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Submitted By: TEAM JAABAZ VELLORE INSTITUTE OF TECHNOLOGY Vellore -632014
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Complete Design Report Team Jaabaz

Oct 26, 2014

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Page 1: Complete Design Report Team Jaabaz

Submitted By:

TEAM JAABAZ

VELLORE INSTITUTE OF TECHNOLOGY

Vellore -632014

Page 2: Complete Design Report Team Jaabaz

Preface

This report describes the processes undertaken by team ‘Jabaaz’ in designing, constructing and testing a vehicle that will compete in the state of Oregon for SAE Mini Baja Competition in May-2009. The report encompasses all the aspects of the virtual development of the car. The team is also participating in VIRTUAL MINI BAJA competition organized by SAEINDIA. The purpose of this competition is to simulate a “real world” engineering design project in which collegiate teams design and manufacture a prototype of a “rugged, single seat off-road recreational vehicle intended for sale to the non-professional week-end off-road enthusiast. The design should be durable, safe, easy to maintain and must be able to negotiate rough terrain in all weather conditions.

The team would like to acknowledge the significant contributions by the team members. We would also like to thank our guide Mr. P. Sakhivel for his constant support and our beloved Dean, Dr. S. Narayanan for his incessant encouragement.

Page 3: Complete Design Report Team Jaabaz

Contents

1) Roll Cage

i) Roll Cage Design Objectives

ii) Fixing Minimum dimensions of the roll cage

iii) Base Model Selection

iv) Selection of final model and building a pipe model prototype

v) Pipe model preparation

2) Material Selection

i) Material requirement

3) Packaging constraints and subsequent changes in the model

4) Finite element analysis of the model

i) Force estimation in loading conditions

ii) Finite element modelling

iii) Finite element analysis

iv) Conclusion of Finite element analysis

v) Transient Load Analysis

5) Determination of center of mass

6) Stability Analysis

i) Effect of banking of road

ii) Maximum gradient negotiation by vehicle

7) Suspension Design

i) Suspension Geometry

ii) Design methodology: suspension

iii) FEA analysis of the arms

8) Braking System

i) Brake calculations

ii) Brake Circuit design

Page 4: Complete Design Report Team Jaabaz

iii) Innovation

iv) Finite element analysis of the disk brakes

9) Steering Kinematics

i) Correct steering angle

ii) Steering Mechanism

iii) Oversteer tendency

10) Effects of different parameters on vehicle dynamics and stability

11) Gear ratio calculation

12) Hub Assembly Analysis

13) Manufacturing Strategy

14) Conclusion

Page 5: Complete Design Report Team Jaabaz

• General Vehicle Configuration

1. Roll Cage

a) Roll Cage Design(Objectives)

The function of the roll cage is to protect the driver and support all operator control systems, front and rear suspension systems, and engine and drive train. The objective of the frame design was to satisfy these functions while meeting the SAE regulations with special considerations given to safety of the occupants, ease of manufacturing, cost, quality, weight, and overall attractiveness. Other design factors included durability and maintainability of the frame.

b) Fixing Minimum Dimensions of the 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 (6 feet 3 in or less) 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.

Fig.1. Extended Arm Position of the Driver

Page 6: Complete Design Report Team Jaabaz

Fig.2. Retracted Arm Position of the Driver

There were two positions of the driver considered:

1) The Upright Position( With straight back & extended arms)

2) The Inclined Position( With inclined back & retracted arms)

Among these postures, the second posture was selected as the probable seating posture of the driver after a feedback from the driver. Also, since the most common, comfortable, safest seat in the vehicle is the bucket type seat, an inclined position was preferred to the upright. Once the driving position was decided, a minimum clearance was left around the driver and probable dimensions of the roll cage were decided.

c) 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 modeling 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.

Also, a torso of the driver was modeled in accordance with the anthropometric charts. According to this chart, the lengths of different portions of the body can be approximated from the table shown below. There are five portions of the body:

1) T chest 2) V waist 3) F hip 4) C inseam to hancle 5) B neck to wrist

Page 7: Complete Design Report Team Jaabaz

H- Height

Fig.3. Anthropometric Chart

There were four models suggested by the design team members, each had a different approach towards solving a design issue. The following are the four models presented:

Fig.4. Model 1

Page 8: Complete Design Report Team Jaabaz

Fig.5. Model 2

Fig.6. Model 3

Page 9: Complete Design Report Team Jaabaz

Fig.7. Model 4

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 next section) were also taken into consideration. The parameters selected for comparison from FEA analysis were maximum values of the stress generated in different tests and location of the maximum stress points.

PARAMETERS MODEL 1 MODEL 2 MODEL 3 MODEL 4

MASS( USING 1020 CARBON

STEEL)

167.48 pounds 194.58 pounds

189.98 pounds

166.77 pounds

CENTRE OF GRAVITY HEIGHT

X = -0.09

Y = 15.76

Z = 11.05

X = 0.18

Y = 17.77

Z = 29.80

X =0.06

Y = 16.35

Z = 24.52

X = -0.10

Y = 17.10

Z = 36.50

VOLUME 606.45 cubic

inches 690.51 cubic

inches 674.18 cubic

inches 604.74 cubic

inches

From the table, it was found, that all the models were almost comparable from the design point of view.

Page 10: Complete Design Report Team Jaabaz

d) Selection of final model & Building a Pipe Model Prototype 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. So, a final design was made in SOLIDWORKS taking into consideration, the design features of each of the four models. The final design is presented below:

Derived from Model 2 Derived from Model 1 Redesigned Engine Tray

Fig.8. Final Model Selected

This model was again analyzed in ANSYS and the results were satisfactory with factor of safety coming above 2 in all the tests that were done. 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.

Page 11: Complete Design Report Team Jaabaz

e) Pipe Model Preparation

Fig.9. Pipe Model

Fig.10. Team engaged in fabricating Fig.11.Pipe Models with tires Pipe Model

2) Material Selection The materials used for the cage must meet certain requirements of geometry as set by SAE, and other limitations as decided by team. 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. Also, paramount are the cost considerations.

Page 12: Complete Design Report Team Jaabaz

a) Material Requirements The rules define the roll cage to be made with materials equivalent to the following specification: • Steel members with at least equal bending stiffness and bending strength to 1018 steel

having a circular cross section having a 25.4 mm (1 inch) OD and a wall thickness of 2.10 mm (0.083 inch).

