Hardness Component Material Strength Elasticity Elongation Strength (BHN) (MPa) (GPa) (%) (MPa) Roll Cage 1018 steel 450 380 265 16 130 Hub 6082 Al alloy 225 186 70 12 75 Adapter EN8 660 530 206 7 120 TABLE NO..64(R) 1.3.02 1.82 3 REAR - TABLE NO.4 Page 5 .3. 1 FRONT IMPACT IMPACT: (8-10G) Page 6 .3. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 3. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 3.2 SIDE IMPACT:(3G) IMPACT Page 7 .3. 3. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 3.3 ROLL OVER:(2G) (2G) Page 8 . 3. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 3.4 TORSTIONAL ANSYS:(3G ANSYS (3G FOR FRONT AND REAR) Page 9 . arms were selected for the front suspension. Compatibility with Desired It must be compatible because Steering suspension geometry is linked with steering geometry. 10” of travel High To ensure ground contact always. LOTUS SHARK Page 10 . and from that design considerations had to be decided Consideration Priority Reason Simplicity Essential Main objective Lightweight Essential Lower weight means Faster car. toe in and out for improving handling. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. The double a-arm suspension is the most feasible design according to our design. 4 SUSPENSIONS: 4. From the above considerations to balance weight and cost savings for the manufacturers. Adjustable Desired To adjust camber. several options for front suspensions were analyzed. Two possibilities for the front suspension were a double a-arm and a single arm McPherson Strut suspension. To do this the operating conditions of the competition had to be researched. Durability High It should be durable and reliable for any condition. thus double A. and comfort and handling for the customer.1 FRONT SUSPENSION: The problem that was encountered was to design a competitive front suspension for the ATV . Shock Absorbing Desired High shocks in the front. To design the front suspension several software packages were utilized to ensure the best possible results. For the best handling characteristics the front wheels must always be in perpendicular contact with the ground. Bump steer and camber gain must be minimized in both ride and roll changes. Fig. No.4. No.1 :.1 Page 11 .2 :.SIMULATION OUTPUT AND ROLL CENTER Fig.4.4.FRONT SUSPENTION ASSEMBLY EXPLODED VIEW The front suspension arms were designed to be as easy to manufacture as possible. while maintaining the high strength as desired. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) simulation software was used to create simulations for both parallel and opposite wheel travel. The Build quality was maintained by welding the A arms mounting brackets at the designed hard points within a tolerance of 1 mm The variation of the toe angle and camber with respect to bump as obtained from Lotus shark GRAPH NO. Lightweight Essential Lower weight means Faster car. 8” of travel High To ensure ground contact always.4.2 REAR SUSPENSION: There were many objectives and considerations to look at during the process of designing and building the rear suspension. This enables the drive train to be pulled from the car for maintenance. The Fox Float loat air shocks have 6 inches of travel. The rear suspension is a full trailing arm design with only one arm per side. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 4. near the end of the arm. Durability High Withstand abusive driving during the endurance race. The trailing arms allow for the full drive train assembly to be removed without interference by the suspension. The rear suspension geometry and modeling was done in Catia and it is as shown below: Fig. and are mounted near the bearing carrier. This allows for maximum suspension travel while staying within the range of the rear axle CV joint travel.REAR SUSPENTION ASSEMBLY EXPLODED VIEW Page 12 . and keeps the overall design of the rear of the car simple Consideration Priority Reason Simplicity Essential Easier to fix. Another reason that trailing arms were used was that the drive train design was to be modular. No.3 :. design. and about half way up the rear main roll hoop. analyze. build. 4 mm Roll Angle 172⁰ Rebound 39.4 mm 117..14 mm 39.3 FOX RACING SHOCKS: Right from the beginning we focused on reducing the weight of the vehicle. The customized Spring and damper assembly of the vehicle was way too bulky to be used in ATV. the fox shocks were selected on the following criteria: Travel of the Shocks. Design Analysis And Optimization Of All-terrain Vehicle (ATV) The car has been driving for two weeks.8 inches of extended length.77kg Camber 1. which is perfect from our design point of view.85⁰ 0⁰ Gain TABLE NO. It provides 6 inches of travel and 19. FOX FLOAT 2 air shocks were selected and procured. Thus. thus team emphasized on Fox Shocks which reduced the weight of the vehicle to a large extent and provided easy adjustable stiffness to the shocks. Cost and availability.14 mm Turning Radius 5m Weight Transfer 90. From the market survey.294 N/mm 19. there has been testing done to see if the suspension reacts the way intended by design. Page 13 .1 4.4. Total extended Length of the shocks. Parameter Values PARAMETERS VALUES Front Rear Caster 6” Suspension Suspension Kingpin inclination 14⁰ Wheel 254 mm 206 mm Static Camber Set as Zero Travel Static Toe In Set as Zero Wheel Rate 9.90 N/mm Roll angle @Speed 30 km/hr Jounce 117. It turns out that the design of the rear suspension is working as well or better than expected. however using a longer moment arm tie rod mount offset this effect. Low Steering Ratio Essential Quick steering response Ackerman geometry High To make understeer. The Ackermann angle was selected by analyzing wheel angles from previous years. Page 14 . The driver had to remove his hand from the wheel at least once to complete the turn. or 70 mm of rack travel per revolution of steering wheel travel. The design considerations are as follows: CONSIDERATION PRIORITY REASON Simple Design Essential Easy to repair during competition Light Weight Essential Lower weight means Faster car. In the normal rack & pinion vehicle the driver had to turn the steering wheel 540º to bring the wheels from the center to lock. 5 STEERING: On the rough terrains it is very essential to have the steering must be light and should give quick response on turns. easier maintenance. Most of the analysis was focused on the steering system. and increasing the steering responsiveness. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. Minimize Bump steer Desired Conserve momentum while Steering Rack and Pinion steering system was selected due to its easy availability.5 to 1.A new system provided a motion ratio of 6. The team also focused on decreasing the amount of steering wheel travel. The primary focus was on decreasing the steering effort. The higher ratio rack has inherently larger steering effort. The goal was accomplished by using a REDUCTION GEARBOX after the pinion . The goal was to allow the driver to use only 290º of steering wheel travel from the center to maximum wheel travel. feasibility to modifications and the cost. STEERING ASSEMBLY Fig.5. No.99 Tie rod Length (mm) 400 TABLE NO.59 Steering Effort (N) 108 Percentage Ackermann 98. No 5.57 Steering Ratio 6..2 :. Steering Rack travel(mm) 57 assembly Steering Wheel lock 109⁰ from centre Turning circle 2. Ackermann geometry Fig. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) The steering calculations are tabulated t as: PARAMETERS VALUES Fig.ACKERMANN GEOMETRY Page 15 .1 :. 5. No.1 Fig.48 Radius(m) Scrub Radius (mm) 36. The driver had to remove his hand from the wheel at least once to complete the turn.5 mm. It is as shown below: below Without Without Part Reduction Reduction gearbox gearbox Steering Ratio 13:1 6.5. The goal for 2014 was to allow the driver to use only 290º of steering wheel travel from the center to maximum wheel travel.1 STEERING RATIO REDUCTION GEAR BOX: In the normally vehicle the driver had to turn the steering wheel 540º to bring the wheels from the center to lock. No. This gearbox consists of two gears: One bigger gear with diameter of 68 mm and the other smaller gear with the diameter of 35.REDUCTION TABLE NO.2 GEARBOX EXPLODED VIEW Page 16 .. The bigger gear is attached to the column of the steering wheel and the smaller gear is attached to the pinion side by the Universal joint. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 5. The goal was accomplished by using a REDUCTION GEARBOX after the pinion.3 :.5:1 Rack travel per revolution 35 mm 70 mm of steering wheel Required Rack travel 70 mm 57 mm (Centre to lock) Rotation of steering wheel 540° 290° (Centre to lock) Steering Effort 68 N 108 N Fig. 5. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. 6 BRAKES: The braking system for the vehicle is responsible for stopping the vehicle at all times and is integral for the driver’s safety. That why the brake must be capable of locking all the four wheels when applied so we incorporated disc brakes in the front and rear. CONSIDERATION PRIORITY REASON Simplicity High Overall goal of vehicle. Light Weight High Lightweight parts to minimize total weight. Performance High Capable of decelerating a 320 kg vehicle. Reliability Essential Reliable to provide hard braking always. Ergonomics Essential Optimal pedal assembly fitment to suit every driver. According to the rim size and the braking calculations we chose to use Bajaj Discover ST discs that will be mounted on the hub in the front. Disc brakes were chosen because of the ace of compatibility, the availability of the replacement parts and the overall effectiveness that the system provides. For the rear design, rear disc brakes of Bajaj Pulsar 220 were used. It provided the required diameter of the disc and the required braking torque could be achieved. The design calculations are tabulated as follows: Page 17 Design Analysis And Optimization Of All-terrain Vehicle (ATV) PARAMETERS FRONT REAR Outer diameter(Custom) 190 218 Effective Rolling 81 95 Radius(mm) Thickness(mm) 3.47 3.47 Material Perlite Grey Cast Iron Radius Of Gyration(mm) 170 280 Moment of Inertia(kg/m^2) 0.289 1.176 Calliper BAJAJ DISCOVER 125 ST Calliper Piston 28 Diameter(mm) Coefficient of friction 0.45 Tandem Master Cylinder Maruti 800 TMC diameter(mm) 19.05 TABLE NO.- 6.1 PARAMETERS VALUES Braking distance(m) 17.66 (Deceleration 0.8kg ) Pedal Force(N) 130 Pedal Ratio 1:4 Inline Fluid Pressure 0.5bar Dynamic load Transfer(kg) 83.63 Single Stop Temp. Rise(⁰c) 22.5 TABLE NO.- 6.2 Page 18 Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. 