chapter-1 1. introduction

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We are a dedicated group of students seeking a Bachelor in Science of Mechanical .... Choosing the optimum FOS is important and BS 5950-. 1:2000 guidelines ...

CHAPTER-1 1. INTRODUCTION Baja is a special kind of four-wheeled vehicle used for recreational and exploration purposes. It is designed for off road usage and for endurance of a rough terrain. In many aspects it is similar to an All-Terrain Vehicle (ATV) except that it is much smaller in size and has safer rollover capabilities. Besides these any Baja vehicle should also be easily transported, easily maintained and fun to drive. A National competition is organized by the Society of Automotive Engineers (SAE India) for colleges throughout the world to design and fabricate their vehicles and then compete against each other. This project was aimed to develop the design of a chassis which is safe, ergonomic and has the lowest possible weight. Competitiveness of the vehicle in terms of ruggedness and manuverability had also been kept in mind at the virtual and final working model of the vehicle which is presented in figure 1.1 and figure1.2

Figure.1.1 Virtual Model of ATV A

preliminary

design

was

Figure.1.2 Real Model of ATV first

prepared

keeping

in

mind

the

guidelines issued by SAE. Indian standards for driver space have been incorporated and a PVC mock-up was developed to evaluate the driver ergonomics. The CAD modelling of the frame and other components was done. This design was checked by Finite Element Analysis after estimating the load and the weight of the frame optimized. A rollover analysis was then carried out to ensure safety in such a situation. The Rollover Analysis involved evaluating static stability and ensuring compliance with the pertinent Indian Standards. Hence, after ensuring safety, the design was finalized and the frame was then manufactured after procuring the tubes from the market. 1

1.1. Team Statement We are a dedicated group of students seeking a Bachelor in Science of Mechanical Engineering that share a passion for motor sports and outdoors activities. We will design keeping in mind concepts for Fit, Form and Function while being able to comply with all of the requirements that SAE has laid out in order to create a functional yet safe and reliable vehicle.

1.2. Ethical Design Statement This project is being executed and analyzed in accordance to the NSPE Code of Ethics for all purposes of design and prototyping of all components and assemblies that will be created as part of this design concept.

1.3. Environmental Impact Statement The design and testing of the prototype is being manufactured in such manner that the impact that it will have to the surrounding environment is minimum and has no permanent adverse effects on the community.

1.4. Project Statement The SAE Mini Baja competition includes over 120 teams from various engineering institutes who design, build, and compete in Baja style off-road vehicles that must survive rough terrain. This competition focuses on the chassis and suspension design of the car, restricting the motor to a standard ten horsepower Briggs and Stratton. We are challenged to design a rugged and reliable off-road capable car, while still keeping cost and maintenance affordable for the theoretical consumer. Competition is focused on the design and innovation of cost effective parts, as well as dynamic events that test the real world function of these vehicles. The cars must follow strict safety and cost regulations, making the design of these cars challenging for student engineers.

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CHAPTER-2 2. DESIGN PLANNING 2.1. RollCage The chassis or Space frame of the mini Baja is called as roll cage. The roll cage serves many critical functions that include linking the power train, control, and suspension systems together. The driver must also be comfortable in order to operate the vehicle effectively, thus driver ergonomics and safety take precedence. Mounting points and the overall frame geometry are crucial design considerations that affect desired characteristics such as the weight distribution and suspension operation. The Roll cage must also be resilient enough to endure all of the loads imposed upon it yet maintain a light weight. It must also be capable of protecting the driver and crucial components of the vehicle while having an impact towards rigid structures or accidents with other vehicles.

2.2. Problem Statement Design and carry out Finite Element Analysis and Rollover Simulations for an off road vehicle chassis with suitable tube diameter, wall thickness and material to withstand different load conditions to have a high Factor of safety, least weight and reasonable cost.

2.3. Design Objectives The primary objective of the Roll cage is to provide a 3-dimensional protected space around the driver that will keep the driver safe. The secondary objectives are to provide reliable mounting locations for components, be appealing, low in cost, and low in weight. In addition roll cage must support all operator control systems, front and rear suspension systems, and engine and drive train. Other design factors included durability and maintainability of the frame.

2.4. Plan Of Work The Plan of Work for the project was divided into 5 major stages. Initially, we prepared a preliminary design based on specifications given using solid works designing software. This then was tested for driver space by building a plastic mock-up model using PVC pipes. It was then followed by refinement in design. Simultaneously the mounting points of the 3

various assemblies which include drive train, suspension, engine, seat, etc. are designed. This design was then checked using Finite Element Analysis with solid works as initial tool and further refinements are been made in the design. Hypermesh and Ansys are used for final FEA analysis with optimized data’s in terms of weight and safety to ensure the impact capabilities and safer roll over capabilities.

2.5. Methodology Chart This chart helps in framing the Gantt chart that will be used to predetermine the time taken to complete the project and pre plan the work. it also acts as a guide therefore a process is done without wasting time in shuttling back and forth for the required data’s and elements in the designing phase. The Figure 2.1 is the stage by stage flow chart that is derived from the plan of work.

Figure 2.1.Roll Cage Building Methodology.

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2.5.1 GANTT Chart The methodology developed previously lead to the planning of the timed schedule of each work, it helped in completing the project on time as shown in figure 2.2.

Figure 2.2 GANTT Chart

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2.6. Design Goals In order to achieve our desired objective various technical goals are been set up. These are listed below and described further. Table 2.1Design Goals Parameters

Properties

Material

Steel

Shape

Circular

Material Elongation

High

Strength to weight ratio

High

Factor of safety

High ( more than 1.25)

Centre of gravity

Low

Attractive design

Aesthetically pleasing

Cost

Low or reasonable

Manufacturability

Easy

Weld ability

Good

Overall vehicle length

Less than 180 inches

Ride height

More than 10 inches

Overall vehicle height

As low as possible

Overall vehicle width

Less than 64 inches

Crash sustainability

Very high

Roll over sustainability

Very high

Endurance sustainability

Very high

Driver ergonomics

Good

The material to be selected must be of the steel grade. Though various materials like composites are available steel will meet the criteria of the Sae rules and the cost and manufacturing views of the other materials are inferior to the steel. Also the steel tube is available easily in market than other materials. Tubes shapes that are commercially available in the market are round, rectangle and square. Of these the most suitable for manufacturing a space frame will be round tubing

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because of its geometric property of withstanding load and undergoing deformation equally in spite of its axis. Rather the thickness of the round tubing variety is larger than the other shapes. The steel material that is selected must have the elongation property high. It must be ductile with high stiffness. even under crucial loading conditions it must elongate rather than sudden breakage. One of the main criteria that to be followed throughout this project is strength to weight ratio. The strength and weight need to be balanced at a very crucial scale since it is the wrestling between safety and performance respectively. The design must also be made to achieve structural rigidity rather than completely depending upon the strength of the material. The FEA analysis and simulation is the closest possibility of the result that the vehicle will undergo in actual scenario. Therefore the ratio of the ultimate strength to the working strength must be higher. But at the same time when this ratio increases proportionally all other considerations also increase. Choosing the optimum FOS is important and BS 59501:2000 guidelines are been used to decide it. Space frame being the component that weights much high and the usage volume is high its individual centre of gravity affects the overallcentre of gravity of the vehicle .it is necessary for the centre of gravity of the vehicle to be lower to ground as much as possible. The roll cage must be therefore designed in such0 a scenario where the centre of gravity is lower to ground. Being designed advantageously over engineering factors never allows the market potential to be low. People mostly attracted towards the aesthetically pleasing vehicle. Good ergonomic design ,stylish look , high speed performance are the few that gets into mind of people when buying vehicle hence it's taken into consideration. Cost as an main factor, cheap in the best policy is followed. Sacrificing the good for the cost is not a good idea. Therefore the best items that will suite the purpose will be selected and then the cheapest and the reasonable will be considered. While the design phases itself virtual assembly of the components will be done so that the manufacturing difficulties will come down. The process, machinery used and feasible

