Design, Analysis and Testing of the Primary Structure of a Race Car for Supra SAEINDIA Competition

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1 Design, Analysis and Testing of the Primary Structure of a Race Car for Supra SAEINDIA Competition P.K. Ajeet Babu, M.R.Saraf Automotive Materials Laboratory, Automotive Research Association of India, Pune Copyright 2012 SAE International K.C.Vora ARAI Academy, Automotive Research Association of India, Pune ABSTRACT An open wheeled open cockpit racing-car with Maruti 800 BS III MPFI engine was designed and built for the SUPRA SAE INDIA 2011 student competition by ARAI academy students. The car won the best engineering design award and second in CAE category. The key focus was to build a car with superior safety and comfort. This work discusses in detail about the design, analysis and validation of the primary structure of the car which includes chassis and impact attenuator. This paper will serve as a reference to the future students who are participating in supra competition. A review of supra sae rules relating to chassis is used to develop realistic parameters. Based on the weight distribution calculations and the hard points the chassis is designed, modelled using cad tool and simulated for longitudinal acceleration, lateral acceleration, braking condition, braking with one circuit failure, bending stiffness and Torsional stiffness using finite element analysis. The chassis is then fabricated and validated in actual condition using data acquisition tools. engineering design, solo performance trials, and high performance endurance test on the track. These events are scored to determine how well the car performs. DESIGN REQUIREMENTS (CHASSIS) - Among other requirements, the vehicle s structure must include two roll hoops that are braced, a front bulkhead with support system and Impact Attenuator, and side impact structures. The Primary Structure of the car must be constructed of Either Round, mild or alloy, steel tubing (minimum 0.1% carbon) of the minimum dimensions specified in the following table[1]. Roll Hoops - The driver s head and hands must not contact the ground in any rollover attitude. The Frame must include both a Main Hoop and a Front Hoop as shown in Figure 1. The impact attenuator protects the driver during frontal crash and it was mounted on the front bulkhead of the car. It was made of aluminium honeycomb structure and designed such that during crash, the average deceleration of the vehicle does not exceed 20g. It is then fabricated and tested using drop test method. INTRODUCTION The SUPRA SAEINDIA 2011 competition challenges teams of university undergraduate and graduate students to conceive, design, fabricate and compete with their own small, formula style, autocross racing cars. Teams are only permitted to use 800 cc BS III engine and power train supplied by Maruti Udyog ltd. JUDGING CATEGORIES - Participating cars are judged in a series of static and dynamic events including: technical inspection, cost, marketing presentation, Figure 1 Roll Hoop requirements Main and Front Hoops - When seated normally and restrained by the Driver s Restraint System, a straight line drawn from the top of the main hoop to the top of the front hoop must clear by 50.8 mm (2 inches) the helmet of all the team s drivers and the helmet of a 95th percentile male (anthropometrical data). 95th Percentile Male Template Dimensions A two dimensional template

2 used to represent the 95th percentile male is made to the following dimensions. The Main Hoop must be constructed of a single piece of uncut, continuous, and closed section steel tubing. The Front Hoop must be constructed of closed section metal tubing. With proper gusseting and/or triangulation, it is permissible to fabricate the Front Hoop from more than one piece of tubing. different sections. The front is a Monocoque, consisting of a balsa-wood core with aluminium skins that securely contain the driver and the front suspension mounts. The back portion is a welded aluminium frame that houses the engine and rear suspension mounts. When both parts are completed, they are joined to form the frame for the race car. DESIGN REQUIREMENTS (IMPACT ATTENUATOR) - The driver s feet must be completely contained within the Major Structure of the Frame. While the driver s feet are touching the pedals, in side and front views no part of the driver s feet can extend above or outside of the Major Structure of the Frame. Forward of the Front Bulkhead must be an energy-absorbing Impact Attenuator. The Front Bulkhead must be constructed of closed section tubing. The Front Bulkhead must be located forward of all non-crushable objects. The Impact Attenuator must be: Installed forward of the Front Bulkhead. At least 200 mm (7.8 in) long, with its length oriented along the fore/aft axis of the Frame. At least 100 mm (3.9 in) high and 200 mm (7.8 in) wide for a minimum distance of 200 mm (7.8 in) forward of the Front Bulkhead. Such that it cannot penetrate the Front Bulkhead in the event of an impact. If the Impact Attenuator is foam filled or honeycomb, a 2.5 mm (0.060 in) solid steel or 4.0 mm (0.157 in) solid aluminium metal plate must be integrated into the Impact Attenuator. The metal plate must be the same size as the Front Bulkhead and bolted or welded to the Front Bulkhead. Figure 2 Torsion of different Roll cage Space Frame technique - The concept is based on 2- dimensional modular space frame[3] sub-structures and utilizes an outer Steel structure with a carbon fiber honeycomb panel inserted between steel members. Attached securely and directly to the Front Bulkhead and not by being part of non-structural bodywork. The attachment of the Impact Attenuator must be constructed to provide an adequate load path for transverse and vertical loads in the event of off-centre and off-axis impacts. If not integral with the frame, i.e. welded, a minimum of four (4) 8 mm Grade 8.8 (5/16 inch Grade 5) bolts must attach the Impact Attenuator to the Front Bulkhead. Impact Attenuator, when mounted on the front of a vehicle with a total mass of 300 kgs (661 lbs) and run into a solid, non-yielding impact barrier with a velocity of impact of 7.0 meters/second (23.0 ft/sec), would give an average deceleration of the vehicle not to exceed 20 g. PREVIOUS DESIGNS CHASSIS - The Chassis of A race car can be of different types and they are as follows: Aluminium Monocoque Livio Nichilo et al[2] of Queen s university has demonstrated the Monocoque design in which the frame of the race car is comprised of two Figure 3 Carbon Fibre frame blow out This type of design presents some advantages over the single, diagonal frame member. One advantage being that the "shear plate" would serve the dual purpose of being a structural frame member and means of closing out the cockpit area as per FSAE rules. The second advantage comes in the ability of the plate to transfer loads throughout the entire structure. The extreme case of this shear plate design being a monocoque structure