7 POWERTRAIN: The goal of the drive train is to transfer power from the engine of the vehicle to the wheels. The power transferred must be able to move the vehicle up steep grades and propel it at high speeds on level terrain. Acceleration is also an important characteristic controlled by the drive train. Calculations were done according to the considerations, looking at gear ratios, engine power and wheel size. After the calculations were re verified no reduction is to be given was decided. Hence direct line was given. Also during design, the angle of the propeller shafts was taken care. The drive train for the car has been radically overhauled to improve overall car performance and correct vulnerabilities. The Drive Train Based of Mahindra GIO was used based on the traction and speed calculations. The system benefited with simplicity and low cost. GIO transmission was used in forward configuration, this year to enhance torque the transmission is used in Reverse configuration. It can be tabulated below : GEAR RATIOS Initial Acceler MULTIPLATE Tractive GEARBOX ation CLUTCH Effort G1 G2 G3 G4 R (m/s²) (N) PIAGGIO APE YES 25.52 15.16 9.25 5.96 30.62 1702.8 2.80 PASSENGER MAHINDRA YES 31.48 18.7 11.4 7.35 55.08 2100 3.76 ALFA CHAMPION MAHINDRA ALFA YES 25.52 15.16 9.25 5.96 30.62 1702.08 2.80 PASSENGER TATA NANO NO 27.6 15.6 10.08 6.64 31.42 1841.58 3.14 MAHINDRA GIO YES 27.66 14.86 8.48 5.55 33.66 1845.58 3.15 MAHINDRA GIO YES 33.66 18.08 10.32 6.76 27.66 2245.93 4.11 IN REVERSE FORCE MINIDOR NO 24.42 14.58 8.22 4.8 23.4 1629.40 2.63 PICK UP Page 19 54 1673. 54 km/hr @Top gear Velocity Max.925 ratio Max.7.18 Nm @ First gear Clutch type Multiplate wet type clutch Gearbox type Trans-axle axle Constant mesh gearbox Shifter type sequencial single wire shifter 2 CVJ Connection Stock maruti 800 DRIVELINE Sleeves on drive shaft for length correction Fig. Specification Gear ratio 4.1 :.12 9.66 5.33 7 23.50 2.55 18. 7.1 .1 The transmission was coupled directly to the engine with a adapter. No. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) AUTORICKSHAW YES 23..27 27.98 1571.35 2.081 15. This assembly is explained below.979 Overall gear 4.8 8. Torque 586.49 MAHINDRA NO 25.73 CHAMPION TABLE NO.ADPATER ASSEMBLY EXPLODED VIEW Page 20 . 13mm Factor of safety 2.2 GRAPH NO. Equivalent 193. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) 7.7.7.. Deformation 0.46 mpa Max.1 Page 21 ..74 TABLE NO.46 mpa Stress Max.1 ANALYSIS OF THE ADAPTER: ADAPTER PARAMETERS VALUES Max. Shear Stress 104. 1 :. Tire configuration. For the rear. This provided the required strength with lower weight. The ideal tire has low weight and low internal forces. aluminum sheet of 0. it must have strong traction on various surfaces and be capable of displacing water to provide power while in mud.1 BODY PANELS: To reduce the weight of the vehicle. For the side panels. No. Page 22 . traction is one of the most important aspects of both steering and getting the power to the ground. aluminum sheets were used instead of Mild steel sheets. requirements are better traction and larger diameter. The Front wheel hub was made from Aluminum this year to reduce the weight. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0. 8.5 mm were used. The Rear wheel hub is the OEM part and modifications were made to assemble disc onto the hub. 8 WHEEL ASSEMBLIES AND BODY PANELS: In an all-terrain terrain vehicle. Maruti Alt Altoo bearings were used. thus. Fig.7 mm thickness was used. Therefore. tread depth. In addition. For the firewall aluminum sheet of 0. tires with specifications of 25x10x12 were selected. weight. The 10 10-inch diameter of the rim will allow the brake components to fit inside the wheel. smaller maller diameter tires were used to allow better maneuver control. and rotational of inertia are critical factors when choosing proper tires. thus. For the front.FRONT AND REAR WHEEL ASSEMBLY EXPLODED VIEW 8. These sheets were bolted to the chassis. tires with specifications of 21x7x10 were selected. Bearings were selected according to required design. exciting racing without risking major injury. including arm restraints. Design Design Parameter Parameter Std. as well as maximum optimization of their functions during an emergency. Value Value Value Value Angle at Steering Wheel 110⁰-130⁰ 110° .1 Drivers should be able to experience fast pace. Std. 6. An LED brake light warns other drivers of deceleration. a number of safety features have been added to further reduce the possibility of personal injury.. In addition. 320 elbows Dia. 4 Head Clearance . and two kill switches were all placed for easy access and use. fire extinguisher. A Six-point safety harness keeps the driver adequately restrained.9. Roll cage padding protects driver’s head from impact.5” (inches) TABLE NO. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. The remaining standard safety equipment. (mm) Angle of Steering Angle at knees 120⁰-150⁰ 130° 20⁰-45⁰ 20⁰ Wheel Clearance from vehicle . The ergonomics aspect of the SAE Baja vehicle is crucial in ensuring that the car will both meet all of the rules stated in the SAE rule book as well ensuring that all of the components of the car will function properly when assembled together. A safety helmet and neck support protects the driver. Page 23 . It is essential that each member of the team is able to safely and comfortably operate the vehicle. Car 80 meets or exceeds all of the minimum safety requirements composed by the Society of Automotive Engineers and the event coordinators. 