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fabrication must be considered .The design for manufacturability guidelines will be followed to achieve the desired. The material that is selected for the fabrication of the roll cage must be a material that can be welded together. Even though the roll cage can be bolted to one another the safety factor is very low .good welding process shall be carried out in order to obtain good strength. The overall length of the vehicle is been restricted to be less than 108 inches so that during the turning, performance and the handling of the vehicle will be easy. Also a compact designed vehicle should be outcome. Since it is a Single seated the driver comfort alone can be concentrated. Baja is an all-terrain vehicle and therefore lot of the obstacles will be presented in the ride way. The min height of these obstacles is considered and the suspension is de-swinged. This affects the chassis to have the ride height of min 10 inches. In order to reduce the centre of gravity of the vehicle towards the ground as low as possible the overall vehicle height must be lowered. But at the same time the driver must be in the vehicle very comfortably seated. In case of Roll over or heavy heaving scenario the top of the vehicle must not intrude the drivers head by any mean. Therefore the height of the vehicle is carefully designed. For easy turning and compact design the overall vehicle width is been restricted to 64 inches. Also the SAE rule doesn’t permit the vehicle outer to outer with to be more than 64 inches. From the suspension geometry design the nose of the roll cage must be restricted to 14 inches. This gives us an challenging design phase so that the driver ergonomics and the engineering constrains rebel. An all-terrain vehicle is mostly undergoes rough terrain accidental scenarios. Also Baja shall also be considered as a racing vehicle which will undergo accidents. In general the crash of the vehicle is a normal phenomenon. The main function of the roll cage is to protect the driver from any intrusion primarily and vehicle major components secondarily. The FEA analysis is to study and obtain a good result of a roll cage that will survive the worst crash. One of the most worst scenario of the all-terrain vehicle is that roll over .Due to the inconsistent terrain the vehicle roll over upside. It has a very major impact on drivers safety. FEA analysis is done to ensure roll over safety. Driver with such an all-terrain vehicle must be able to drive comfortably. It is dependent upon not only the suspension but ride height, vision of the driver , ride height, vibration , etc. Therefore ergonomics is considered. 8

CHAPTER-3 3. DESIGN PHASE 3.1. Fixing Minimum Dimensions Of Roll Cage The minimum dimensions of the roll cage were decided taking the driver into consideration. Since the primary task of the roll cage is to protect the driver in case of any accident, driver comfort ability was given paramount importance. The roll cage should be able to accommodate a person of height comfortably. The tallest member in the group was selected as the driver and the roll cage was designed taking the tallest member into consideration. The first task of the is to decide the seating position and then, using the anthropometric charts to further make the posture further suitable from ergonomics point of view.

3.2. Base Model Selection The task of designing the basic roll cage was taken up by the members of the design team. To ensure maximum number of ideas and different types of designs, the task of modelling was given to each and every group member. The design parameters were space considerations,

manufacturability,

safety

Features,

cost,

quality,

weight,

better

ergonomics, pleasing aesthetic looks as already stated in design goals. Also, a torso of the driver was modelled in accordance with the anthropometric charts developed. According to this chart, the lengths of different portions of the body can be approximated by taking the average from the table listed in the ergonomic, anthropometric section. There were three models suggested by the design team members, each had a different approach towards solving a design issue. The following are the three models presented. The figure 3.1, 3.2 and 3.3 are the chronological development of the roll cage design

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Figure 3.1 Model One

Figure 3.2 Model Two

Figure 3.3 Model Three To compare the models, and to come out with a final base model, a table of comparison was thought off. This table was made taking some parameters into consideration like weight, height of centre of gravity, etc. Also, results of the FEA analysis (presented in Preliminary FEA section) were also taken into consideration. 10

From the table 3.1, it was found, that all the models were almost comparable from the design point of view. Each of the models has its own advantage and disadvantages. But with the obtained results the selection procedure proceeded further. Table 3.1 Base Model Comparison. Parameter

Model 1

Model 2

Model 3

Mass

78 kg

72 kg

62 kg

Manufacturability

Easy

Moderate

moderate

CG( vertical location)

420.12 mm

520.89 mm

460.76 mm

Attractive design

Low

Fine

Good

Volume

6146987.5mm3

7739783.3mm3

8572709.82mm3

Driver ergonomics

Good

Moderate

Good

3.3. Ergonomics 3.3.1. Adequate Operator Space Dimensions of certain members were decided using standards published by the International Standards Organisation. These standards are: 

ISO 7250 Basic Human Body measurements for Technological Design.



ISO 3411 Human Physical Dimensions of operators and minimum operator space envelope.

These standards facilitated the weaving of a vehicle frame around the human body. One of the major design criterion that was used in the design of the chassis is the idea of driver ergonomics. Ergonomics is the study of how to layout and design the driver controls and safety Features of the car according to the needs of the driver in order to optimize human well being and overall system performance in a given situation. For the application of Baja, it was necessary to create a driver envelope that would not only fit the planned drivers, but allow comfort, safety, and stability to the driver for a period of time of up to higher driving hours. During This time period driver should not become tired or uncomfortable, while still remaining safely within the vehicle. 11

Factors that were taken into consideration were the seat location and inclination, the location of the steering wheel, the design of the foot box area so that the driver will be able to properly operate the vehicle in all driving scenarios. To allow for proper movement of the drivers feet to control the gas and brake pedal, the foot box was designed to be as small as possible while allowing the driver the proper amount of space to operate the controls safely. In an investigation of the 95th percentile man, it was found that the average man with shoes on has a foot width of roughly 4.5 inches. Because of these parameters, the final dimensions of the foot box were chosen to be 14 inches wide at the bottom and left relatively open at the top to allow each of our competition drivers to drive comfortably as well as allowing our other drivers to safely fit in the car for testing. The size and shape of the foot box also allowed for an optimum placement of the brake pedal and assembly low and between the lower frame members maintaining a low centre of gravity. The brake pedal and assembly was also extensively designed in order to maximize driver efficiency in a normal as well as panic braking scenario. The next area of the driver envelope that was investigated was the location of the steering wheel within the cockpit.The location of the steering wheel must be within a comfortable distance of the driver’s chest to prevent either driver interference or overextending of the arms. If the driver’s arms are over extended, the muscles in the arms will tire more rapidly as well as decreasing the amount of force that the driver can put into the wheel, which could lead to problems in an emergency situation. However, if the wheel is placed too close to the driver then the driver may not be able to exit the car within the required time of five seconds, even without the addition of a quick release steering wheel. Through a great deal of ergonomic research and physical testing, it was determined that the ideal location would be approximately 19.5 inches forward from the bottom of the rear roll hoop and 20 inches up from the lower frame members. This location will give the driver sufficient advantage to apply force to the wheel as well as allowing space for the driver to exit. Another rule that must be taken into account when designing the steering system is that the driver's wrists are attached to the steering wheel with safety straps. This has a major influence on the type of mechanism that is used to relay driver input to the steering by limiting the angle that the wheel can be turned. If the 12

wheel must be turned more than 90 degrees to reach the lock on the steering mechanism, and the driver is unable to adjust their hands, than this will create an extremely uncomfortable angle as well as increase the possibility of injury should an accident happen at full lock. The final ergonomic parameter that was considered was the location and inclination of the driver's seat. The seat is crucial to supplying enough support to the driver's back to allow him to stay upright with a clear view of the track ahead, to apply the proper forces to the gas and brake pedals, and support the shift in the driver's weight while cornering or landing from a jump. To properly determine the inclination of the seat a great deal of research was done both online and by a trial and error approach with many members of the team being tested. With this basic information, a physical test was conducted with a few members of the team and it was found that roughly 20 degrees of inclination was comfortable for those tested and was chosen. In addition to the seat back and inclination, aluminium bolsters are necessary in order to deal with the cornering forces that will be seen during the operation of the vehicle.

3.1.2. H-Point The H-point (or hip-point) is the theoretical, relative location of an occupant's hip, specifically the pivot point between the torso and upper leg portions of the body, either relative to the floor of the vehicle or relative to the height above pavement level and pertinent to seating comfort, visibility from the vehicle into traffic and other design factors. Itis shown in figure 3.4

Figure 3.4 H-Point Technically, the measurement uses the hip joint of a 90th percentile male occupant, viewed laterally.As with the location of other automotive design "hard points," the 13

H-point has major ramifications in the overall design of a vehicle, including roof height, aerodynamics, visibility, comfort, ease of entry and exit, interior packaging, safety, restraint and airbag design and collision performance. As an example, higher H-points can provide more legroom, both in the front and back seats. The Society of Automotive Engineers (SAE) has adopted tools for vehicle design, including statistical models for predicting driver eye location and seat position as well as an H-point mannequin for measuring seats and interior package geometry.