3 where loads are transferred through panels of aluminium or carbon fiber sandwich material. The 2007 Western FSAE car used a rectangular prism made of aluminum honeycomb for their impact attenuator. The total crush strength of this design was found to be 2.65 MPa. Their design reduced the overall weight of the attenuator by about 476%. The team at the University of New Hampshire also came up with a similar design that met the criteria for FSAE impact attenuation. Last year s Worcester Polytechnic Institute FSAE car used aluminum honeycomb as well. The team at San Jose State University used a steel truncated pyramid for their crash protection. This design included a hollow interior within the impact attenuator, effectively reducing the weight. The strength of the steel met the requirements necessary for the 2008 FSAE rules. Figure 4 Kevlar taped Carbon Fibre panel Super Light weight space frame - Masao Ishihama[4], et al designed a formula type racing-car for the Formula SAE competition. Lightweight design was achieved separately in two areas, i.e. the pipe frames and bulkheads, by employing two different strategies. In the pipe frame design, the stretch method was used. The baseline design employed the minimum number of pipes required to make the structure statically stable. After this, reinforcements were attached one-by-one, until the rigidity reached a satisfactory level. The Oxford Brookes University received and innovative design award for their impact attenuator, which was made from balsa wood. No exact details could be found regarding their design but it had to have met the standards set forth by the FSAE in order to participate in competition. DESIGN CHASSIS - Weight Distribution Bulkhead : 2.8 Kg Front Assembly cage : 8.30 Kg Cockpit : Kg Front Roll Hoop : Kg Side Pod : 8.9 Kg Main Roll Hoop : 12.3 Kg Engine : 15.9 Kg Welding : 5 Kg Graph 1Front to Rear Weight Distribution chassis Figure 5 Space frame modifications for stiffness improvement IMPACT ATTENUATOR - In order to get an accurate idea of what exactly is to be created; research was done to see what cars have used in the past. Many different designs have been used in the past. The goals of this project include utilizing the concepts of ones used before and making improvements on them or formulating an entirely new design.