Calculating the strength and stiffness about the axis that gives the lowest value ensures that the tubes with a non-circular cross-section will be equivalent even in a worst case loading situation. 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 Sy I/c, (for 1020 steel the values are; Sy=207 MPa, E=205 GPa). E = Modulus of elasticity I = Second moment of area for the cross section about the axis giving the lowest value Sy = Yield strength of material in units of force per unit area c = Distance from the neutral axis to the extreme fiber (diameter here)

Strain

Fig.12. Stress- Strain curve of AISI 1020 steel While the rules set many factors of the material’s geometry there are other limitations. These limitations include the method of fabrication and industry standards for the material. The frame will be built using a bent tube construction and TIG welded joints. TIG welding becomes difficult at wall thicknesses less than 0.035 inches. The tubing bender that will be used for the fabrication can bend a maximum of 1.5 inch diameter tube with a 0.120 inch wall thickness. It also requires that the tube have a minimum wall thickness of 0.055 inches. The geometry is also limited by industry standards. It is important to utilize commonly 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. The most commonly available materials for this type of tubing are 1020 Mild Steel and 4130 Chromoly Steel. The benefit of using 4130 Chromoly steel is that it is 17.5% stronger than the 1020 Mild Steel. The 4130 Chromoly has the same Modulus of Elasticity (E) and density as the mild steel, so using it does not affect the weight or

Page 13: Complete Design Report Team Jaabaz

stiffness in members with the same geometry. However the increase in Yield Strength affects the bending strength. As the bending strength is affected not only by cross sectional moment of inertia of the material but also by the radius, the 4130 allows the usage of a larger diameter tube with a smaller wall thickness. This in turn can allow a reduction of weight. Additionally, the 4130 Chromoly steel is more ductile than the 1020. This means that the 4130 will deform more before its ultimate failure. A chart showing the associated modulus of elasticity, yield strength, and elongation at break values for the mild and Chromoly steel is shown in the following table.

Material Modulus of Elasticity(MPa)

Yield Strength

Elongation At Break

AISI 1020 Steel 205 GPa 207 MPa 16.5%

AISI 4130 Steel

205 GPa 243.2 MPa 25.50%

While the 4130 Chromoly is not corrosion resistant, neither is the 1020 Mild steel. As both materials will have to be painted or otherwise coated for use, this will not be a factor in the material comparison.

Fig.13.Comparison of Bending Strength and Wall thickness of the tube for AISI 1020

and 4130 Chromoly

After reviewing, it was evident that the best choice would be to use 4130 Chromoly tubing with a 1.125 inch diameter and a 0.058 inch wall thickness. But, these tubes are very costly to manufacture and MIDHANI (Mishra Dhatu Ispat Nigam, Hyderabad), a Govt. of India company was contacted, the prices quoted per meter of the tube was too high to be affordable. So, in order to stick with the budgetary constraints, it was decided that AISI 1020 steel would be used for fabrication of the roll cage.

Page 14: Complete Design Report Team Jaabaz

3) Packaging Constraints & Subsequent Changes In Model

The base model which was selected was modified to suit the requirement of packaging and at the same time ensure that the model was safe and well within the safety limit. The main packaging concerns were:

1) The A-arm mounting points had to be taken into consideration. These mounting points were plotted in the SOLIDWORKS and the design was slightly modified to follow these mounting points. The mounting points were decided from the results of suspension-analyzer software.

2) In the rear end of the vehicle, a provision was provided for the mounting brackets of the gearbox to hold it in position.

3) During the placement of the rear A-arms, the mounting points if joined by straight lines were obstructing the driveshaft. Hence, the inverted U shaped tube was thought of to avoid the obstruction.

4) The driver should be able to get out of the roll cage easily. In this regard, the RRH and the RHO were connected by a member which can be used by the driver as a handle which can help the driver to get out of the car easily.

These packaging constraints are shown in the model below:

.

Fig.14(i). Salient design features of the Roll-Cage

Provision for the driver to hold and get out

Page 15: Complete Design Report Team Jaabaz

Fig.14(ii). Exploded View of the engine-transmission assembly

4) Finite Element Analysis of the Models The following tests were used to check the design 1) Front Impact Test 2) Rear Impact Test 3) Front Wheel Bump 4) Rear Wheel Bump 5) Heave 6) Rollover

a) Force estimation for loading conditions

• Estimation of Impact Force

For a perfectly inelastic collision, Energy Transferred,

E = 1/2 (m1m

2/m

1+m

2) (u

2-u

1)2

Where, m1

and m2

are the two colliding masses with velocities u2

and u1

respectively. Since both m1 and m2 are two vehicles with similar masses and the vehicle m2 is at rest,

=> m1=m

2 & u

2=0

=> DE = 1/4 m1u

1

2

F = DE/t Where‘t’ is impact time

F =1/4 * m1u

1

2 * 1/t

Weight of vehicle = 250kg

Page 16: Complete Design Report Team Jaabaz

Weight of driver = 75Kg m

1 =250 + 75 =325 Kg

Maximum Speed of Vehicle, u1 = 10m/sec

In most crashes t is of the order of 1000 ms.

F = 275 /4 *102 /1

F = 8125 N • Hence for design purposes force is taken to be 8500N.

Also, design output is for no plastic deformations. The vehicle should remain

in the elastic region. • The Safety of the driver in case of crash is taken care of by safety equipment

which includes special helmets, foam padding on bars and seat belts. • The Design Factor of Safety, FS

d was taken as 2. This relatively high value

is taken to account for the uncertainty in the nature of forces.

• Estimation of Wheel Bump Forces 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. The worst case 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.

2F = m1 * g F =1/2 m1 * g

F= 1/2* 325 * 10 F = 1625 N

Hence, Designing for F = 1.1 * 1625 = 1700N (approx).Where 1.1 is the Stress Factor of Safety. • Estimation of Loading Forces While Heaving The Entire Weight of the vehicle will be transmitted to the two points in each case. Hence F =3250N • Forces In Case of Rollover Another situation can arise when the Chassis undergoes rollover. Hence the same force as that in is applied. i.e., F =8500N. The safety in case of rollover is also covered using the concept of clearance zone given by NHTSA (National Highway & Traffic Safety Administration). This force of 8500N is applied at the top front corner points of the chassis.

b) Finite Element Modeling

Page 17: Complete Design Report Team Jaabaz

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. The geometric model in Solid works 2006 was converted into IGES format which was then imported in ANSYS. The model was imported Solid works 2006 by converting into IGES file consisting of datum Points and Lines.

Fig.15. Analysis Procedure in ANSYS

Element: 3D Elastic Pipe Elements (Pipe16) Line Elements

Material Properties Linear, Elastic, Isotropic

Exy=2.08 x 1011

N/mm2 ν= 0.28(Poisson’s Ratio) Element Properties

Area = 1.28 x 10-4 m2 Outside Diameter=28.448mm (1.12 in) Pipe Thickness=3.048mm (0.12in)

c) Finite Element Analysis

The aim is to carry out a design check of the given Mini Baja chassis under estimated loading conditions and to minimize the weight of the frame keeping a Safety Factor of 2.Here, as a model calculation, analysis of model 4 is shown. Material of the tubes chosen was AISI 1020, Hot Rolled with properties: Sut(Ultimate Strength) = 379 Mpa Syt(Yield Strength) = 207 Mpa

Density = 7800 Kg/m3

Front Impact Test

• Model Used: Full Model • Loading: F= 8500N on Front Corner points. • Boundary Conditions: Rear Corner Points ,All DOF=0

Results: Stress: Max Stress= 111.54 MPa

Factor of Safety: Incorporated Factor of Safety = Syt/Smax = 207/111.54 = 1.85(Roughly equal to 2)

Page 18: Complete Design Report Team Jaabaz

Hence, the Chassis will be safe under Frontal Impact.