3.1.3. Anthropometry Anthropometry refers to the measurement of the human individual. It plays an important role in industrial design, clothing design, ergonomics and architecture where statistical data about the distribution of body dimensions in the population are used to optimize products. In addition, it is interesting to investigate the relationship between anthropometric data of the subjects and the driver’s seat adjustment with the comfortable postural angle measured.various driving postural angles are shown in figure 3.5 and 3.6

Figure 3.5 Driving Postural Angles

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Figure 3.6 Driver’s Anthropometric Measurements From the literature review different body heights persons were tested. Anthropometric measurements were made in the driver’s position according to Figure shown above and measurements are tabulated. 95% male heights hold its special consideration in the ergonomic analysis.

Figure 3.7 Driver’s Desired Seating Position The recorded data is compared with the result data of the literature review and the best position is selected from average of the following table 3.2.

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Table 3.2 Anthropometric Measurements Sizes

Anthropometric

Anthropometric

according measurements

in measurements

to Fig. 1

in Males(Body height in cm)

Females(Body height

in Average (cm)

cm) A

38.4

38.3

38.7

39.4

39.7

39.8

39.1

B

14.8

14.8

15

15.2

15.7

16.9

15.4

C

34.6

34.8

35.3

35.3

36.6

38.1

35.8

D

47.4

47.5

47.7

48.2

48.4

48.4

47.9

E

78.2

83.7

88

88.3

93.7

98.1

88.3

F

69.1

72.2

76.1

76.3

82.6

88.4

77.5

G

51.3

54.4

56.6

60.9

62.4

64.9

58.4

H

50.5

50.5

50.5

50.5

50.5

50.5

50.5

I

20.4

21.7

24.3

24

24.7

24

23.2

J

44

45.7

48.8

48.1

50.3

52.5

48.2

K

52.4

52.6

53.2

53.3

53.5

53.4

53.1

L

57.7

58.1

58.5

58.8

59.6

59.8

58.8

M

10.3

13.8

12.1

11.5

13.7

14.2

12.7

A manikin model from PROE is adjusted with the values as per the table above and a revised manikin model is developed which is shown in the figure 3.8, this model will be used for further analysis and design considerations.

Figure 3.8 Driver Manikin Model

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3.1.4. Building Prototype This model was again refurbished with the SAE guidelines and ready to be tested. A solid works model has been developed with updated model as shown in the figure3.9. Finally, in order to actually see the roll cage design in physical form, a 1:1 scaled pipe model was built using PVC pipes where the ends were fastened using cellophane tapes and the driver was made to sit inside it. Thus the comfort ability and ergonomics apart from aesthetics were observed. The objectives of building the mock-up were the following: 

To assess the preliminary design for driver ergonomics.



To improve the designs aesthetic senses.



To get a hands on feel of the design.



To visualize the serviceability of the design.

Figure 3.9 Preliminary Design CAD Model

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Figure 3.10 Building Mock Up Model

Figure 3.11 Checking Mock Up Model

The design was found to meet all the above objectives which is illustrated in the above figures.

3.2. SAE Guidelines SAE has laid down a set of guidelines and rules that every vehicle should follow. These guidelines are based on recommendations and tests conducted by design professionals. For creating a preliminary design these guidelines were followed to include members in the frame of the chassis. No additional members were added initially, so that the frame with the minimum weight is obtained.

3.3. Strengthening Mechanisms In an attempt to alleviate stress concentrations resulting from chassis geometry as well as improve the frame’s torsion stiffness, several strengthening mechanisms were incorporated into the design. These included gusseting at nodes and stressed body panels.

3.3.2. Gussets Gussets are pieces of sheet steel that are welded tangential to two tubes intersecting at a node. They reduce nodal stresses by distributing load farther down the intersecting members. They were added to joints that required bracing and in such a way that they would be loaded in this manner. The gussets are placed within 5 inches from the nodal point. The primary purpose of these gussets was to increase the roll cage safety factor and provide better protection for the driver in a roll over scenario. They also help to increase the

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overall frame stiffness which will benefit vehicle control and feel during normal or bumpy driving conditions.

3.4. Selection Of Final Model Based on the results of the comparison table which compared the mass, centre of gravity position etc. apart from the strengths and stress analysis results obtained from FEA, it was found that all the models had comparable performances. Finally model three is selected to serve as the base model. Additionally few other Features are also considered into the design with the guidance from the sae rule book. Those considerations are listed below. drivers safety guidelines are shown in figure 3.12

Figure 3.12 Driver Safety Guidelines  The driver’s helmet to be 15.24 cm (6 inches) away from top of the vehicle.  The driver’s torso, knees, shoulders, elbows, hands, and arms must have a minimum of 7.62 cm (3 in) of clearance from the envelope created by the structure of the car.  Roll cage members having a bend radius > 15.2 cm (6 inches) may NOT be longer than 71.1cm (28 inches) unsupported.  The RRH must be a minimum of 73.6 cm (29 in) wide at 68.6 cm (27 in) above the driver’s seat.  The start of bend c must be greater than 12 inches from the drivers helmet.  The SIM shall be between 20.3 cm (8 inches) and 35.6 cm (14 inches) (as measured vertically) above the area of the seat in contact with the driver  The angle between the FBMUP and the vertical should be less than 45 degrees.  These rules are also been explained detain in the section element of roll cage. 19

3.5. Geometry Of Roll Cage Design The general frame geometry as well as individual tube geometries were given considerable design consideration to throughout the chassis optimization process. The overall frame size was minimized as much as possible in order to reduce weight. The firewall area was designed to meet the SAE rule book specification which states it must be greater than 29 inches wide at 27 inches high measured vertically from the driver’s seat, but was designed to taper off at the top and bottom in order to minimize the surface area and thus reduce wind drag. Members that are primarily loaded in tension or compression and thus maximizing their effectiveness in force resolution is resolved. Specific tube geometries such as the front and rear damper mounts were designed to best handle the loads that would be imposed upon them. The curved hoop geometry of the rear damper mount member was modelled after the typical shock tower design found in many desert buggy designs. The intention is to resolve the immediate damper force in the one member such that only smaller bracing members are required for additional stiffness. The hoop design places Both sides of the damper mount member in tension and thus requires a tube with less cross Sectional area to resolve the load. The front damper mount was designed to feed loads directly into the front hoop, a shape that is both strong and rigid. The design of the frame is defined by the design safety rules set out by the SAE. First, the rules set specific requirements on the building material’s type and geometry. They also define the specific requirements of the frame geometry. The requirements were referenced when making decisions regarding the material selection, design geometry and any additional modifications to the design. It is important that the reader understand that these constraints were in place during the design of the vehicle frame, as well as the interaction between the frame design and other factors, such as drivetrain, suspension, and driver safety and restraint. The usage of parametric design was extremely important with this design. As so many factors interact in the design of the frame, the parametric properties allowed the change of a single part to automatically change the design of all parts interacting with it. This is especially important when ensuring the roll cage envelope considerations as well as the suspension and drive train mounting.

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CHAPTER-4 4. DESIGN RESULTS 4.1. Elements Of Roll Cage

Figure 4.1 Elements of Roll Cage 

Rear Roll Hoop (RRH).



Rear Hoop Overhead Members (RHO).



Front Bracing Member (FBM).



Lateral Cross Member (LC).



Front Lateral Cross Member (FLC).



Lateral Diagonal Bracing (LBD).



Lower Frame side (LFS).



Side Impact Member (SIM).



Fore / Aft Bracing (FAB).



Under seat Member (USM).



All other required cross members.



Any tube used to mount safety belts.

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4.1.1. Rear Roll Hoop (RRH) The RRH is a structural panel behind the driver’s back, and defines the back side of the roll cage. The driver and seat must be entirely forward of this panel. The RRH is substantially vertical, but may incline by up to 20 degree from vertical from points A to points B towards rear side in parallel with seat backrest. The RRH is made up of a maximum of four sections: two LCs at highest and lowest points and two continuous, no break vertical members. This may be one continuous hoop/tube. The driver’s seat may not intrude into the plane(s) of the RRH. The RRH must be a minimum of 73.6 cm (29 in.) wide at 68.6 cm (27 in.) above the driver’s seat (H-point of driver).the RRH is shown in figure 4.2.