9 ERGONOMICS AND SAFETY: Ergonomics is the science of equipment design intended to maximize productivity by reducing driver fatigue and discomfort. 1 :. 10 ELECTRICALS: The electrical system was proposed to work on many road vehicles. reverse light and kill switch. horn.ELECTRICAL CIRCUIT Page 24 . The electrical circuit for the vehicle is as shown below: Fig. No 10. The electrical circuitry is to be done mainly for the brake light. No. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0. maneuverable. 11 CONCLUSIONS: The Team used extensive physical testing. After initial testing it can be seen that our design should be a strong competitor in this year’s competition. and prototype construction to create a vehicle that is fast. 11. No.PVC MOCKUP ERGONOMICS CONSIDERATION Page 25 . and reliable. Several team members attended the time to time workshops arranged by BAJA to gather her ideas and information about what design choices were successful and how they could be incorporated into our design. Fig. hours of simulation and analysis. No.2 :. There will be extensive testing done to prove the design and durability of all the systems on the car and make any necessary changes up until the leaves for the competition.1 ::. Fig. 11. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0. 12 CAD MODELS: FRONT VIEW SIDE VIEW REAR VIEW TOP VIEW Page 26 . Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0. O. Knuckle Knuckle 7 6 5 210 Soldering of 7 2 2 28 Dislocati Knuckle on Detachm 8 6 4 192 Perfect Shaft 8 2 3 48 ent from length different Power Drive Shaft ial Train Muff 10 8 4 320 Align using V 8 3 3 72 weld block& Drill hole failure in Muff coupling for excess weld material Penetration . 13 DESIGN FAILURE MODE EFFECTIVE (DFME) ANALYSIS: SYSTEM COMPONE POTENT S O D R ACTION TAKEN S O D R NT IAL P P FAILUR N N E MODE Optimum design consideration to Trailing Torsion 7 7 3 147 reduce tensional& 7 4 2 56 Arm bending force. co- Bending linear line of action Suspension of wheel& spring centerline. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0. Leakage 9 4 3 108 Refill of brake oil 9 2 2 36 of oil from fluid line Engine Engine Air Cloggin 9 252 Rerouting of Air 10 60 Intake g of Air 7 7 4 196 intake above the 7 2 3 42 filter 8 224 driver seat through 8 48 the firewall Pedal Failure 10 4 3 120 Replacement of 10 2 2 40 of linkage linkage Page 27 . co-linear of Master line of action the Brake Cylinder pedal & pushrod.F. Steering Pinion Pinion 9 5 2 90 Replacement of 9 2 2 36 Failure Pinion Failure Push rod should of Push 8 3 6 144 have adequate 8 2 4 64 Tandem rod D. 26. Move the vehicle over surface Maximum Transmission having inclination of 100 . 14 DESIGN VALIDATION PLAN Design System Parameter Method Of Checking Value Vehicle is to be taken to a surface Minimum with loose soil and a circle with Turning steering wheel locked at full travel Steering 4. The distance Diameter between the two diametrically opposite points is to be measured. Climbing then allow the vehicle to climb the Page 28 . Damper is mounted in the designated position and point of Jounce- maximum designed travel is 117.14mm travel. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. Load is to be added to the spring Front- Spring while holding it in a vertical block.73N/mm Stiffness Load required to cause unit Rear- deflection is to be noted.4mm Travel marked.9 m Circle is to be negotiated. Difference between the initial and final position of hub is to be noted. 40. Hub is moved upwards Rebound- manually till the point of maximum 39. 33⁰ Gradient Transmission to be set on first gear.10N/mm The vehicle is to be loaded on a jack and front wheels and spring are Suspension removed. Distance is to be measured from the line to the front once the vehicle is brought to a complete halt. The vehicle is to be at a predetermined speed while crossing Stopping that line. A reference line is to made. Design Analysis And Optimization Of All-terrain Vehicle (ATV) Capacity slope.until vehicle would not climb the slope. Previous angle is measured. Tachometer is to be held at 54 km/hr the wheels and maximum reading is noted. Another method is to Weld is seen Roll Cage Weld Test take the welded joint and impact to fail. with multiple hammer blows until failure. Vehicle is to be loaded on a jack. The welded joint is to be taken and tested on a Universal Testing Machine. Page 29 . Transmission is shifted to final gear and full throttle is given for 20 Top speed seconds. Maximum force is to be Brakes 17. Then subsequently increase the slope by 50 .5 m Distance applied on brake pedal when front wheels cross the line. The failure is to be observSZed. from which the driver is to start braking. Failure of weld or material is noted. Speed is calculated by using noted rpm. 15 TECHNICAL SPECIFICATIONS ENGINE Type Briggs & Stratton 10HP OHV Displacement.82 firewall (mm) Z:165. 7. cc 305 Max. 19 @ 2500 Nm @ rpm Max Power. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) CHAPTER N0.r.5 @ 3600 kW @ rpm Transmission Mahindra Gio Gearbox 4 forward 4 reverse speed Steering Front Double Wishbone Rear Trailing Arm Brakes Overall Performance Targets Hydraulic Disc Brakes Dimensions Light Weight Buggy Length (mm) 2286 Width (mm) 1600 Best Driver Safety and Ergonomics Height (mm) 1400 Weight Kerb Weight (Kg) 240 Gross Weight (Kg) 320 Wheel size 20% Front 20% Front (inches) 21x7x10 Left Weight Distribution Right Rear (inches) 25x10x12 Centre of Gravity 30% Rear 30% Position w.t. Torque.33 center of base of Y: Y:-45.26 Page 30 . X :109. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. 16 TESTING: BRAKE TEST JUMP TEST FINAL VEHICLE Page 31 . Design Analysis And Optimization Of All-terrain Vehicle (ATV) CHAPTER N0. Gillespie 2 Windsor 3 Mille ken & Millikenh 4 Automobile Engineering volume 1-volume 2 By Kirpal Singh 5 Google Search 6ARAI India Page 32 . 17 References: 1 Vehicle Dynamics By Thomas D. 2 = 117. the front weight is 80 kg.81 =3139.73 N/mm Wheel rate/ wheel travel = (kw) Wheel rate is the actual rate of a spring acting at the tire contact patch kw = ks × (M.2 N ? K=? (? =allowable travel=117. Design Analysis And Optimization Of All-terrain Vehicle (ATV) APPENDIX – A SUSPENSION DESIGN KINEMATIC ANALYSIS – Sample calculation front suspension- A arm suspension: The weight of the vehicle is 200 kg but because of 40:60 ratio of weight distribution between front and rear suspension.for trailing arm(from internet reference) kw = ks ×(M.R)2……………………….4 = 26.4 mm) 3139.R)2 × sin(θs)…………. F = 80×4×9.for A-arm ( from Windsor as reference) for an offroad vehicle ideal value is 8 to 12 inch unit-N/m or lbs/inch Page 33 . 4 mm ?? ? ?? ?? ? ? ? ? Motion ratio = ? ? ? ? ? ?? ? ? ? ? 152. Design Analysis And Optimization Of All-terrain Vehicle (ATV) CALCULATIONS:- A) Front suspension- A-arm double wishbone- Motion ratio= 0.73 × 0.62 = 9622.4 wheel travel = 0.6 wheel travel = 254 mm = 10 inch Wheel rate (kw) = ks × M.R = d1/d2 = 369 497 = 0.6 Shock travel = 152.6 N/m B) Rear suspension:- Trailing arm suspension- Motion ratio d1= 369 mm d2= 497 mm M.R2 = 26.74 Page 34 . 94 mm Wheel rate (kw) = ks×(M.? Unit – Nm/deg or lbs/inch Front suspension A-arm double wishbone ?? ×? ? kØ = ? ×? ? .74 Wheel travel = 205.4 Wheel travel = 0.482 ×9622. Design Analysis And Optimization Of All-terrain Vehicle (ATV) ?? ? ?? ?? ? ? ? ? Motion ratio = ? ? ? ? ? ?? ? ? ? ? 152. ?? ×? ? Roll stiffness (kØ) = ……………………………(from internet) ? ×? ? .R)2×sin(θs) = 40.34 Nm/deg Page 35 .74)2 ×sin650 = 19.10 × (0.8 = 2 ×57.39 N/m Roll stiffness – (kØ) Amount of roll moment needed to roll the suspension by one unit of rotation guidelines.90 N/mm = 19901.3 = 19.? 0. 39 N/m 0.from internet ? Rebound – it is the downward movement or extension of suspension component.882 ×19901.3 = 134.48 Nm/deg Jounce – It is the upward movement or compression of suspension component.? t = 885 mm = 0. ? ? = ………………….88 m kw = 19901. Design Analysis And Optimization Of All-terrain Vehicle (ATV) Rear suspension Trailing arm suspension ?? ×? ? kØ = ? ×? ? .39 kØ= 2 ×57. Rebound : jounce = 3:1 Calculation:- A) Front suspension- A-arm double wishbone- ? Jounce(at 4g load) ? = ? Page 36 . 42 mm ? Rebound (at 4g load) ? = = 39. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 4×9.73 = 117. ??? N.81×80 = 26.14 mm ? Natural frequency:- The natural frequency is rate at which an object vibrates when it is not disturbed by an outside force.F = ? ??? ??? ? ? ? ?? ???? ? = 1.81×120 Jounce(at 4g load) ? = ? = 40.(from internet) ? ??? ??? ? ? ? ?? ???? ? Ideal value = 1 to 1.14 mm B) Rear suspension:- Trailing arm suspension- ? 4×9.10 = 117.44 mm ? ? ×? .? ? ×? ? Rebound (at 4g load) ? = = ? ? ? .F = √ ?? ?? ? ? ? ? ? ? ?? A) Front suspension- A-arm double wishbone suspension ??? N.45 Hz Page 37 ..5 Hz ? ? ? ? ? ? ? ? ?? N.? ? ×? = 39.F = ………………………………. 73+0.73 ×0. Front suspension- A-arm double wishbone suspension- ? ?. Calculation:- A). rear suspension :- Trailing arm suspension ? ?.10+0. per unit vertical displacement of the sprung mass relative to the ground at a specific load.421 = 0.? ? Ride rate = ? ? +?? 26.51 Page 38 .10 ×0.414 B).? ? Ride rate = ? ? +?? 40.51 = 40.F = √ ?? ?? ? ? ? ? ? ? ?? = 1.421 = 26. Design Analysis And Optimization Of All-terrain Vehicle (ATV) B)Rear suspension- Trailing arm suspension ? ? ? ? ? ? ? ? ?? N.6 Hz Ride rate The change of wheel load at thecentre of tire contact. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) = 0.30 for front suspension Camber gain = jounce + rebound = 117.14 = 156.16” = 2.44 + 39.58 mm = 6. Front suspension: suspension:- Camber gain = 1 inch = 0.503 Camber gain:- The amount of angle change in front spindles as suspension travels inward or outward from the centre of car A.460 Roll centre analysis Page 39 . = Steering Wheel Lock Angle / Road Wheel Angle S. Steering Wheel Centre to lock Angle – 290° 3.48m C. STEERING EFFORT S.R.59 D. STEERING RATIO S. Rack Used :. Design Analysis And Optimization Of All-terrain Vehicle (ATV) APPENDIX – B STEERING DESIGN CALCULATIONS:- A) STEERING TYPE : RACK & PINION 1.TURNING RADIUS (R) 2 = (R1) 2 + (C) R = 2.MARUTI 800 B. Rack Travel: – 57mm 2.E.R = 6. = Weight On Front Wheel/ Moment Ratio Page 40 . PERCENTAGE ACKERMAN % Ackerman = (Angle Of Inner Wheel – Angle Of Outer Wheel) / Angle Inside Wheel For 100% Ackerman % Ackerman = 98. Design Analysis And Optimization Of All-terrain Vehicle (ATV) S.99 Fig.E. = 108 N E . Ackerman Geometry Page 41 .TIE ROD LENGTH – 400mm F. Bump Steer Correction H. steering reduction gear box concept implemented in vehicle. Reduction box Exploded view Page 42 . Fig. keeping tie rod parallel