Fig.16. Front Impact Test (Von Misses Stress Plot)

Rear Impact Test

• Model Used: Full Model • Loading: F= 8500N on Rear Corner. • Boundary Conditions: Front Corner Points, All DOF=0 Top Front Corner, All DOF=0

Results: Stress: Max Stress= 205.9 MPa

Factor of Safety: Incorporated Factor of Safety = Syt/Smax = 207/205.9 = 1.005 < 2

Hence, the Chassis will not be safe under rear Impact

Page 19: Complete Design Report Team Jaabaz

Fig.17. Rear Impact Test (Von Misses Stress Plot)

Front Wheel Bump Test • Model Used: Full Model

• Loading: F= Weight=1700N on front wheel. • Boundary Conditions: All DOF =0 at Rear Wheels and Opposite Front Wheel. Results: Stress Max Stress= 134.497 MPa

Factor of Safety

Incorporated Factor of Safety = Syt/Smax = 207/134.497 = 1.53 < 2 Hence, the Chassis will not be safe during a Front Wheel Bump.

Page 20: Complete Design Report Team Jaabaz

Fig.18. Front Bump Test (Von Misses Stress Plot)

Rear Wheel Bump Test • Model Used: Full Model • Loading: F= Weight=3250N (Vehicle + Driver Weight) on Rear Right Wheel. • Boundary Conditions:

1. All DOF =0 at Front Wheels and Rear Left Wheel Results:

Stress Max Stress= 399.21 MPa

• Factor of Safety Incorporated Factor of Safety = Syt/Smax = 207/399.61 = 0.52 <2 Hence, the Chassis will not be safe during a Rear Wheel Bump

Page 21: Complete Design Report Team Jaabaz

Fig.19. Rear Bump Test (Von Misses Stress Plot)

Heave Loading • Model Used: Full Model

• Loading: F= 2 X Weight=6500N is to be supported. Force on Two Front Corner Points= F/2 =3250N Force on Two Rear Corner Points= F/2 =3250N

• Boundary Conditions: All DOF =0 on all Key points on the Upper Surface of the Frame.

Results: • Stress

Max Stress= 174.11 MPa

Factor of Safety Incorporated Factor of Safety = Syt/Smax = 207/174.11 = 1.18 < 2

The model needs to be improved by adding more members upfront to decrease chances of failure under heave loading.

Page 22: Complete Design Report Team Jaabaz

Fig.20. Four Wheel Heave Test (Von Misses Stress Plot)

Twisting Load • Model Used: Full Model • Loading: F= 1700N on Front wheel F=3250N on Rear wheel •Boundary Conditions:

All DOF =0 on Key points( Front Left & Rear Right)

Results: • Stress:

Max Stress= 270.151 MPa Stress Distribution

Factor of Safety, Incorporated Factor of Safety = Syt/Smax = 207/270.151 = 0.99 < 2 Clearly, the model will fail. This can be attributed to stress concentration at rear engine tray.

Page 23: Complete Design Report Team Jaabaz

Fig.21. Four Wheel Heave Test (Von Misses Stress Plot) Rollover Loading • Model Used: Full Model • Loading: F= 8500N on Top Front Points (Rollover is toppling along the axis of the

vehicle here) • Boundary Conditions:

All DOF =0 on all Key points on the Bottom Members of the Frame.

Results: • Stress:

Max Stress= 239.354 MPa Stress Distribution

Factor of Safety, Incorporated Factor of Safety = Syt/Smax = 207/239.354 = 0.86 < 2

Page 24: Complete Design Report Team Jaabaz

Fig.22. Roll over Test (Von Misses Stress Plot)

d) 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. This addition of bracing at the key stress concentration points does help in meeting the safety requirements. Also, it was decided to include gussets in the weld so that the effective welding area increases. The factor of safety taken is 2 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.

Page 25: Complete Design Report Team Jaabaz

d) Transient Load Analysis In order to actually simulate the on track loading conditions time varying loads were applied. These time varying loads were considered to be ramped. Ramped load step means that the load steadily increases from zero to maximum loads in constant steps. The plots for the various times are shown. Stepped Versus Ramped Loads When we specify more than one substep in a load step, the question of whether the loads should be stepped or ramped arises.

• If a load is stepped, then its full value is applied at the first substep and stays constant for the rest of the load step.

• If a load is ramped, then its value increases gradually at each substep, with the full value occurring at the end of the load step.

Fig.24.Stepped Versus Ramped Loads

• Model Used: Full Model • Loading: F= 1700N to -1700N on Front left and Right Points • Boundary Conditions:

All DOF =0 on all Key points on the rear side of the Frame.

Fig.25. Load of 1700N is applied in 20 time steps from 1700N to -1700N

Page 26: Complete Design Report Team Jaabaz

Fig.26.Nodal Displacements at 5th, 10th, 15th and 20th Load Steps

Page 27: Complete Design Report Team Jaabaz

5) 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.

From the mass properties option, the C.G height was roughly found out to be 16.5 inches.

Page 28: Complete Design Report Team Jaabaz

6) Stability Analysis

a) Effect of banking of road We have a car cornering at a speed ‘v’ and radius ‘r’, measured in a horizontal plane passing through C.G on a banked road with bank angle α .

Fα= centrifugal force = mv2/r = waα Fy = - Fα cosα + w sinα. Fz = Fα sinα + w cosα

� Fy = w(-aα cosα + sinα) w’ = Fz = w(aα sinα + cosα) Effect of banking is to change the tire load. Lets say that the car is approaching a banked turn at v = 45 kmph = 12.5 m/s w = 350 X 9.81 = 3433.5N r = 5 m Fα = Fc = mv2/r = 10937.5 N Net force in the horizontal direction : Fy = Fc cosα – w sinα = 10937.5 cosα – 3433.5 sinα This force is balanced by frictional force on ground. It is given by f = µN So Fy = µN Where µ = 0.1 Now N = Fz = Fcsinα + w cosα � N = 10937.5 sinα + 3433.5 cosα � 10937.5 cosα – 3433.5 sinα = 0.1(10937.5 sinα + 3433.5 cosα) � α = 66.860

This is without considering the weight transfer during cornering.