Figure 4.2 Rear Roll Hoop 4.1.2. Rear Roll Hoop Lateral Diagonal Bracing (LDB) The RRH must be diagonally braced as in figure 4.3. The diagonal brace(s) must extend from one RRH vertical member to the other. The top and bottom intersections of the LDB members and the RRH vertical members shall begin at a point along the vertical portion of the RRH within 12.7 cm (5 in.) vertically of point BR or BL and extend diagonally to a point no farther than 12.7 cm (5 in.) above point AR or AL respectively. The angle between LDB members and RRH vertical members must be greater than or equal to 20 deg. Lateral bracing may consist of more than one member.

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Figure 4.3 Lateral Diagonal Bracing

4.1.3. Roll Hoop Overhead Members (RHO) The forward ends of RHO members (intersection with the LC) define points CR and CL. Points CR and CL must be at least 30.5 cm (12 in.) forward of a point, in the vehicle’s elevation view, defined by the intersection of the RHO member and a vertical line rising from the after end of the seat bottom. This point on the seat is defined by the Seat bottom intersection with a 10.1 cm (4 in.) radius circle which touches the seat bottom and the seat back. The Top edge of the template is exactly horizontal with respect to gravity.

4.1.4. Lower Frame Side Members (LFS) Lower frame side members defines the lower right & left edges of roll cage & joins RRH LC and the points forward of The driver’s heel to a front lateral cross member (FLC).

4.1.5. Side Impact Members (SIM) The two Side Impact Members define a horizontal mid-plane within the roll cage. These members are joined to the RRH at points S and extend horizontally to points SF forward of the driver’s toes. The SIM must be between 20.3 cm (8 in.) and 35.6 cm (14 in.) (as measured vertically) above the area of the seat in contact with the driver.

4.1.6. Under Seat Member (USM) The two LFS members must be joined by the Under Seat Members. The USM must pass directly below the driver where the template in RC3 intersects the seat bottom. The

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USM must be positioned in such a way to prevent the driver from passing through the plane of the LFS in the event of seat failure.

4.1.7. Front Bracing Members (FBM) Front bracing members (continuous, with no break) must join RHO, SIM and LFS. The upper Front Bracing Members must join points C on the RHO to the SIM at or behind points SF. The lower Front Bracing Members must join points AF to points SF. The FBM must be continuous tubes. The angle between the FBMUP and the vertical must be less than or equal to 45 degrees.

Figure 4.4 Front Bracing Members

4.1.8. Roll Hoop Bracing (FAB) The hoop must be braced on both right and left sides. From a side view, the bracing must be triangulated, with the maximum length of 101.6 cm (40 in.) between attachment points. The angles of the triangulation must be minimum 20 degrees. A bent tube cannot exceed 81.3 cm (32 in.) between attachment points.

4.1.9. Front Bracing If front bracing is used it must connect FBM up, LFS and the SIM. Front bracing must be attached as close as possible to the top of the roll cage.

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4.1.10.

Rear Bracing If rear bracing is used it must be attached as close as possible to the top of the roll

hoop along the outer perimeter. The bracing must be triangulated and connect back to the RRH below the SIM.

4.1.11.

RHO/FBM Gusseting

If the RHO and FBM are not fabricated from a continuous tube, a gusset is required at point C. Gussets shall be made of steel plate, be triangular from a side view, have length at least 3 times the tube diameter and have a minimum thickness of 3.2mm (0.13 in.) The gussets shall be welded to the sides of the tubes and not directly in the plane of the tubes making up each joint as shown in figure 4.5.

Figure 4.5 Gusseting

4.2. Design Specifications 4.2.1. Geometry Table 4.1 Geometry Object Name

Roll cage

State

Fully Defined

Definition Type

DesignModeller

Length Unit

Meters

Element Control

Program Controlled

Display Style

Part Colour

25

BoundingBox Length X

0.7891 m

Length Y

1.251 m

Length Z

1.9219 m

Properties Volume

6.4681e-003 m³

Mass

49.804 kg

Scale Factor Value

1

Statistics Bodies

40

Active Bodies

40

Nodes

145387

Elements

71573

Mesh Metric

None

4.2.2. Vehicle Design  Attractive to customers  Easily operated  Easy to manufacture and maintain  Safety to the highest degree

4.2.3. Dimensions and Configurations  Have four wheels  Carry one person 190 cm (6’3”) tall weighing 113 kg (250 lbs.) at minimum  No more than 162 cm (64 in) at the widest point.  90 inch maximum length for vehicle; however, SAE recommends the car be no longer than 108 inch. in length.

26

4.2.4. Capability  Safe operation over rough land terrain, including but not limited to rocks, sand, jumps, steep inclines, mud and shallow water  Safe operation over any type of weather including but not limited to rain, snow, or ice  Adequate ground clearance, minimum 10 inches

4.2.5. Engine  Briggs & Stratton 10 HP OHV Intek Model

4.3. Functional Specifications This portion of the design describes the objective of the project as the function. The sub functions of this project describe the constraints set forth by the Society of Automotive Engineers, the engineering authority proposing the competition. The means of the functional analysis describes the processes that were considered in the design of the MiniBaja.

27

CHAPTER-5 5. SELECTION OF BUILDING MATERIALS 5.1. Tube Size Selection The lesser the diameter the lesser the weight, therefore it is better to choose the least available diameter which satisfies all the design considerations. But at the same time large diameter thin walled tube will increase strength by decreasing weight comparatively. Lot of consideration involve in the selection of tube size selection and thickness. a circular tube of the same thickness and same dimensions i.e. 25 dia as against 25 square will be stronger in all respects save for bending. A square tube of same dimension will be thinner to achieve equal mass and may not necessarily be stronger in bending. The primary advantage of round tubing (besides the slightly higher UTS) is that it is slightly more efficient in resisting buckling per unit weight.

5.2. Tube Shape Selection The most generally available tubes shapes are round, square and rectangle. The round shape is used for the space frames. Round is stronger than square in compression and torsion. Tension and shear they are about equal and bending varies based upon where the force is applied. To be specific with bending forces, square is stronger when the force is applied in line with an edge (the tube acting basically like a fully boxed I-beam) whereas it's weaker when the force is applied 45 degrees to the edge, in which case the square tube tends to flatten and lose strength accordingly.The geometry is also limited by industry standards. It is important to utilizecommonly available tubing sizes and materials. Tubing is available in standard fractional sizes to the 1/8th of an inch: 1, 1.125, 1.25, 1.375, and 1.5. The wall thickness is limited to the common tubing Gauges. In this case these are: 0.035, 0.049, 0.058, 0.065 and 0.083 inches.

28

CHAPTER-6 6. MATERIAL SELECTION 6.1. Materials The key design decisions of our frame that would greatly increase safety, reliability and performance is material selection. To ensure that we chose the optimal material, we did extensive research and compared materials in multiple categories. Our key categories for comparison were strength, availability, weight, and cost. Design considerations aside, the driving factor behind chassis material selection were the SAE competition vehicle regulations. The materials used for the cage must meet certain requirements of geometry as set by SAE, and other limitations as decided by team. Steel is a ubiquitous material choice for in mass produced chassis, custom auto racing roll cages, and other SAE Baja car frames because of its high strength, low cost, and high Weldability. Steel is very responsive to thermal processing which provides for a higher strength to weight ratio and thus less material is required for construction. After the base material had been selected, the team then had to choose which alloy would best suit the vehicle requirements. The SAE Rule Book uses AISI 1018 steel properties as a base for many of their required strength and stiffness equivalencies, so this alloy was considered first. 1018 is a very common alloy that is cheap and readily available in multiple geometries and wall thicknesses. A maximum carbon content of .30% is required for good Weldability, so the low carbon content of .18% in the 1018 steel was acceptable. As the frame is used in a racing vehicle, weight is a crucial factor and must be considered. The proper balance of fulfilling the design requirements and minimizing the weight is crucial for a successful design. The rules define the roll cage to be made with materials equivalent to the following specification. Circular steel tubing with an outside diameter of 25mm (1 in) and a wall thickness of 3 mm (0.120 in) and a carbon content of at least 0.18%. A steel tube with bending stiffness and bending strength exceeding that of circular steel tubing with an outside dia of 25mm and a wall thickness of 3 mm and a carbon content of 0.18%.The rules go on further to define bending strength and stiffness by. Bending stiffness is proportional by the EI product and bending strength is given by the value of ᵠ I/c. 29