4 Vehicle Weight Part Wise Chassis :65.5Kg Front Upper Wishbone x 2 :5.0 Kg Front Lower Wishbone x 2 :5.8 Kg Rear Upper Wishbone x 2 :5.0 Kg Rear Lower Wishbone x 2 :5.4 Kg Tire With Wheel x 4 :48 Kg Power Pack :140Kg Gear Linkage :2.0Kg Tank with Pump :8.0 Kg Steering system :3.7 Kg Front Knuckle x 2 :5.0 Kg Rear Knuckle x 2 :8.0Kg Seat Belt :3.4 Kg Front Hub x 2 :8.2 Kg Rear Hub x 2 :5.8 Kg Exhaust :3.5 Kg Pedal Assembly :5.0 Kg Wiring :1.0 Kg Battery :15.0Kg Engine Oil & Coolant :5.0 Kg Miscellaneous :8.0 Kg Total : Kg Torsional stiffness is one of the most important properties of a vehicle chassis. Lack of chassis Torsional stiffness affects the lateral load transfer distribution, it allows displacements of the suspension attachment points that modify suspension kinematics and it can trigger unwanted dynamic effects like resonance phenomena or vibrations. It is in fact known that in order to obtain good handling performances, stiffness, together with lightness and weight distribution is one of the most important properties of a chassis. Table 1: Torsional Stiffness of different vehicles[6] Vehicle Chassis Torsional Stiffness[Nm/deg] Formula SAE car Passenger car Winston cup racing car Sports car Formula one car Chassis Design Considerations- Design of the race car began with the frame. There are several factors that must be considered when designing the frame. Stiffness - Normally, a race car chassis should be as stiff as possible torsionally. This is to facilitate easier suspension tuning. When determining the handling qualities of a race car, one of the most effective methods of adjusting the amount of over steer and under steer is the adjustment of roll stiffness front-to-rear. By increasing front roll stiffness while decreasing rear roll stiffness, both rear tires are more equally weighted than the front tires. The force on the outside front tire quickly overwhelms the traction available to it, and the car under steers. Conversely, with a large amount of rear roll stiffness and a small amount of front roll stiffness, the inside rear tire is lifted during a turn, the amount of available rear traction is reduced, and the car over steers. By tuning the stiffness of the anti-roll bars, it is possible to affect the balance of the car. However, Torsional flex in the frame adds another spring to this two-spring system. Weight - As discussed earlier, wherever possible, weight should be minimized. All tubing sizes not dictated by the rules were chosen to be as light as possible while remaining structurally sound and suitably stiff. Just as important as weight is mass moment of inertia. A car with a lower mass moment of inertia will be able to turn more quickly. In order to reduce mass moment of inertia, all weight on the chassis is pushed as far as possible towards the centre of the vehicle. Fitment and Packaging - Possibly the most difficult criterion to satisfy is fitment. This criterion determines the functionality of the chassis. The chassis must accommodate the driver, as well as the engine, suspension components, and templates while remaining as light and small as possible. While a problem with structural integrity or stiffness can usually be solved by simply varying the wall thickness or diameter of a tube, the challenge of fitting all components into the smallest space possible rarely has clear or straightforward solutions. Multiple iterations are performed for fitment problems[5]. CHASSIS CONSTRUCTION METHODS (TUBULAR SPACE FRAME) - The most common frame type, the tubular Spaceframe, is a structure composed of many small, usually round tubes bent to shape and welded together. Tubular Spaceframe do not require specialized machinery or equipment for manufacture, and they are inexpensive and can be constructed from a wide variety of readily available materials. The SUPRA SAEINDIA rules dictate many of the tubing sizes for a steel tubular Spaceframe, and construction of any other type of chassis requires proof that the alternate structure is as strong as or stronger than a similar tubular Spaceframe structure. CHASSIS MATERIAL CONSIDERATIONS - The most common material for tubular Spaceframe, steel retains its strength and ductility after welding. It is inexpensive, easy to find, and easy to cut and grind. The rules dictate

5 tubing sizes for steel, and the use of any other material requires the completion of a structural equivalency form. For ease of construction, the chassis is made from steel. In order to determine the type of steel to be used, further analysis is necessary. In the interest of safety, the rules dictate most of the tubing sizes used. CHASSIS ANALYSIS MESHING - The element type is SHELL 63 and the size of the meshing size is of 10. Figure 6 Meshed model of Chassis CASE 1: STATIC STRUCTURAL ANALYSIS - Static structural analysis is done to analyze the car in a static condition. The Engine is mounted at three points and each points will carry a load of 360 N, Each suspension point carries 70 N, The driver sits in the seat and the seat is attached to the chassis at three points and the total weight is distributed to three points in 260 N, The Battery of the car is attached at two points and the load is 125 N each, The radiator and fuel tank mounting point load is 120 N and 130 N respectively. Figure 8 Stress plot CASE 2: Lateral Acceleration Analysis The formula car when enters a turn or exits experiences a force of 2.5g. All the components of the car experience the same force. The direction of the load is given in Y direction. The analysis results show that the maximum stress developed is MPa and thus it is safe which is shown in Figure 9. The deformation shape is given in Figure 10 The Chassis is constrained in all degrees of freedom at and the displacement plot is shown in Figure 7. The stress analysis is done and the maximum von misses stress is MPa which makes the chassis safe in static condition. Figure 8 shows the stress plot. Figure 9 stress plot Figure 7 Displacement plot Figure 10 Displacement plot