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If we consider weight transfer :

Lateral weight transfer = lateral acceleration(g) X weight X C.G height / track width Acceleration = v2/rg = 12.52/(5 X 9.81) = 0.2548 g C.G height = 18.5 inches Weight on the front half of the vehicle = 35 % of total weight Thus later load transfer = 0.2548 X (0.35 X 350) X (18.5 X 25.4/1000)/(58 X 25.4/1000) = 7.80 kg = 76.518 N Weight on the rear half of the vehicle = 65 % of total weight Load transfer = 14.959 kg = 146.74 N So effective weight on the wheels is : Front outside = (.35 X 350 X 9.81)/2 + 76.518 = 677.308 N = 69.05 kg Front inside = (.35 X 350 X 9.81)/2 - 76.518 = 524.244 N = 53.45 kg Rear outside = (.65 X 350 X 9.81)/2 + 146.74 = 1262.62 N = 128.7 kg Rear inside = (.65 X 350 X 9.81)/2 + 146.74 = 969.147 N = 98.79 kg Now we balance the friction force on the outside tire with the horizontal component of Fc and w. Friction force f = 0.1(summation of( mv2/r sinαααα + wcosαααα)) = 0.1[(69.05 X 12.52 X sin α/5.7366) + (677.308 cosα) + (53.45 X 12.52 X sin α/4.2634) + (524.344 cosα) + (128.7 X 12.52 X sin α/6.5748) + (1262.6275 cosα) + (98.79 X 12.52 X sin α/3.4252) + (969.1475 cosα)] => f = 1140.464 sinα + 343.34 cosα Now this is equal to Fy :

� 10937.5 cosα – 3433.5 sinα = 1140.464 sinα + 343.34 cosα � α = 66.64.

Thus we see that there is not much difference in the angle if we neglect weight transfer. This shows that because of large track, there is very little weight transfer to the outer wheels and hence the performance is satisfactory.

Page 30: Complete Design Report Team Jaabaz

b)Maximum gradient the vehicle can negotiate :

Here we are assuming that car is resting on the slope. µNa + µNb - mgsinθ = 0 µ = 0.25

� 0.25(Na + Nb) - mgsinθ = 0 and along y direction :

(Na + Nb) - mgcosθ = 0 From the above two equations we get : tanθ = .25 => θθθθ = 14.036o This behaviour is independent of the C.G height. To incorporate C.G we need to know the longitudinal weight transfer. Longitudinal weight transfer = acceleration(g) X weight(kg) X C.G height / wheelbase Acceleration is decided by our requirement. If suppose the slope is of length 25m and the time in which it is to be covered is taken to be 12 seconds with the vehicle approaching the slope at a speed of 10 km/hr we get acceleration as: u = 10km/hr = 2.77 m/s t = 12 sec s = 25 m Using Newton’s motion equation : s = ut + ½ at2 we get a = 0.2327 m/s2 = 0.02372 g Thus, Longitudinal weight transfer = 0.02372 X 350 X 9.81 X (15 X 25.4/1000) / (68 X 25.4/1000)

� 179.65 N Weight in front tire = (350 X 0.35 X 9.81/2) – 179.65/2 = 511.037 N Weight on rear tire = (350 X 0.65 X 9.81/2) + 179.65/2 = 1205.712 N

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Therefore, 2µ(Na + Nb) - mgsinθ = ma Let the road be slippery, so the coefficient of dynamic friction µ = 0.1. So now the equations will be: 2 X 0.1(5.11.0375 +1205.7125) – 350 X 9.81 X sinθ = 350 X 0.2327

� θθθθ = 4.37o If we take µ = 0.3 then: θθθθ = 16.030

Page 32: Complete Design Report Team Jaabaz

7) Suspensions Design and Wheel Kinematics

Kinematics describes the movement caused in the wheels during vertical suspension travel and steering, whereas elasto-kinetics defines the alterations in the position of the wheels caused by the forces and moments between the tyres and the road. While designing the suspension certain standards are followed like German standards like DIN 70000 and Din 74250 as well as The majority of the front suspensions used today are

1. Mac Pherson strut 2. SLA( short- Long-arm)

Fig.27. Front suspension packaging

Following are the General Front suspension design issues that must be accounted for:

• Packaging parameters that are fixed. These are determined by the space which other systems take.

• Package the wheel, tires, brakes and bearings • Tire size and rim diameter and width must be decided(- decide wheel offset,

fitting of brake calliper, location of brake rotor and lower ball joint, wash rack clearance)

• Decide Kingpin dimensions and kingpin angle in front view (decided by scrub radius, spindle length. A compromise has to be made. Eg. - If a certain scrub radius is wanted then you have to establish 2 fixed point’s accordingly-1. Lower ball joint and 2. Ground contact point of the kingpin. In a rear-wheel drive the lower ball joint can be pushed as far as possible and a kingpin angle lesser than 8 degrees is acceptable. More the kingpin angle more is the car lifted when steered. Also a longer spindle length means more lift. Camber of wheels is also a function of kingpin angle and caster angle. With no kingpin angle there is no camber angle with steer lock. With positive caster and no kingpin angle, the wheel gains negative camber on the outside wheel and positive camber on the inside wheel. Caster adds favourable angle to the effects of kingpin angle. Thus low kingpin angles are desirable as it subtracts from negative camber gain due to caster on the outside wheel.

• Rack position has to be decided. This will depend on engine location and orientation, what kind of drive it is. Rack mounting stiffness versus upper or

Page 33: Complete Design Report Team Jaabaz

lower control arm mounting stiffness. Lateral displacement of ball joints in relation to tie-rod outer pivot during cornering will cause steer angle. It is better for stability to have lateral force deflection toe-out than toe-in. A high mounted rack should be behind wheel center and a low mounted rack should be ahead of wheel center.

• Structural requirements for suspension design must be considered when packaging each element of the total system. Eg.- Control arms that have one leg straight across from the ball joint are superior in system stiffness to arms. Linkage ratios for spring, shock and stabilizer bar as close to 1:1 will produce more direct load paths thus improving system stiffness while providing a lighter overall design.

Compared to the Mac-pherson Strut type suspension, the SLA is better as it requires less space than the former. Hence, it is better suitable in this scenario.

a) Suspension Geometry The stability and effective handling of a vehicle depends on the optimum steering and suspension geometry which particularly includes the parameters like the wheel camber, castor and the kingpin inclination. For the ease of designing, the design parameters are subdivided into front view geometry parameters and side view geometry parameters.

• Front View Geometry The front view swing arm instant centre is uniquely determined by the desired roll centre height and roll camber. The desired roll camber sets the front view swing arm length as follows:

The above procedure is implemented and suitable suspension geometry in front view is drawn. The following are the vehicle parameters that are decided before the construction:

1) Vehicle Track Width - The vehicle track width should be kept as large as possible as it avoids possible weight transfer during cornering. The lesser

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the weight transfer, the better is the stability of the vehicle. But large track width has a detrimental effect on the manoeuvrability. Hence, an optimum value is selected, the value is assumed to be 58 inches considering the packaging constraints and in accordance with the rulebook.

2) Ground Clearance- The ground clearance of the vehicle is assumed to be 14 inches. This is done so as to avoid large boulders and obstructions from affecting the safety of the driver. Also, adequate ground clearance allows a Baja vehicle to wade through deep mud pits, which is usually an integral part of the competition.

3) Tire Size- The tire size was assumed to be 23 inches in diameter. The width of the tire was 7 inches and the rim size was 10 inches in diameter. The tire was actually selected after it was successfully used in the first Mini-Baja India event.