MATERIAL

OUTER

THICKNESS

DIAMETER

MOMENT OF

BENDING

BENDING

INERTIA

STRENGTH

STIFFNESS

MASS

MM

MM

MM4

N/MM2

N/MM2

KG

ASTM A252

25.4

2.1

10514.6593

360147.7791

2155505156

39.05978

ASTM A252

31.75

2.1

21600.36365

591883.9804

4428074549

49.70482

ASTM A252

38.1

2.1

38600.82947

881436.2635

7913170041

60.34987

ASTM A252

44.45

2.1

62782.79318

1228819.574

12870472602

70.99492

ASTM A252

50.8

2.1

95412.99122

1634041.385

19559663201

81.63996

ANSI 4130

25.4

2.1

10514.6593

360147.7791

2155505156

39.05978

ANSI 4130

31.75

2.1

21600.36365

591883.9804

4428074549

49.70482

ANSI 4130

38.1

2.1

38600.82947

881436.2635

7913170041

60.34987

ANSI 4130

44.45

2.1

62782.79318

1228819.574

12870472602

70.99492

ANSI 4130

50.8

2.1

95412.99122

1634041.385

19559663201

81.63996

AISI 1018

25.4

2.1

10514.6593

360147.7791

2155505156

39.05978

AISI 1018

31.75

2.1

21600.36365

591883.9804

4428074549

49.70482

AISI 1018

38.1

2.1

38600.82947

881436.2635

7913170041

60.34987

AISI 1018

44.45

2.1

62782.79318

1228819.574

12870472602

70.99492

AISI 1018

50.8

2.1

95412.99122

1634041.385

19559663201

81.63996

IS 3074

25.4

2.1

10514.6593

360147.7791

2155505156

39.05978

IS 3074

31.75

2.1

21600.36365

591883.9804

4428074549

49.70482

IS 3074

38.1

2.1

38600.82947

881436.2635

7913170041

60.34987

IS 3074

44.45

2.1

62782.79318

1228819.574

12870472602

70.99492

IS 3074

50.8

2.1

95412.99122

1634041.385

19559663201

81.63996

ASTM 106B

25.4

2.1

10514.6593

360147.7791

2155505156

39.05978

ASTM 106B

31.75

2.1

21600.36365

591883.9804

4428074549

49.70482

ASTM 106B

38.1

2.1

38600.82947

881436.2635

7913170041

60.34987

ASTMA106B

44.45

2.1

62782.79318

1228819.574

12870472602

70.99492

ASTM 106B

50.8

2.1

95412.99122

1634041.385

19559663201

81.63996

Table 6.1 Material Strength Comparison

30

6.2. Chassis Material Optimisation These ruling left the team with only few options. The comparison chart of various materials that we analysed under the rules specified by BAJA SAE and other factors such as Weldability and availability etc.are as follows in the table 6.2 Table 6.2 Material Properties MATERIAL

CHROMOLY4130 ASTM A252

AISI

IS 3074

1018 Carbon content( % ) Yield

ASTM A106B

0.32

0.18

0.18

0.20

0.30

strength 395

350

365

372

383

strength 560

455

450

473

466

(MPa) Tensile (MPa) Elongation (%)

25

20

20

5

20

Cost in rupees /m

550

650

600

800

500

Availability

Easy

Medium

Easy

Difficult Easy

With these available data in hand, a decision matrix was created for selecting the optimized material which led to the following graph as in figure 6.1 and 6.2 Table 6.3 Decision Matrix MATERIAL

Chromoly

Astm

Aisi

Is

Astm

PARAMETER

4130

a252

1018

3074

a106b

Carbon content ( % )

5

1

1

3

4

Yield strength(MPa)

5

1

2

3

4

Tensile strength (MPa)

5

2

1

4

3

Elongation (%)

5

2

2

1

2

Cost in rupees /meter

2

4

3

5

1

Availability

1

3

5

1

5

TOTAL

23

3

4

17

19

31

Figure 6.1 Material Comparison

Figure6.2 Material Parameter Comparisons From this matrix in Fig 6.1 and 6.2, we conclude that the best available material is Chromoly AISI 4130.

32

CHAPTER-7 7. FEA MODEL 7.1. Stress Analysis In order to ensure the structural integrity of the chassis and driver safety through the rigors of competition, a 3D mathematical model of the frame was created and then tested using finite element analysis software. The finite-difference computational simulation created utilized to calculate damper loads for predicted jump landing heights. In addition to ensuring chassis integrity, the FEA analysis was also utilized to refine the frame stiffness and minimize the overall weight.

7.2. Load Scenario Determination Before attempting to calculate the input loads, it was necessary to determine what kind of obstacles would be encountered during vehicle operation. This was done by reviewing test and competition videos of other SAE Baja vehicles on the internet as well as driving the past years vehicle. This allowed the team to get a very broad sense of what terrain the vehicle would be required to overcome. The vehicles were regularly seen going off of jumps and landing from one or more vertical meters in the air. Given the repeated success and toughness of the old car which was designed for 1 meters landings, it was determined to use similar criteria.It was determined that the 3G side load used in the 2012 design was also a relevant and necessary Loading scenario in case the car should be drifting sideways around a corner and have either the front or rear outside wheel come into contact with a rock or tree. Finally, it was deemed appropriate to include a one wheel landing scenario as reviewing multiple test videos revealed that many jumps are uneven and cause the vehicles that launch of them to roll in the air and land un-evenly on one side. The one wheel landing and 3G sides loading were thought of as worst case scenarios and therefore the resulting peak design stresses are limited by the yield strength instead of the fatigue limit.

33

From the load scenario determination various test that to be conducted in the finite element analysis are 1. Front impact test 2. Rear impact test 3. Side impact test 4. Heave test 5. Roll over test 6. Front bump test 7. Rear bump test 8. Twist ditch The explanation and calculated force are explained in the input load determination section.

7.3. Determination Of Centre Of Mass In order to check the likelihood of the Mini Baja to undergo tripped rollover, the height of the centre of mass of the vehicle is required. This was found by treating assigning various materials to the components and calculating the centre of mass of the full vehicle assembly using the option “Mass Properties” in SOLIDWORKS as shown in figure 7.1.

Figure 7.1 Centre of Gravity X=-0.88 inch(approximated to 0) Y=481.20 inch Z=153.06 inch. 34

7.4. Input Load Determination Front Impact Test Consider two vehicles approaching each other opposite direction. To solve this consider the each vehicle separately dashes on a fixed obstacle like 1st type as shown in figure 7.2. Here mass of vehicle 1=MA, mass of vehicle 2=Mb

Figure 7.2 Front Impact Scenario

Force on vehicle 1 and 2is given by FA=MA*aA, FB=MB*a Since both forces acts in opposite to each other FA=-FB= Vr= VA+VB (+ sign since the vehicle moves towards each other) Vr=15.2778+0=15.2778 m/s F=

=19839.94 N

F=19839.94N.