to Upper A-Arm A Arm shown in geometry. Fig. Bump Steer Consideration To minimize bump steer. Design Analysis And Optimization Of All-terrain terrain Vehicle (ATV) G. 5:1 Rack travel per revolution 70 mm of steering wheel Required Rack travel 57 mm (Centre to lock) Rotation of steering wheel 290° (Centre to lock) Steering Effort 108 N Page 43 .Design Analysis And Optimization Of All-terrain Vehicle (ATV) Part Implementation in vehicle Steering Ratio 6. Pedal braking force- =Total input to each TMC =Pedal force*No.52) =819. Pedal Force.4:1 4.. Coefficient of friction =ű = 0.45 E. Design Analysis And Optimization Of All-terrain Vehicle (ATV) APPENDIX – C BRAKE CALCULATIONS:- A.130N B. C.73N =83. TMC. TMC 1.445/1.19. Stopping Distance- = V²/2ά = 17.8)*3500*(0. Piston dia.05mm 3. Dynamic Load Transfer- W =(ά/g)*w*(H/L) =(0.Maruti 800 2.of TMC =130*4 =520 D.63.56kg Page 44 . Pedal ratio. 2+28 ……. =10.25 =48. Actual torque = Ideal torque * Brake Force =39 * 1.23*10^-³) =2244.44 Pitch dia.05²)) = 18. Rolling radius- T = F*R But.75kg.m Torque is divided by 2 wheels.m R= T/F = 24.375/2.75/2 =24.375kg.75N H.86 Actual dia =217.25bar G. 48. Design Analysis And Optimization Of All-terrain Vehicle (ATV) F. (28=calliper dia) =245mm Page 45 .25*10^⁵)*(1. Calliper/Brake force- F =P*A =(18. Inline Pressure- P = F/A = 520/((Π/4)(19. 4=4mm K. Temperature Rise- =22. Front wheel speci. Disc dia.5°c J. Design Analysis And Optimization Of All-terrain Vehicle (ATV) I.83 4. Size -21*7*10 3.- 1.=21. Disc thickness-4mm Page 46 . Type-Disc brake(custom) 2. Front Pistriction-220mm 5. Front wheel speci. Disc thikness-3.- 1. Type-Disc brake(custom) 2. Disc dia.=190mm 4. Front Ristriction-260mm 5. Size -25*10*12 3. 29 m All calculations for 1st gear . taking available gear ratios in consideration.29) = 1702. 1. Design Analysis And Optimization Of All-terrain Vehicle (ATV) APPENDIX – D POWERTRAIN DESIGN CALCULATIONS:- Calculations for Gear Box Selection – Taking following Assumptions: IE = 1 λ = 1.35×25. Piaggio Ape Passenger FzA = Total Available Traction Fzex = Excessive Traction =(FzA-FzB) FzB = Total Driving Resistance=534 FzA = (Engine Torque ×Gear Ratio)÷(1000× rdyn) = (19.8-534 Page 47 .8 N FzEX = (FzA-FzB) = 1702.3 ηtot = 1 rdyn = 0.52)÷(1000×0. 8 N Acceleration =FzEX =mF ×λ×a =a = ( FzEX)÷ (mF ×λ) = 1168.48)÷(1000×0. 1.80 m/s2 Gradient Angle on 1st gear FzEX = mF ×g×sin(ast) =1168.81× sin(ast) = (ast )= 22.29) =2245.8÷(320×1.93 N FzEX =FzA –FzB =1711.04° According to above calculations same procedure for following vehicles.93= mF ×g×sin(ast) = (ast) =33.11 m/s2 Gradient Angle 1711. Design Analysis And Optimization Of All-terrain Vehicle (ATV) = 1168.35×33.3) =a = 2.29) = 2100 N Page 48 .9° 2.35×31.8=320×9. Mahindra Alfa Champion FzA= (19.66)÷(1000×0.Gio In Reverse Calculations FzA=(19.93 N Acceleration =FzEX = mF ×λ×a =a = 4. 47 N Acceleration (a)= 3.58 N Acceleration(a)=3.35 N Page 49 .58 N Acceleration(a)=3.40 N FzEX =1095.58 N FzEX =1307.90° 3.42° 6. Auto Rikshaw FzA=1571.8 N FzEX =1168. Force Minidor Pick Up FzA=1629. Design Analysis And Optimization Of All-terrain Vehicle (ATV) FzEX = 1566.80 m/s2 (ast)=22.14 m/s2 (ast)=24. TATA NANO FzA=184. Mahindra Gio FzA=1845.58N FzEX =1311.76 m/s2 (ast)= 29.8 N Acceleration(a)=2.93° 2.14 m/s2 (ast)=24. Mahindra Alfa Passenger FzA=1702.40 N Acceleration(a)=2.60° 4.69° 5.63 m/s2 (ast)=20. 76 29.80 22.93° 3) Mahindra Alfa Passenger 1702.50 N Acceleration(a)=2. Vehicle Name Initial Acceleration Gradient No. Design Analysis And Optimization Of All-terrain Vehicle (ATV) FzEX =1037.9° 2) Mahindra Alfa Champion 2100 3.73 m/s2 (ast)=29.80° Passenger Transmission used in the vehicle on the basis of Acceleration and Traction Mahindra Gio in Reverse Configuration Page 50 .50 2.11 33.50 N FzEX =1139.28 ° FINAL VALUE FOR TRANSMISSION SURVEY CHART Sr.35 2.60° 5) Mahindra Gio 1845.9° 4) TATA NANO 1841.58 3.14 24.42° 8) Auto Rickshaw 1571.49 19.8 2.14 24.93 4.04° 7) Force Minidor Pick Up 1629.40 2. Tractive Angle Effort 1) Piaggio Ape Passenger 1702.29° 9) Mahindra Champion 1673. Mahindra Champion Passenger FzA=1673.8 2.29° 7.35 N Acceleration(a)= m/s2 (ast)=19.58 3.60° 6) Mahindra Gio in Reverse 2245.73 21.63 20.80 22. 51° Reverse Gear = 24.37 N 3rd Gear = 688.93 N 2nd Gear = FzA2 = (19.59 N 4rt Gear = 451.08)÷(1000×0.38° 2nd Gear= (ast) = 3° 3rd Gear= (ast) = 2.35×32.69° Page 51 . Design Analysis And Optimization Of All-terrain Vehicle (ATV) Traction available on each gear of Mahindra Gio in Reverse Configuration .055 N Gradient on each gear when used in reverse configuration – 1st Gear = (ast) = 33° 2nd Gear= (ast) = 12.29) =2245.66)÷(1000×0. 1st Gear = FzA1 = (19.82 ≈ -1.29) =1206.35×18. 77 kmph ( n3 = 580 rpm) V2 (kmph)= 19.32×1.283]÷3.1 V4 = 53. Page 52 .76×1.6× π/30×3700×0.35 T1 =638.1 V3 = 34.6×π/30×ηmax)rdyn]÷iA ×iE = [3.66 kmph ( n1 =176.55 N.84 kmph (n2 = 331 rpm) V1 (kmph)= 10.66 rpm= Roadwheel rpm) Torque Available on 1st Gear =G1 ×Max Torque of Engine T1 =33 ×19. Design Analysis And Optimization Of All-terrain Vehicle (ATV) Speed Calculations On Each Gear- V4 (kmph)= [(3.6× π/30×3700×0.08 kmph (n4 = 885 rpm) V3 (kmph) = [3.m Using 2 CVJ joints at Gear Box side – Maruti 800 Sleeve arrangement for drive shaft length correction is made.283] ÷10. 