Fig.28. Front View Geometry Constructed in Solidworks

The front view instantabeous center height is set by projecting a line from the tire center ground contact patch through the desired roll center. The instant center must lie on this line. Now, lines are projected from both the ball joints to the instant center. These become the centerlines of the upper and lower control arm planes as projected into the vertical plane through the wheel center. Packaging requirement would decide on the length of the control arm but it should be made as large as possible. The length of the upper control arm in relation to the lower adjusts the shape of the camber curve. If they are the same length, the camber vs wheel travel will be a straight line. If the upper is longer than the lower, the curve must be convex with its curvature towards positive camber. If the upper is shorter than the lower, the curve will be concave towards the negative camber. As the upper is made progressively shorter, the camber increases. The ideal curve has progressive camber in bump with less camber change in droop. The tie rod and rack location should be roughed in by projecting a line through the tie rod outer point and the front view instant centre. The correct tie rod length is then established

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for a linear ride toe curve or by the method of instantaneous centres as described in the steering section of the report. The front view swing arm instant center location controls the roll center height, the camber rate, and tire lateral scrub. The Instantaneous Center can be located inboard of the wheel or outboard of the wheel. It can be above ground level or below ground. The location is up to the designers, performance requirements.

• Roll Center Height

The roll center height is found by projecting a line from the center of the tire-ground contact patch through the front view instant center. This is repeated for each side of the car. Where these two lines intersect is the roll center of the sprung mass of the car, relative to the ground. It is not necessarily at the centerline of the car, especially with the asymmetric suspension geometry or once the car assumes a roll angle in turn. The roll center is controlled by the instant center heights above or below ground, the distance away from the tire that the instant center is placed inboard or outboard of the tire contact patch. The roll center establishes the force coupling point between the unsprung and sprung masses. When a car corners, the centrifugal force at the center of gravity is reacted by the tyres. The lateral forces at the CG can be translated to the roll center if the appropriate force and moment are shown. The higher the rolls center the smaller the rolling moment about the roll center. The lower the roll center the larger the rolling moment.with higher roll centers the lateral force acting at the roll center is higher off the ground. This lateral force times the distance to the ground can be called as nonrolling overturning moment. So roll center heights are trading off the relative effects of the rolling and nonrolling moments. Another factor in establishing the desired roll center is the horizontal-vertical coupling effect. If the roll center is above the ground level the lateral force from the tire generates a moment about the instantaneous center, this moment pushes the wheels down and lifts the sprung mass and is called jacking. If the roll center is below the ground level then the force will push the sprung mass down. In either way the sprung mass will have a vertical deflection due to the lateral force.

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Fig.29. Roll Center( 3 Instantaneous Center method)

The roll center height may be derived for short swing arm suspension by consideration of similar triangle concept:

_h = r_ t/2 l

Where, h= roll center height t = track width

r = wheel radius l = swing arm length

The swing arm length can be found out from the formula: Fvsa (front view swing arm) length = (t/2) / (1- roll camber) Roll camber is the slope of the graph between the camber changes versus the degree of roll. This graph is obtained as below using the package SUSPENSIONANALYZER. Using, the software, as described below, the camber change curve for roll of 2 degrees is found out.

Fig.30. Graph between camber change and roll degrees

From the slope of the graph, the roll camber can be found out. Roll camber= 0.3/0.5= 0.6 Therefore, Fvsa = (58/2)/ (1-0.6) = 72.5 inches

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Hence, the height of the roll center is given by h = r x t/2=11.5x29/72.5 = 4.6 inches l

• Camber Change Rate While the roll center is a function of the front view swing axle length and height, the camber change rate is only function of the front view swing axle (fvsa) length. If the control arms of the suspension were to be replaced by a single link that ran from the knuckle to the instant center, then the amount of camber change that was achieved per inch of ride travel would be given as Camber change ( deg./in.) = arc tan(1/fvsa length). Thus the longer the arms, lesser is the camber change.

Hence, the camber rate change is given by the Arctan (1/fvsa)= 0.79 degrees per inch of bump or droop

• Rate Of Change Of Front View Swing Arm Length The instant centers move with wheel ride travel. The rate is a function of the absolute and relative lengths of the control arms in the front and side views. A camber curve can be made to have more or less camber change with wheel travel by altering the length of the upper arm even though it is aimed at the same instant center at ride height. By this the arm length at ride height is same but is shortens faster or lengthens slower with wheel travel.

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• Scrub

Tire Scrub is another variable or front view. This is the lateral motion relative o the ground that results from vertical motion of the wheels. Scrub occurs in every suspension system. The amount of scrub is a function of absolute and relative lengths of the control arms and the position of the front view instant center relative to ground. When the instant center is at any position other than ground level then scrub is increased. If it is above ground and inboard then tire will move outward as it rises. If it is below ground level and inboard then the opposite happens with the rise in tire. The amount of movement is function of the control arm length and absolute height from the ground. On a rough road the wheel path is not a straight line if there is scrub. Significant amounts of scrub introduce lateral velocity component at the tire which, when added to the forward velocity, change the tire slip angles. This in turn laterally disturbs the car. The same slip angles will also add viscous damping to the ride motion.

The scrub radius is used in order to give self straightening effect. The self aligning/straightening torque is the product of the traction/braking force and the scrub radius. The scrub radius is assumed to be 25mm for the given suspension arrangement and it is considered to be positive. Kingpin Axis Inclination is about 6 degrees.

• Side view geometry The side view geometry deals with the geometry when looked from the sides. The desired instantaneous centre is established first. This is a result of calculating the desired anti-features, the side view swing arm length that is acceptable and whether the wheel path must be receding in bump or not.

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Using the above procedure, the side view geometry is drawn and is shown below. Typical assumptions before drawing the side view geometry were:

1) The caster angle was assumed to be 5 degrees. 2) The anti squat and anti dive features were not included in the design.

The justification of these two assumptions would be provided after presenting the diagram.

Fig.31. Side View Geometry

In the side view geometry, the following parameters are important.

1) Caster Angle: The inclination of swivel ball joint axis or the kingpin axis in the fore and aft direction, so that the tire contact center is either behind or in front of

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the imaginary pivot center point produced to the ground, is known as the caster angle.

The effect of caster angle can be seen, when the steering is partially turned on one lock. The trail or lead distance between the contact patch center and the pivot center rotates as the steered wheels are turned so that the forward driving force and the equal but opposite ground reaction are parallel and form a couple. This causes self straightening effect. The self straightening effect increases as the angle is increased.

Fig.32. Castor angle and self aligning torque

2) Anti Squat and Anti Dive features

The anti squat and anti dive features are incorporated in the suspension design to negate the effects of weight transfer during the acceleration and braking and the subsequent effect on comfort ability of the occupant in the vehicle. By, altering the suspension geometry, the anti squat and anti dive properties can be incorporated. In a double wishbone suspension design, the arm mounting points are tilted so that when produced, they meet at an imaginary point in side view.