35

Rear Impact Test

Figure 7.3 Rear Impact Scenario

Similar to that of front impact test type2 but the vehicle 1 is stationary as shown in figure 7.3 Vr= VA+VB (+ sign since the vehicle moves towards each other but VA=0) Vr=0+15.2778=15.2778 m/s FA=-FB= Vr= VA+VB (+ sign since the vehicle moves towards each other) Vr=15.2778+0=15.2778 m/s F=

=19839.94 N

F=19839.94N

36

Side Impact Test Very similar to rear impact but the position of vehicle changes as shown in figure 7.4

Figure 7.4 Side Impact Scenario

Vr= VA+VB (+ sign since the vehicle moves towards each other but VA=0) Vr=0+15.2778=15.2778 m/s FA=-FB= Vr= VA+VB (+ sign since the vehicle moves towards each other) Vr=10+0=10 m/s F=

=10491.7285 N

F=10491.7285N

37

Heave Test When all the four wheel drop or rise simultaneously it is called as the heaving. So dropping the vehicle from a height (drop test)is performed as showed in figure.7.5

Figure 7.5 Heave Test Scenario Potential energy (P.E)=Mgh Thereby this potential energy converts into kinetic energy when dropped and reaches the ground with impact velocity Impact velocity V=√ V=√

ie dropped from a height of 1 m

V=4.427 m/s K.E = K.E = K.E= 3429.70 Considering the vehicle a free falling object

Therefore s=0.99 38

Force = K.E/s=3433.49 F=3433.49 N

Roll Over Test Condition for vehicle to roll over dynamically. Figure 7.6 shows various forces acting on the vehicle.

Figure 7.6 Roll Over Forces The vehicle rolls about the pivot point of the wheel The moments of the forces in the system must be in equilibrium to avoid side roll over Centrifugal force * perpendicular distance from C.G= Weight force * perpendicular distance from C.G Determining the safe cornering speed to avoid roll over in a curvature of 3m turning radius(3mts max turning radius) *d(c.g )=m*g* d(c.g) Substituting the values and v is derived=24.98Kmph or 6.939 m/s Critical angle of slope and height of adjacent wheel to be up for static side rollover

39

Figure 7.7 Horizontal Roll Over Scenario From trigonometric relation

=31.430 =90- =58.5698



H=52.90in or 1343.73 mm Critical Angle Of Slope And Height Of Adjacent Wheel To Be Up For Static Slope Rollover

Figure 7.8 Vertical Roll Over Scenario 40

From trigonometric relation

=36.345 =90- =53.65



H=51.54 in or 1309.27 mm

Rollover Forces To analyse force consider vehicle falling upside down from a height of 1 meters as shown in figure 7.9.

Figure 7.9 Roll Over Force Potential energy (P.E)=Mgh Thereby this potential energy converts into kinetic energy when dropped and reaches the ground with impact velocity Impact velocity V=√

41

V=√

ie dropped from a height of 1 m

V=6.264 m/s K.E = K.E = K.E= 6670.62 Considering the vehicle a free falling object

Therefore s=0.99 Force = K.E/s=6738 F=6738 N

Front Bump And Rear Bump Forces

Figure 7.10 Front And Rear Bump Scenario An assumption is made that when the vehicle passes over a bump, the entire weight of the vehicle will turn into two point loads at the two points where the wheel force is transmitted to the chassis, through the suspension as shown in the figure 7.10. The worst case 42

will be when the suspension fails and the entire force is transmitted. As the requirement is not for the Chassis to fail in case the suspension fails. These two point loads will be equal to the weight of the chassis. For low circuit speed race cars the 2 g force considerations are used Hence, 2F = m *2g F=½m*g F= ½ * 340 * 9.81*2 F = 3335.4 N

Twist Ditch Or Torsion Rigidity

Figure 7.11 Twist Ditch Scenario Apply bump force to one wheel only F=3335.4

7.5. Element Attribute 7.5.1. Elements 3D Elastic Beam Elements (Beam4) Tetrahedral Elements 4 Nodes 43

7.5.2. Material Properties Linear, Elastic, Isotropic Exy=2.05E5 , ν= 0.30

7.5.3. Element Properties AREA = 8.03E-5 IZZ

= 5.96E-9

IYY

= 5.96E-9

TKZ

= 0.021

TKY

= 0.021

IXX

= 1.19E-8

ADDMASS = 6.153

7.6. Finite Element Modelling In order to carry out a design check of the preliminary designs developed by the design team in Finite Element Analysis, a finite element model was developed using the package ANSYS workbench. The geometric model in Solid works was converted into STEP format which was then imported in ANSYS.

7.7. FEAAssumptions And Consideration 

Element category- h-element



Element shape- tetrahedral



Element type – beam



FEAanalyze approach- Top down approach



Element character- solid



Nodes/element- 4 nodes



result type – Von- misses stress and total deformation



All weldments are assumed as parent material

44

CHAPTER-8 8. PRELIMINARY FINITE ELEMENT ANALYSIS The 3D chassis model was created using Solid Works and the FEA was performed using the Solid works simulation analysis. All analyses were run using a Beam Mesh which partitions the annular frame members into tiny sections and then calculates the load distribution and resulting stresses. The beam mesh is a very convenient Feature of Solidworks simulation analysis because the calculation time is very short and it allows for multiple iterations very quickly. This is not the accurate and required result but helped in the study of the force distribution through the members of the roll cage. The finite element analyses of the preliminary stage are 

Beam element



Axial and bending strength



Free coarse mesh type.

In order to account for and minimize the possible inaccuracies, the mesh size was refined as small as possible. The boundary conditions for the chassis simulations were defined by constraining all suspension mounts opposite to the end being loaded. This was to ensure a worst case loading scenario by providing the largest possible lever arm between the reaction and load points. Constraining the frame at the centre of mass was also considered, but this method would incur additional unrealistic stress concentrations due to the unrealistic associations the software would assume between frame nodes and the CM. The CM constraint would influence load propagation to occur relative to one point instead of a section of members at the opposite end of the vehicle and thus the calculated displacements and stresses would be inaccurate. As mentioned before, the control arms were included into the model and the loads were input at nodes perpendicular to ends of the members

. 45

8.1. Processing Preliminary FEA For the preliminary FEA the following conditions are used Table 8.1 Preliminary FEA Model used:-

Fullmodel

Mesh type:-

Beam

Mesh control:-

Controlled

Total nodes:-

1535

Total elements:-

1344

Front Impact Test Loading= 19839.94 N

Figure 8.1 PFEAFront ImpactDisplacement Result Max stress=202.9 N/mm2 Factor of safety=1.922

46

Figure 8.2 PFEAFront Impact Stress

Rear Impact Test Loading= 19839.94 N

Figure 8.3 PFEARear ImpactDisplacement

Figure 8.4 PFEARear Impact Stress

Result Max stress=127.2 N/mm2 Factor of safety=3.066

Side Impact Test Loading= 10491.7285 N

Figure 8.5 PFEASide Impact Displacement Result Max stress=239.3 N/mm2 Factor of safety=1.629

47

Figure 8.6 PFEASide Impact Stress

Heave Test Loading= 3433.49 N

Figure 8.7 PFEA Heave Test Displacement

Figure 8.8 PFEA Heave Test Stress

Result Max stress=69.5 N/mm2 Factor of safety=5.611

Roll Over Test Loading= 6738 N

Figure 8.9 PFEARoll OverDisplacement

Figure 8.10 PFEARoll Over Stress

Result Max stress=60.5 N/mm2 Factor of safety=6.44

48

Front Bump Test Loading= 3335.4 N

Figure 8.11 PFEAFront BumpDisplacement

Figure 8.12 PFEAFront Bump Stress

Result Max stress=70.1 N/mm2 Factor of safety=5.7

Rear Bump Test Loading= 3335.4 N

Figure 8.13 PFEARear BumpDisplacement

Result Max stress=67.21 N/mm2 Factor of safety=5.4

49

Figure 8.14 PFEARear Bump Stress

Twist Ditch Loading= 3335.4 N

Figure 8.15 PFEATwist DitchDisplacement

Figure 8.16 PFEATwist Ditch Stress

Result Max stress=123.1 N/mm2 Factor of safety=3.168

8.2. Preliminary Analysis Result Conclusion The peak displacement discrepancies are a result of the included suspension members because in some loading scenarios they may experience a greater overall deflection than the chassis members. The peak stress discrepancies are a result of stress concentrations produced by inaccurate node geometry and therefore imprecise stress distributions. These analysis inaccuracies are intrinsic to the beam mesh, software restraint, and solution methods utilized by Solid Works and are therefore believed to be highly inflated from what would actually occur. In cases where these untrue peak stresses occurred, the highest average probed stress was taken to be true. The most problematic analyses were the roll over and torsion rigidity. In these scenarios the loads propagated from the rear of the chassis to the front and large stress concentrations developed at the bends in the side impact and front damper mount members. These stress concentrations were a result of the curved geometry inciting a large stack up error in the nodal stress calculations and were accounted for by probing average surrounding values. Even though the nodal stress concentrations were considered to be an erroneous part of the finite element analysis, during construction the nodes where these

50

singularities occurred were heavily reinforced with gussets to ensure that a catastrophic failure would not occur. Another important factor to discuss is the lack of correlation between peak stresses and displacement. That is to say that peak stresses do not always occur at the location of a peak geometrical displacement. This is because the peak displacement often occurs far away from the restraint points due to the incremental displacement of intermediate frame members, essentially a result of stack up. Frame member displacements vary widely depending on their length, shape, and resultant load vector. For example, if a long member is placed in bending such that it acts like a cantilever, the results will yield a relatively large displacement at the unrestrained end, but the member may not experience any high resultant stresses. The areas of peak stress often occur at very stiff joints. This is because the noncompliant joints must resist the imposed load to effectively transmit it throughout the surrounding chassis members. Thus the joints become point stresses and can result in massive stress concentrations that are often times not present in the actual system being modelled. The lack of correlation between displacement and peak stresses realization was important because it provided insight into frame member allocation and allowed the team to better refine the overall chassis stiffness matrices thus maintaining an even load distribution.