6 CASE 3: LONGITUDINAL ACCELERATION ANALYSIS - The formula car when experiences a force of 2g during the race event. All the components of the car experience the same force. As the weight transfer is to the rear the load direction at X direction. The analysis results show that the maximum stress developed is MPa and thus it is safe which is shown in Figure 11. The deformation shape is given in Figure 12. Figure 13 stress Plot Figure 11 Stress Plot Figure 14 Displacement Plot Figure 12 Displacement Plot CASE 4: BRAKING ANALYSIS - The formula car experiences a force of 2g during the braking process. The weight transfer is to the front and it is around 88% at the front and rest at the rear. The load direction at X direction. The analysis results show that the maximum stress developed is MPa and thus it is safe which is shown in Figure 13. The deformation shape is given in Figure 14. CASE 5: BRAKING WITH ONE CIRCUIT FAILURE - The SUPRA rule book mentions that there should be two independent brake circuits so that there is at least two wheels which can be locked properly in case of any accident. Thus we have done the analysis of chassis in a case of brake failure of one circuit. To simulate the real condition the constraint at two diagonally opposite set of suspension point is removed. The maximum stress developed is 143 MPa, 12 MPa higher than without failure. The stress diagram and deformation mode is given in Figure 15 & 16 respectively. Figure 15 Stress plot

7 Figure16 Displacement plot CASE 6: BENDING ANALYSIS - Bending stiffness is one of the most important criteria of a chassis. To analyze the chassis for bending stiffness the constraint point chosen is at the bottom of the front hoop and the loads are acted at the downward direction of 1000 N each side. Figure 17&18 shows the result of bending analysis. Figure 18 Displacement plot CASE 7: TORSIONAL ANALYSIS - Torsional stiffness for a typical formula SAE car is around 1500 Nm/deg. If the Torsional stiffness is low then the following will be the consequences: The control of lateral load transfer distribution is difficult and the vehicle does not respond as expected to setup changes Displacements of the suspension attachment points occur, so that the desired control of the movement of the tires cannot be guaranteed Dynamic effects like vibrations can occur. Fatigue phenomena are more marked. Ride quality is poor. The chassis is tested for its Torsional rigidity by constraining the end four points at the rear and twisted at the front by applying a load of 1000 N. The various results are shown in the figure totally the chassis is analyzed in ANSYS for eight different loading conditions and the factor of safety is obtained as two. Figure 19&20 shows the result of bending analysis. Figure 17 Stress plot Figure 19 Stress plot

8 Figure 20 Displacement plot Figure 22 Third mode of Chassis CASE 8: MODAL ANALYSIS - The modal analysis of the chassis is very important to find whether the chassis is in resonance at any instant. The frequency of vibration due to road undulations ranges up to 25 Hz. The result of the modal analysis shows the frequency to be around 60 Hz which explains us that the chassis is safe from resonance and as the load is added when subsystem is assembled. Thus the chassis is analyzed for various critical conditions and the results for maximum displacement and von misses stress are plotted as follows:- Table2 Results of Chassis analysis Figure 20 First mode of Chassis Condition Displacement,mm Von Misses Stress, MPa Acceleration 2g Cornering 1.5g Braking 1.5g Braking one circuit failure 1.5g Bending Torsional Static Structural Figure 21 Second mode of Chassis

9 IMPACT ATTENUATOR DESIGN - After completing extensive research on previous designs and physics of a collision, preliminary design concepts were formulated. These designs were then compared and contrasted against each other based on cost, safety, reliability, feasibility and weight[7]. Finite element analysis was conducted to get an idea of how the attenuators would deform under the high stresses of a collision. The complex honeycomb structure proved problematic during the analysis[10]. The program was unable to mesh the pieces of the honeycomb pattern and complete a full test. A solution for this was to model each layer as a solid block but giving it the same properties. Due to the magnitude of force during impact and the non-linear deformation of honeycomb the program we used was unable to produce accurate and reliable results on our models. However a depiction of how the model would crush based on the test results. The next step was to manufacture the two designs and prepare them for physical testing. The method of testing chosen was the drop test. Preliminary Calculations - Initial Conditions: crane and freely dropped using a quick release mechanism. Besides the HSP, videography as well as digital photography of the impact attenuator before and after the drop test was taken. Figure 22 Drop Test set up V impact =7 / V final =0 / =9.8 / 2 =300 =20 =196 / 2 Kinetic Energy: =1/2 ( ) 2 = / 2 =7350 J By Conservation of Energy, Kinetic Energy is equal to potential energy =.Calculating the Desired Drop Height: = = / ( ) =7350/ ( ) =2.5 =8.2 Time of Impact: t=vimpact/ac t=.036s Impulse and Force: Im=M (Vimpact Vfinal) Im=2100 kg m /s F=Im/t =58,333 VALIDATION (Drop Testing of Impact Attenuator)- Instrumentation High Speed Photography (HSP) specifications: Redlake HG100 K, CMOS type, x 600 pixels camera along with SAI Image Express Motion Plus analysis software. A weight of 250 Kg was raised 2.5 metre into the air and released onto each prototype design. This height and weight were chosen to meet the requirement of the car reaching a velocity of 7 m/s at the point of impact. During the test a high speed camera recorded the test in slow motion so that the deformation patterns of each design could be analyzed. The impact attenuator was kept on hard plate and drop weight was lifted to 2.5 m height using Goliath Fig 23.Specimen one before testing Fig 24.Specimen one after testing