Fig.33. Side view geometry for anti-squat/dive

When the vehicle accelerates forward, the reaction to the driving torque pivots the suspension arm about the axle in the opposite direction to the input torque. Thus, the arm swings downward and opposes the upward lift of the body.

By, addition of this feature, the major problem is that motion of wheels takes place in two planes of reference. The first plane of reference is the front plane where the tire moves

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left and right on the bump. The second motion is in the side view, where the wheel moves in the fore and aft direction. This type of motion adds an unnecessary complexity to the motion.

An ingenious method of reducing the squat/dive is:

a) Reducing the weight transfer in the longitudinal direction.

Longitudinal weight transfer = accleration(g) X weight(kg) X C.G height / wheelbase Hence, by increasing the wheelbase without affecting the maneoueverability is the task. A wheel base of 68 inches suits this requirement. Also, the height of the center of gravity height is kept low at 16.5 inches.

b) Using a spring with high stiffness in the shock absorbers. This avoids large lowering or lifting of the body while accelerating or braking. Hence, the side view geometry is kept as simple as possible so that the mounting points are horizontal and the side view instantaneous center is found at infinity.

b)Design Methodology: Suspension

The methodology followed for designing a proper suspension system is an iterative one. An iterative design process, assumes some parameters, performs an analysis of the affect of these on the performance.

The parameters influencing the stability and performance of a suspension system can be classified into known and dynamic.

The common dynamic parameters are:

i) Installation Ratio/ Motion Ratio - It is the ratio of the displacement of the wheel to the compression/ expansion of the spring.

ii) Spring Stiffness- The stiffness of the spring changes progressively during the bump and jounce.

iii) Camber angle change- The suspension geometry and the instantaneous center position guides this as explained earlier. The camber angle change is found out. Sometimes, this changes non-linearly.

iv) Toe in/Toe out Characteristics- Generally, toe-in is more sought after because while cornering, the load transfer will cause the slip angle to increase on the more heavily laden wheel. This in turn increases stability.

v) Roll Center Height- The distance between the roll centre and the center aligning torque is the radius arm of the destabilizing torque. The dynamic roll center height which is induced by the suspension geometry controls this distance between the roll center and the center of gravity.

These parameters are most sought after as they affect the stability of the vehicle. The design stage involves the following stages:

1) Finding out the coordinates of the mounting points of the suspension A-arms with respect to a coordinate system.

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2) Entering the known parameters which affect the vehicle’s dynamic performance like the wheel base, track, vehicle weight distribution, C.G height, tire size etc.

3) Performing the analysis with the help of the software tool like SUSPROANALYZER which helps in the simulation of the conditions of roll, dive and steer.

4) Analysis of the output in terms of graphs between the parameters like the wheel travel vs camber change or roll center height variation etc.

This is explained by the graphical output below:

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Fig.34. Rear Suspension Model/ Parameters

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Fig.35. Front Suspension Model/ Parameters

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Three different types of conditions were simulated:

1) + 2 degrees roll 2) + 2 degrees dive 3) + 2 degrees steer

The results were generated as shown below:

1) For rolling conditions

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2) During steering of the wheels

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3) During bump/ jounce conditions

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• Analysis of the result

An example of how the result of the analysis is useful is taken. During the bump/ jounce test, the roll center height variation takes place. As rightly explained by the graph, the variation for rolling between + 2 degrees roughly 6 inches overall. The height of the C.G is around 16.5 inches. Thus, the distance between the C.G and the roll centre decreases during the bump travel. Hence, a qualitative conclusion can be undertaken that the destabilizing torque value would not be sufficient due to a shorter moment arm.

Fig.35. Roll Center Height variation while bump/droop

c)FEA Analysis of the arms

The need for finite element analysis of the A-arms was felt on pre-empting the worst case scenario which the suspension can undergo when subjected to road loads. The maximum road shock would be transmitted to the roll cage through the arms when the shock absorbers fail. This causes the stiffness of the shock absorbers to increase rapidly, such that the elastic shock absorbing capacity is compromised.

• Model Used: A arm • Loading: F= 6500 N

• Boundary Conditions: All DOF =0 at frame mounting point • Element Type- Solid 45

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Fig.37. Analysis of A-Arm in COSMOS EXPRESS (Von-Misses Stresses)

Fig.38. Simulating Load Conditions(ANSYS)

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Fig.39. Meshed Model in ANSYS

Fig.40. Von Misses stress distribution in ANSYS

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8. Braking System

The type of braking system is decided taking into consideration the comparison between disk brakes and drum brakes.

Advantages of disc brakes over drum brakes: 1. The friction surfaces in a disc brake are directly exposed to the cooling air, whereas

in the drum type, the friction occurs on the internal surfaces from which heat can be dissipated only after it has passed by conduction through the drum.

2. The friction pads in case of disc brakes are flat compared to the curved ones in drum type thus there is uniform wear in the linings of the disc brake friction pads.

3. Unlike drum brakes, in disc brakes there is no loss in efficiency due to expansion. 4. Disc brakes have comparatively better anti-fade characteristics. 5. Disc brakes have a simple design compared to that of drum brakes due to which – a)

servicing the brakes is easier, b) changing friction pads is easier and, c) it weighs less

Fig.41. Floating Brake Caliper

a)Brake calculations For the brake calculation we take: Pressure from the master cylinder is considered as 3 bar. The diameter of the master cylinder is taken as 1 inch The diameter of the wheel cylinder of the brake cylinder at the disc is taken as 1.18 inch The diameter of the wheel is taken as 23 inch = 5.84 m The disc diameter is assumed to be the diameter of the rim of the wheel which is 180 mm thus we calculate the maximum force we are getting. The force of the piston rod (Fk) = pressure from master cylinder (p) X piston area of the master cylinder (Ah)

� Fk = p X Ah � Fk = 3 X 105 X 2X3.14X0.1272 = 15201.22 � Fk = 15201.22 N

Clamping force on the wheel cylinder (Fs) = Fk X (diameter of master cylinder(dh)/diameter of wheel cylinder(dr))^2

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= 15201.22 X (25.4/30)2 = 10897.06 � Fs = 10897.06 N

Brake force at brake disc(Fb) = Fs X coefficient of friction on lining(µ) X no of cylinders(z) = 10897.06 X 0.28 X 2 = 6102.35

� Fb = 6102.35 N Brake force at the periphery of the tyre(F) = Fb X (radius of brake disc(rt)/ radius tyre(rd)) = 6102.35 X ( 3.54/23) = 939.23 N

� F = 939.23 N So the total force on the vehicle due to braking is = 939.23 X 4 = 3756.93 N Thus the braking force is 10612.80 N. The mass of the vehicle is taken as 325 Kg.Using this we calculate the deceleration on the vehicle. F = m X a � 3796.93 = 325 X a � a = 3796.93/325 = 11.55 m/s2 Thus the deceleration produced is 11.55 m/s2 Assuming the vehicle will have a velocity of 30 km/hr = 8.33 m/s, using Newtons motion equation we find what the stopping distance will be : V2 = 2 X a X s � 8.332 = 2 X 11.55 X s � s = 69.38/23.10 = 3 m