51

CHAPTER-9 9. FINAL FEA ANALYSIS 9.1. Pre Processor The final FEA uses the similar kind of the elements approach as of in the preliminary FEA but uses a dedicated software Hypermesh for the meshing the roll cage. The element size and the attributes are controlled to obtain good result which are checked using the quality index. Figure 9.1 and 9.2 shows the screen shot of the mesh quality and full mesh view.

Figure 9.1 Quality Index of Mesh After the file is imported into the Ansys and analysis is done

Figure 9.2 Roll Cage Mesh In Hypermesh

52

9.2. Processor And Results Front Impact Test • Model Used: Full Model • Loading: F= 19839.94 N on front Corner points. • Boundary Conditions: Rearcorner points, All DOF=0

Figure 9.3FFEAFront ImpactDisplacement

Figure 9.4FFEAFront Impact Stress

Results Max Stress= 341.73 MPa

Min stress=379.7 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 1.272935 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Rear Impact Test • Model Used: Full Model • Loading: F= 19839.94 N on Rearcorner points. • Boundary Conditions: Front Corner Points, All DOF=0

53

Figure 9.5FFEARear ImpactDisplacement

Figure 9.6FFEARear Impact Stress

Results Max Stress= 292.42 MPa

Min stress=302.05 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 1.487586 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Side Impact Test • Model Used: Full Model • Loading: F= 10491.7285N on right side outer most Corner four points. • Boundary Conditions: on Left side outer most Corner four points., All DOF=0

Figure 9.7FFEASide ImpactDisplacement 54

Figure 9.8FFEASide Impact Stress

Results Max Stress= 282.61 MPa

Min stress=264.01 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 1.539224 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Heave Test • Model Used: Full Model • Loading: F= 3433.49 N on centre of gravity point. • Boundary Conditions: on suspension points. All DOF=0

Figure 9.9FFEAHeave Test Displacement

Figure 9.10FFEAHeave Test Stress

Results Max Stress= 53.95 MPa

Min stress=65.33 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax = 8.063021 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Roll Over Test • Model Used: Full Model 55

• Loading: F= 6738 N on top frame points. • Boundary Conditions: base frame six points, All DOF=0

Figure 9.11FFEARoll Over Displacement

Figure 9.12FFEARoll Over Stress

Results Max Stress= 33.2 MPa Min stress=28.8 MPa Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax = 13.10241 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Front Bump Test • Model Used: Full Model • Loading: F= 3335.4 N on top frame four points. • Boundary Conditions: Front suspension, All DOF=0

Figure 9.13FFEAFront BumpDisplacement 56

Figure 9.14FFEAFront Bump Stress

Results Max Stress= 257.26 MPa

Min stress=154.66 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 1.690896 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

Rear Bump Test • Model Used: Full Model • Loading: F= 3335.4 N on top frame four points. • Boundary Conditions: Rear suspension, All DOF=0

Figure 9.15FFEARear BumpDisplacement

Figure 9.16FFEARear Bump Stress.

Results Max Stress= 129.66 MPa

Min stress=132.05 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 3.354928 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact.

57

Twist Ditch • Model Used: Full Model • Loading: F= 3335.47 N on suspension pointsalternate in direction • Boundary Conditions: Rear Corner Points, All DOF=0

Figure 9.17FFEATwist DitchDisplacement

Figure 9.18FFEATwist Ditch Stress

Results Max Stress= 154.26 MPa

Min stress=46.3 MPa

Factor of Safety: Incorporated Factor of Safety = ᵠt/Smax= 2.819914 FOS is greater than 1.25 Hence, the Chassis will be safe under font Impact

9.3. Conclusion Of Finite Element Analysis After the analysis of results, additional bracings were added to the frame mainly to the RRH and RHO. The performance after addition of bracings was found to be satisfactory. The addition of bracings further strengthens the structure by increasing the effective area under load, thus reducing the intensity of pressure. Also, it was decided to include gussets in the weld so that the effective welding area increases. The factor of safety taken is 1.25 here to account for the uncertainties in the assumptions. Following this and the suitable design changes, the model was found to be satisfactory in the strength aspect and thus enhancing the driver safety. 58

CHAPTER-10 10.CHASSIS OPTIMIZATION Chassis optimization was concurrent to the finite element analysis and served to optimize the safety, stiffness, stress distribution, and weight of the frame.In order to optimize the tubular structure, specific tubes or areas were focused on during each FEA simulation. The SAE rule book and driver envelope specifications (ergonomics) dictated the placement of many of the main structural tubes, so the bulk of refinement occurred with tube size, wall thickness, and location of bracing members. During each analysis, the areas of high stress were concentrated on and alleviated by adding bracing members to tune the stiffness and redistribute load to other areas. Additionally, the tube size and wall thicknesses were adjusted until they were all uniformly stressed. This was a difficult process because as the chassis was optimized for one loading scenario, the changes made affected the stress distributions in the other load cases. Accordingly, the front and rear jump landing load scenarios were accounted for first by ensuring the maximum resultant stresses were below the fatigue limit. Next, the worst case load scenarios were designed for by increasing the tube size until the maximum stresses were below the minimal predicted yield stress.

10.1.Result Of Chassis Structural Optimisation Front Impact Test

Figure10.1 FrontImpact Optimisation 59

Table 10.1 Front Impact Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.33549 kg

Optimized Mass

69.383 kg

Rear Impact Test

Figure10.2 Rear Impact Optimisation Table 10.2 Rear Impact Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.48536 kg

Optimized Mass

69.154 kg

60

Side Impact Test

Figure10.3 SideImpact Optimisation

Table 10.3 SideImpact Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.49597 kg

Optimized Mass

69.904 kg

Heave Test

Figure10.4 Heave Test Optimisation

61

Table 10.4 Heave Test Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.49079 kg

Optimized Mass

69.752 kg

Roll Over Test

Figure 10.5 Roll Over Optimisation Table 10.5 Roll Over Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.38712 kg

Optimized Mass

70.594 kg

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Front Bump Test

Figure10.6 Front Bump Optimisation Table 10.6 FrontBump Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.69937 kg

Optimized Mass

68.863 kg

Rear Bump Test

Figure 10.7 Rear Bump Optimisation

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Table 10.7 Rear Bump Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.40593 kg

Optimized Mass

69.649 kg

Twist Ditch

Figure 10.8 Twist Ditch Optimisation Table 10.8 Twist Ditch Optimisation Definition Target Reduction 20. % Results Original Mass

80.584 kg

Marginal Mass

0.51498 kg

Optimized Mass

69.351 kg

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10.2.Result Of Structural Analysis From the chassis optimisation using ANSYS it’s been determined to use two different members along the roll cage thus bringing the weight lowered. The consideration of all other factors for designing the roll cage is also made. The final model of the varied roll cage tube dimensions are shown in figure 10.9.thus a weight of 62.215 kg is achieved.

Figure 10.9 Optimised Roll Cage Table 10.9 Optimised Roll Cage Primary member (shown in blue)

Secondary member( shown in white)

Outer diameter=33.38 mm

Outer diameter=31.75 mm

Thickness=3.38 mm

Thickness=2.1 mm

The above table 10.9 gives the final suitable diameter and wall thicknesses of the tubular members that can be used in the Roll cage.