10 Graph 3 Acceleration plot Sample two Fig 25.Specimen two before testing CONCLUSION The two most important safety features of a SUPRA Car is the chassis and Impact attenuator. A review of rules relating to chassis & Impact attenuator is used to develop realistic parameters. The chassis is analyzed for eight different worst case scenarios and the Impact attenuator design is validated by conducting drop test. The actual vehicle dimensions are as follows: Fig 26.Specimen two after testing The average deceleration of the sample one is g and thus the Impact attenuator design meets the requirement of average deceleration not to exceed 20 g s. Graph 2 Acceleration plot Sample one Vehicle Weight Driver Weight Total Vehicle Weight Wheel Base Front Track width Rear Track width Ground Clearance Front Ground Clearance Rear Vehicle weight actual a) Vehicle Without driver : 345 Kg : 69 Kg : 414 Kg : 2027 mm : 1560 mm : 1524 mm : 100 mm : 105 mm Right Hand Front : 59 Kg Left Hand Front : 58 Kg Left Hand Rear : 117 Kg Right Hand Rear : 111 Kg Total : 345 Kg b) Vehicle With driver Right Hand Front : 76 Kg Left Hand Front : 76 Kg Left Hand Rear : 133 Kg Right Hand Rear : 129 Kg Total : 414 Kg

11 8. Andrew Deakin, David Crolla, Juan Pablo Ramirez and Ray Hanley. The Effect of Chassis Stiffness on Race Car Handling Balance, , SAE TECHNICAL PAPER SERIES, School of Mech. Eng., The University of Leeds. 9. Brian Auer, Jared McCombs and Edwin Odom. Design and Optimization of a Formula SAE Frame, , SAE TECHNICAL PAPER SERIES, University of Idaho. Figure 27 SUPRA CAR (Team Sultanz of Speed) ACKNOWLEDGEMENTS The Author would like to thank Bharat Petroleum Corporation Limited (BPCL) to financially support the project and Mr. Shrikant Marathe (Director, ARAI) for supporting the team in validation of components/assemblies and the entire vehicle. 10. A report on 2009 Formula SAE Race car, WORCESTER POLYTECHNIC INSITUTE, Project Number: A MECH 460 Final Design Report, 2006 Formula SAE Chassis Design, Queen s University, Department of Mechanical Engineering 12. Fundamentals of Vehicle dynamics by Thomas D Gillespie REFERENCES 1. SAEINDIA Publication, SUPRA SAEINDIA 2011, Rule Book, Release 2 on 2nd August Livio Nichilo and Marc Auger, Tomas Konvalina. Assessment of a Glued Aluminium Monocoque and Suspension for Formula SAE Style Race cars, , SAE TECHNICAL PAPER SERIES, Dept. of Mech. Eng., Queen s University 3. A. Henningsgaard and C. Yanchar. Carbon Fiber Reinforced Steel Spaceframe Techniques, , SAE TECHNICAL PAPER SERIES, Department of Mechanical Engineering University of Minnesota Twin Cities. 4. Masao Ishihama, Shingo Iizuka, Kazuo Tanahashi, Aki Higeuchi and Mayumi Fukuda. Optimization of Super-Lightweight Space-Frame Vehicle Structure, , SAE TECHNICAL PAPER SERIES, Kanagawa Institute of Technology. 5. Eva Mariotti and Badih Jawad. Formula SAE Race Car Cockpit Design An Ergonomics Study for the Cockpit, , SAE TECHNICAL PAPER SERIES, Lawrence Technological University. 6. Enrico Sampo, Aldo Sorniotti and Andrew Crocombe. Chassis Torsional Stiffness: Analysis of the Influence on Vehicle Dynamics, , SAE TECHNICAL PAPER SERIES, University of Surrey 7. Worcester Polytechnic Institute, FSAE Impact Attenuator by Jon Hart Craig Kennedy Todd LeClerc Justin Pollard,

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