46 Mpa Max Deformation 0.13 mm Factor of Safety 2.74 Page 53 .Design Analysis And Optimization Of All-terrain Vehicle (ATV) Adapter ansys: PARAMETER VALUE Max Equivalent 193.46 Mpa Stress Max Shear Stress 104. Page 54 . Modena. A particular focuses given to weight reduction in automotive chassis design applications following the experience matured at Mille Chili Lab. DESIGN Marco Cavazzuti and Luca Splendi (joint with Luca D'Agostino. SET UP.it ABSTRACT Improvements in structural components design are often achieved on a trial- and-error basis guided by the designer know-how. A different turn of mind that could boost structural design is needed and could be given by structural optimization methods linked with niter elements analyses. These methods are here brief introduced . such an approach is likely to allow only marginal product enhancements. Dario Costi and Andrea Baldini) MilleChili Lab. Despite the designer experience must remain a fundamental aspect in design. Dipartimento di Ingegneria Meccanica e Civile. Italy Universit_a degli Studi di Modena e Reggio Emilia millechililab@unimore. Enrico Torricelli.and some applications are presented and discussed with the aim of showing their potential. Design Analysis And Optimization Of All-terrain Vehicle (ATV) APPENDIX – E STRUCTURAL OPTIMIZATION OF AUTOMOTIVE CHASSIS:THEORY. Pedersen [2]. and Duddeck [3] are of interest. Despite this. In the following some of these techniques will be introduced and their application to chosen automotive structura ldesign problems discussed. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 1. In particular. These techniques are an effective approach through which large structural optimization problems can be solved rather easily. yet they are not always well known and applied in industry. with the term structural optimization methods we refer to: (i) topology optimization. To cite a few applications in the automotive _eld the works of Chiandussi et al. and car bodies respectively . Page 55 . (iv) size optimization. the literature over the topic is quite rich and is addressing both theory and applications.Structural optimization methods are rather peculiar ways of applying more traditional optimization algorithms to structural problems solved by means of _nite elements analyses. They address the optimization of automotive suspensions. (ii)topometry optimization. [1]. (iii) topography optimization. (v) shape optimization. INTRODUCTION Optimization techniques are very promising means for systematic design improvement in mechanics. crushed structures. STRUCTURAL OPTIMIZATION In the de_nition of any optimization problem a few elements are necessary. 2.g. and size. The optimization was performed in three stages: topology. bottom view (c) optimum layout Figure 1: Ferrari F458 Italia front hood: reference model and new layout from the optimization results. Topology optimization was _rstly introduced by Bends_e and Sigmund and is extensively treated in [4].in our case. it has developed in several directions giving birth to rather different approaches. (vi) the optimization algorithm (e. stiffness and/or displacement stargets). in structural optimization this is commonly a gradient- based algorithm. targets and objectives are evaluated (e. topometry. top view (b) reference model. for a given set of variables. it could be said that the various structural optimization methods essentially differ from each other in the choice of the variables of the optimization problem as follows. such as MMA).g.. this method does not apply to 3D elements where the concept of thickness could not be de_ned.1. the variables being the element-wise thicknesses. mass minimization). Page 56 . The optimization was performed in three stages: topology. The variables are then given by the element-wise densities. in structural optimization this is often given by the mesh) (ii) variables. (v) the mean through which. _nite elements analyses). and size. bottom view c) optimum layout Figure 1: Ferrari F458 Italia front hood: reference model and new layout from the optimization results. the most simple and known of which is the SIMP (Single Isotropic Material with Penalization). Topology Optimization In topology optimization it is supposed that the elements density can vary between 0 (void) and 1 (presence of the material). Topometry Optimization The idea behind topometry optimization is very similar to that of topology optimization.g. these are: (i) design space or space of the possible solutions (e. (iv) optimization constraints (e.2. (a) reference model. (iii)objective(s) (e.g.g. top view (b) reference model. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 2. optometry. 2.Trying to simplify in a few words a rather complex and large topic.(a) reference model. Of course. The results have been re-interpreted into more performing thin-walled cross-sections. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 2.2. Thus. Page 57 .3. The targets relate to bending and torsion static load cases. the variables are given by the set of the elements o_sets from the component mid-plane. yet in the respect of all the performance requirements (Fig. Topography optimization was used to improve the beads disposition in the panel. The objective of the optimizations was mass minimization. 2. in terms of beads and thickness. the damping material distribution is not known during the numerical veri_cation stage. The solution was re_ned through size optimization. 