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CHAPTER-11 11.MANUFACTURING STRATEGY 11.1.Fabrication One of the criteria that had to be taken into consideration by the team throughout the design process was the manufacturability of each design feature incorporated into the vehicle. This was because the team felt that it was necessary to fabricate as much of the vehicle in house by utilizing the college manufacturing facilities in order to best fulfil requirement. The total range of abilities required to fabricate the car were fixturing, cutting, bending, notching, fitting, welding, and machining. All of these skills were represented by different team members which made collaboration and team work vital for the successfully completion of the car.

11.2.Chassis Fabrication The chassis fabrication process enlisted all of the different skills possessed by each individual team member. The team worked diligently to setup the frame table and necessary fixturing in order to ensure that the chassis geometry would be accurate to the 3D model after completion. Tube notching, fitment, and cleanliness were a top priority and guaranteed good weld penetration as well as maximum joint strength. Quality was always a top priority for the team throughout the entire chassis build and because of this it was sometimes necessary to alter the design during construction in order to meet certain time constraints.completely fabricated vehicle is shown in the figure11.1.

Figure 11.1 Fabricated Chassis

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11.3.Frame Table and Coordinate system The first step to assembling the chassis was to acquire a frame table as in figure 11.2. The table was made up of t thick plywood sheet. A coordinate system was drawn on the table top in order to properly orient the fixtures and jigs as well as accurately position frame members on the table. The coordinate system was drawn with the base plane at the bottom of the roll hoop as well as the central axis of the car. which was used for both the chassis manufacturing as well as accurately locating jigs such as the one for suspension mounts.

Figure 11.2 Frame Table

11.4.Frame building methodology 11.4.1.

Bending To construct the space frame chassis it was necessary to bend the tubes very

precisely in order to achieve the proper geometries. This was accomplished by using the machine bender available external to college facility in the shop shown in figure 11.3. Because of the minimal wall thickness of some frame members, localized tube buckling occurred when bending them to angles typically above 15 degrees.

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Figure 11.3 Bending In order to achieve high angle bends in these tubes, it was necessary to use a bending alloy which provided additional support and helped to prevent buckling. A tactic that was used to ensure that the tubes were bent to the proper angle was printing out full scale drawings of the bent tube geometry. In some frame members containing compound bends it was necessary to ensure that the tube remained level in the bender throughout the entire bending process.

11.4.2.

Notching Another major portion of chassis fabrication was the notching and fitment of the

tubular frame members. Tube notching was performed by hand using a motorized angle grinder, pedestal Grinder and smaller handheld grinder as shown in figure 11.4.

Figure 11.4 Notching

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To make notching the tubes easier and increase the notch geometry accuracy, templates were printed and wrapped around the members. To make these templates it was necessary to use the new sheet metal features in the 2013 version of Solid Works. This template is then printed out and wrapped around the tube at the proper spacing in order to grind down the desired shape. After notching, the tubes were then placed in their respective locations in the chassis and ground further to ensure a proper fitment and make the welding process easier.

11.4.3.

Welding All of the chassis components were bonded through TIG (Tungsten Inert Gas

Welding) performed by team members shown in figure 11.5. Due to the fact that there are a large number of intricate parts coupled with the thin wall material that the chassis was made from, TIG was the preferred method because of the intrinsically high amount of control. The TIG welder uses a water cooled torch with 2% throated tungsten of varying diameters as well as cup sizes varying from 4 to 8. The diameter and cup size for the torch was chosen based up the type of joint that was to be welded as well as the thickness of the material involved. After the tubes were properly notched, the ends were hand sanded to remove the mill scale.Then the ends were cleaned and were ready to be welded. These manufacturing and cleaning processes ensured complete weld penetration and thus allowed for high strength joints. As stated in the Chassis Design section, it was necessary to use AISI 4130 filler rod in all of the chassis welds.

Figure 11.5 Frame Welding

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11.5.Cost report Labour cost Cutting cost Outer diameter

= 25.4mm

Inner diameter

= 22.098mm

Cost of cutting

= Rs. 5/10mm cut = Rs. 0.5/mm

Length of 1 cut

= πd =π*25.4 =79.79 mm

Cost per cut

= Rs. 39.98

No. of cuts

=35

Cost of cutting

= Rs.1399.3

Bending cost Cost per bend

= Rs. 20

No. of bends

= 10

Cost of bending

= Rs. 200

Welding cost Outer diameter

= 25.4mm

Inner diameter

= 22.098mm

Cost of welding

= Rs. 8/10mm cut = Rs. 0.8/mm

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Length of 1 weld

= πd=π*25.4 =79.79 mm

Cost per weld

= Rs. 63.83

No. of welds

= 35

Cost of welding

= Rs.2234.05

Fish mouthing Cost of fish mouthing

=Rs.106

Total labour cost

= Rs. 106 +200+2234.05+1399.3 =Rs. 3939.35

Material cost Length of Roll Cage

= 40 m

Cost of Chromoly 4130

= Rs.700/m

Total Cost of Chromoly material

= Rs.28000/-

Overall Roll cage cost

= Rs. 28000 + 3939.35 =Rs.31939.35.

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CHAPTER-12 12.CONCLUSION The team gained knowledge in the field of designing through the virtual design and analysis with optimum usage. The team’s goal was to produce a design that met or exceeded the SAE criteria for safety, durability and maintainability as well as provide features that would have mass market appeal to the general off-road enthusiast such as performance, comfort and aesthetics. Design decisions were made with each of these parameters in mind. Computational design and analysis software solid works,pro-e,Hypermesh and ANSYS were used to verify whether each part of the design met or exceeded its stated objective. Use of these design tools also allowed the team to address and rectify conflicts between any interfacing before fabrication, saving both time and cost. Design goals were met, resulting in a final product that will withstand the rigors of off-road travel while providing the driver with the necessary comforts. The vehicle is appealing to the customer in design, driver comfort and safety, and maintainability. The vehicle is appealing to the producer in manufacturability and reliability. The use of a high strength TIG welding allows the frame to be both light weight and resilient. Using bends in the frame geometry provides strength and allows for a faster fabrication process. The approach that we followed is iterative in nature and processes like reverse engineering are adopted in order to select various systems from the ones, existing in the market. This step would ensure standardization and reliability would follow as a by part. Our top priority would always be the safety of the driver and working in this direction, we will strive to add aesthetic value and a sense of ergonomics to the vehicle.

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APPENDIX 1

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APPENDIX 2

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REFERENCES

1. SAE 2013, “Mini Baja 2013 Rules” (http://bajasaeindia.org/Rule%20Book%20BAJA%20SAEINDIA-2013.pdf). 2. Stanakova and Jovan vu. (2010) ‘Impact of anthropometric measurements on ergonomic driver posture and safety’, periodicumbiologorum Vol. 112, no 1, 51–54, 2010. 3. DarlianaMohamad, Baba MdDeros, Dzuraidah Abdul Wahab. (2010) ‘Integration of Comfort into a Driver’s Car Seat Design Using Image Analysis’, American Journal of Applied Sciences 7 (7): 937-942, 2010. 4. ISO 7250 Basic Human Body measurements for Technological Design. 5. ISO 3411 Human Physical Dimensions of operators and minimum operator space envelope. 6. Kazimi S.M.A, Solid Mechanics, revised First editionTata McGraw-Hill Publishing Co, New Delhi,1981 7. Lechner, G., Naunheimer, H., Ryborz, J., and Day, S. Automotive Transmissions: Fundamentals, Selection, Design and Application. Springer-Verlag, Berlin. 1999. 8. Milliken, William F. and Milliken, Douglas L., Race Car Vehicle Dynamics. SAE International. 1995. 9. Ramamrutham S and Narayan R Strength of Materials, DhanpatRai and Sons, New Delhi,1997 10. Keith J. Wakeham, Introduction To Chassis Design, January 2, 2009 11. www.roymech.co.uk/Useful_Tables/.../Safety_Factors.html 12. Chandrupatla T.R., and Belegundu A.D., “Introduction to Finite Elements in Engineering”,Pearson Education 2002, 3rd Edition.

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