Thus the stopping distance is 3 m

b)Brake Circuit

Diagonal split is used in the tandem master cylinders here because in case of failure, at least two wheels will help the vehicle to stop and reduce the yaw movement of the vehicle. In this arrangement, the front left wheel caliper is connected to the same port of the master cylinder as the rear right wheel caliper. These calipers are connected to a common port through a T shaped valve called as splitter. The splitter diverts the flow from the main line of the master cylinder to the individual lines of each wheel without sacrificing the pressure in the line. The diagonal arrangement is also advantageous in the sense that it prevents yawing of the vehicle in the case of brake failure. In the case of brake failure, if one the brake circuits fail, still one of the circuits is functional and the braking force is applied to the diagonal wheels. This diagonal arrangement of brake forces ensure that the moments produced by the moving wheel about the braked wheels are opposite in direction in both front and rear and thus cancel out each other, hence preventing the yaw movement of the vehicle.

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Hence, the yawing movement can be reduced by the introduction of this type of arrangement.

Fig.42.Diagonal brake circuit arrangement

c)Use of single disk arrangement in rear (Innovation)

With the introduction of caliper in each wheel assembly, it becomes virtually difficult to

package all the components in a single wheel assembly. In the rear wheels, there is a

further problem as suitable arrangements have to be made for drive shafts to fit inside the

hub assembly. To simplify the arrangement, it was thought that the wheels in the rear

should be connected to a single disk and caliper should apply braking force to the common disk to apply brakes.

From the view of packaging, this arrangement is very advantageous as there is no need

to find the mounting points in the hub assembly for the calipers and there is no restriction

on the size of the disk. Also, there is no need of protecting the brake assembly during the

events where the tires are submerged in mud. But, the problem which is faced is that if

we place a single caliper on a single disk, then in the event of failure of one the circuit, the rear wheels would have no braking at all.

Hence, the concept of diagonal braking is applied to the problem. Another caliper is

introduced in the rear disk. In total, there are two calipers in the rear disk. Each caliper in

the rear is connected with a caliper in the front through the splitter to a port in master

cylinder. In this case, even if one of the circuits fails, both the rear wheels will have at least reduced braking effect because of one of the calipers still functioning.

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By, the use of ingeniously designing the braking circuit, the space constraint in the braking rear side is well optimized with the braking performance.

Fig.43.Three dimensional Visualization of the brake circuit

Fig.44.Innovative brake design in the rear

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d)Finite Element Analysis of Disk brakes The role of FEA analysis in disk brakes helps us to ascertain the importance of holes in the design of the disks. Following the details of the FEA analysis: Type of Element used: PLANE 55 Since the thickness of disk is very small when compared to its diameter, hence the analysis is two dimensional (2D). For this reason, the plane element is chosen in place of a solid element as in case of a 3D analysis. Boundary Conditions- The boundary condition chosen is that the temperature of the outside edge is 40۠ C and temperature of the inside edge is also 40۠ C. Thermal Load- The temperature of the contact patch is assumed to be 700 C. It is assumed to be annular region.

Fig.45. Application of thermal load and boundary conditions

The graphical plot of the temperature is also given. Here, temperatures at individual node points are calculated by solving the conduction equation. The temperatures at different points are calculated with and without holes.

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Fig.46. Temperature distribution without ventilating holes

Fig.47. Temperature distribution with ventilating holes

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• Analysis of the result By using 8 holes of 10mm diameter in the disk, the temperature in the disk is reduced as can be seen from the contour plot where more area is now under reduced temperature. Hence, by using holed disk, the life of the disk can be improved from thermal degradation. The braking system desired for the vehicle is a disk brake on all four wheels. The front disks are inboard and the rear disks are outboard. By using two calipers in the rear instead of one on each, sufficient space has been saved in the rear hub assembly and the complexity has been reduced. The brake circuit chosen was a diagonal split one which ensures the safety of the driver even when the primary circuit fails. The handbrakes would be of mechanical type and would not be a part of the hydraulic brake circuit. To prevent the yaw movement of the vehicle on braking and to further improve the stability, it was decided to go for outboard braking in the rear.

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9.Steering Kinematics The steering system allows the driver to control the direction of vehicle travel. This is made possible by the linkages that connect the steering wheel to the steer able wheels and tires. The steering system may be either manual or power. The steering system has three major components: 1) The steering wheel and steering shaft that transmit the driver’s movement to the steering gear. 2) The steering gear that increases the mechanical advantage while changing the rotary motion of the steering wheel to linear motion. 3) The steering linkage that carries the linear motion to the steering arms. Like in designing any subsystem, some suitable targets were thought off, the means to achieve them were found out and the effects of the system’s performance on other systems were analyzed. Typical target for a Mini Baja vehicle designer is to try and achieve the least turning radius so that the given feature aids while maneuvering in narrow tracks, also important for such a vehicle is that driver’s effort is minimum. This is achieved by selecting a proper steering gear. The next factor to take into consideration deals with the response from the road. The response from the road must be optimum such that the driver gets a suitable feel of the road but at the same time, the handling due the bumps is not affected. Lastly, the effect of steering system parameters on other systems like the suspension system should not be adverse.

Fig.48. Steering Assembly Components

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Fig.49. Steering Assembly Components (Exploded View)

a)Correct Steering Angle

The perfect steering is achieved when all the four wheels are rolling perfectly under all the conditions of running. While taking turns, the condition of perfect rolling is satisfied if the axes of the front wheels when produced meet the rear axis at one point. Then this point is the instantaneous centre of the vehicle. The requirement is that the inside wheel is made to turn through a greater angle than the outer wheel. The larger the steering angle, the smaller is the turning circle. There is however a limit to the maximum steering angle. It has been found that steering angle of the inner wheel can have a maximum value of about 44 degrees. The figure below shows the position of the wheels for correct steering.