2). Rear Bench The rear bench of a car is fundamental to isolate acoustically the passengers compartment from the engine. Size optimization is applied to control the thickness of the aluminum plate and of the vibrational-damping material.4. where the material is added iteratively to counteract the _rst normal modes. Size Optimization Size optimization is the same as topometry optimization. The concept is yet similar to the previous cases and. A series of topometry optimizations followed to _nd the optimal thickness distribution and identify the most critical areas. have been optimized through size and topography optimizations at the same time.1. compliance when closing the hood. Generally. deformations under aerodynamic loads. Topography Optimization Again topography optimization can be applied only to 2D or shell elements and aims at _nding the optimum beads pattern in a component. whereas several thickness variables were created locally for the damping layer. APPLICATION EXAMPLES 3. but in this case the number of variables is greatly reduced in that the shell thicknesses of components are considered in place of the single elements of the domain. 1). The presence of damping material should be limited to essential parts due to its relatively high weight. In the end. while the _rst normal mode frequency was constrained to be outside the range of interest (Fig. but is decided later during the experimental analysis. 3. simply speaking. The bench of Ferrari F430 has been analyzed with the objective of reducing the weight while maintaining the same vibrational performance of the reference panel. 3. the weight was reduced by 12 %. just one thickness variable was created for the aluminum layer because its value should be uniform along the plate. Automotive Hood The internal frame of the Ferrari F458 front hood has been studied aiming at reducing the weight while keeping the same performance target and manufacturability of the reference model. A suitable preliminary architecture has been de-_ned by means of topology optimization. In this study vibration-damping material distribution and panel design. The objective of the optimization is still the weight reduction while the performance requirements regard handling and safety standards. The initial design space is given by the provisional vehicle overall dimensions of Ferrari F430 including the roof (Fig. or design space (b) optimum chassis con_guration (c) optimum roof con_guration (a) domain.3. (iv) local sti_ness of the suspension. (iii) modal analysis. and gearbox joints.1 (blue) to 1. 3(b) and 3(c). or design space (b) optimum chassis con_guration (c) optimum roof con_guration Figure 3: Automotive chassis topology optimization. The results for the chassis and the roof are shown in Figs. blue stands for low deformation/thickness.(a) domain. in detail: (i) global bending and torsional stiffness’s. (ii) crashworthiness in the case of front crash (a) Size optimization variables (b) Optimum con_guration (c) Damping material optimum thickness subdivision deformed shape distribution Figure 2: Rear bench coupled optimization. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 3.0 (red). the density range from 0. 3(a)). In the results. engine. red for high. In the results. Automotive Chassis Topology optimization has been applied to the design of an automotive chassis. Page 58 . A more detailed discussion on a combined methodology for chassis design including topology. topography and size optimizations was presented in [5] by the authors. Their potential has been shown to be large and it is believed that their spreading in mechanical design could boost innovation in industry considerably. To be noted that the different methods have different characteristics and in a design process it is recommended to rely on more than just one technique. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 4. CONCLUSIONS A quick overview on structural optimization methods has been given including various application examples. For instance. but they give useful hints to the designer in view of the product development and engineering. whose outcome could be further re_ned through size and shape optimizations. Examples in the automotive _eld have been provided. topology and topometry optimizations are more suitable for an early development stage. On a general basis these techniques do not deliver the shape of the _nal product. Page 59 . Baldini. E. D. Multidisciplinary optimization of carbodies. Duddeck. Springer. Chiandussi. 3] F.2004.Torricelli. Page 60 . Bertocchi. [4] M. Moruzzi. Cavazzuti. 44:45-56. Sigmund. and P. Structural and Multi disciplinary Optimization. methods and applications. [2] C. 2008. REFERENCES 1] G. Structural and Multidisciplinary Optimization. 5] M. I. B. 35:609-617. and A. P. Pedersen. W. Bends_e and O. A. Ibba. Design Analysis And Optimization Of All-terrain Vehicle (ATV) 5. Topology optimization: theory. Advances in Engineering Software. Topology optimization of an automotive component without _nal volume constraint speci_cation. Gaviglio. 2004. 35:375-389. 193:653-678.Computer Methods in Applied Mechanics and En- gineering. Costi. E. 2011. Crashworthiness design of transient frame structures using topology optimization. High performance automotive chassis design: a topology optimization based approach.