Referring to the figure, for correct steering:

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Cot Ø= y+c where Ø= angle of outside lock b Cot θ= y_ where θ=angle of inside lock b Hence, Cot Ø - Cot θ= c_ b This equation represents the basic condition for the steering mechanism to be perfect rolling of all wheels. To solve the above equation, trial and error method is used. In the above equation, c is the distance between pivot centers of the steering tie rods and b is the wheelbase. From the vehicle parameters, c = 47 inches = 1200 mm (approximately) b= 68 inches = 1727 mm (approximately) c/b= 0.70 ( approximately) From the relation, Cot Ø - Cot θ= c_=0.70 b By trial and error method, the approximate values of angles are 35 degrees and 25 degrees respectively. Hence, for perfect rolling conditions and no slipping condition on the tires, the angles of steering are Ø= Outer wheel lock angles = 25 deg and θ= Inner wheel lock angle = 35 deg From the values of these angles, the turning radius for different wheels can also be found out as follows:

i) For the inner front wheel R (Inner Front)= b - (a – c) = __68__ - (58-47) = 113 inches = 9.42 feet Sin θ 2 Sin35 2 Where a= track width= 58 inches

ii) For the outer front wheel R (Outer Front)= b - (a – c) = __68__ - (58-47) = 155 inches = 12.95 feet Sin Ø 2 Sin25 2

iii) For the inner rear wheel R (Inner Rear)= b - (a – c) = __68__ - (58-47) = 92 inches = 7.63 feet Tan θ 2 Tan 35 2

iv) For the outer rear wheel R (Outer Rear)= b - (a – c) = __68__ - (58-47) = 140 inches = 11.69 feet Tan Ø 2 Tan 25 2

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b)Steering Mechanism To achieve the correct steering, two types of mechanisms are used. They are the Davis and Ackermann mechanism. Ackerman mechanism is used generally for applications where the speed is less and where the lateral accelerations are low. This type of geometry is apt for an all terrain vehicle like the Mini Baja where the speed seldom exceeds 40 KMPH because of the terrain. This geometry ensures that all the wheels roll freely without the slip angles as the wheels are steered to track a common turn center. The simplest construction that generates Ackermann geometry is where the rack is located behind the front axle and lines starting at the kingpin axis and extended through the outer tie rod ends when extended intersect the center of the rear axle. The angularity of the steering knuckle will cause the inner wheel to steer more than the outer wheel and a good approximation of the perfect Ackermann is achieved. The above explained method is shown below with a diagram.

Fig.49. Ackermann Geometry

A second way to design-in differences between inner and outer steer angles is by moving the rack forward or backward so that it is no longer on the line directly connecting the two outer tie rod ball joints. Another way to generate toe with steering is simply to make the steering arms different lengths. A shorter steering arm, as measured from the kingpin axis to the outer tie rod end will be steered through a larger angle than one with a longer knuckle. But this effect is asymmetric and applies only to cars turning in one direction, eg. Oval tracks.

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Hence the method of extending the outer tie rod ends to intersect at the rear axle is most preferred. When the vehicle is in straight ahead position, these links make equal angles α with the center line of the vehicle. The dotted lines indicate the position of the mechanism when the vehicle is turning to the left. Let l= length of the track rod r= length of the steering arms. Then, referring to the figure and neglecting the obliquity of the track rod in the turned positions, the movements of A and B in the horizontal direction may be taken to be the same as equal to ‘x’.

Fig.50. Ackermann Geometry (calculations)

Then, from the figure: Sin α = c-l_ r Sin(α+θ) = y+x_ r Sin(α-Ø) = y-x_ r Adding the above two equations, we get Sin(α+θ) + Sin(α-Ø) = 2y = 2Sinα r The unknown quantities while designing are α, length of the tie/ track rod (l) and length of the steering arm( r ).

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• Determination of the steering arm angle α

The angle which the steering arm makes with the centre line can be found out geometrically by drawing the given diagram in Solidwork or by hand.

Fig.51. Determination of the angle

• Determination of the Tie Rod Lengths The method which is used to determine the length of the steering tie rod is the method of virtual centers. Initially, the position of the outer tie rod joint U is assumed by assuming the length of the steering arm as 150mm. Hence the value of r is 150mm. The front view length of the steering arm is found out by the method of projection. From this method K= r Sin α =150 Sin (19.06) = 48.98mm= 50 mm approx. First, the instantaneous center P1 is calculated so that it can be connected to U. The extension of the paths EG and DC gives point P2, which is also required and from which a line to point P1 must be drawn. If UP1 is above GD, the angle α (this angle is applicable only for the above drawing) enclosed by the two must be moved up to P1P2.A line drawn from P1 at the angle α must be made to intersect with the extension of the connecting path UE to the tie rod virtual/instantaneous center P3. To calculate the desired point T, that is the center of the inner tie rod end joint, P3 is connected to C and extended.

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Fig.52.Position of the rack in front the front axle axis

Fig.53.Determination of tie rod length

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Fig.54.Determination of tie rod length in SOLIDWORKS Hence the length of the tie rod is roughly 13.5 inches. From the given data, let us examine the how far the Ackermann criteria is satisfied. The front inner wheel lock angle is given by(θ)=35 degrees r = Steering arm length=150mm l/2=Length of the half track/tie rod=13.5 in= 342.9 mm c= Distance between pivot centers=47in=1193.8mm Sin (19.06+35) + Sin (19.06 -25)=0.70 2Sin α = 0.66 Hence the Ackermann criterion is roughly satisfied by the given design.

• Bump And Roll Steer Steer with ride travel is very common in all terrain scenarios. Steer with ride travel is undesirable because if the wheel steers when it runs over a bump or when the car rolls in a turn, the car will travel on a path that the driver did not select intentionally. Ride/Bump and roll steer are a function of the steering geometry. If the tie rod is not aimed at the instant axis of motion of the suspension system then steer will occur with ride because the steering and suspension are moving about different centers. If the tie rod is not the correct length for its location then it will not continue to point at the instant axis when the suspension traveled in ride. Thus the choice of tie rod location and length is important. If the tie rod height and angle are adjustable it is usually possible to tune most of the ride steer out of a suspension. Curved ride/bump steer plot as shown in figure below are to be avoided because they result in a net change in toe with ride and the steer effect changes from under steer to over steer depending on the wheel ride position. If ride steer plot is curved then a possible

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solution is to raise both ends of the tie rod to move it closer to the shorter, upper A arm, with the tie rod angle can also be adjusted.

Fig.55.Variation of toe in/out during bump/rebound

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• Kingpin

It is common for the wheel to be offset laterally form the point where the steer rotation axis intersects the ground. The lateral distance form the ground intercept to the wheel centerline is the offset at the ground and is called scrub radius. It is considered positive if wheel is outboard of the ground intercept. In the front view kingpin geometry shows the kingpin inclination, spindle length and scrub radius. Some factors to consider are:

• The more the kingpin inclination is tilted from the vertical the more the car will be raised when the front wheels are steered, unless the kingpin inclination is true vertical.

• Kingpin inclination affects the steer-camber characteristics. When the wheel is steered, it will lean out at the top, towards positive camber, if the kingpin is inclined in the normal direction i.e. towards the center of the car at the upper end. The amount of this effect is small.

• Driving or braking forces introduce steer torques proportional to the scrub radius. If the driving or braking forces are different on left and right wheels then there will be a net steering torque felt by the driver.

In side view kingpin geometry gives caster and mechanical trail caster angle results when the steer rotation axis is inclined in the longitudinal plane or side view. Some factors to consider are:

• More trail will give higher steering force. • Caster angle causes the wheel to rise and fall with steer. Unlike kingpin

inclination this effect is opposite from side to side. With equal positive caster on left and right wheels, the effect of left steer is to roll the car to the right, causing a diagonal weight shift. Thus an over steer effect is caused.

• With positive caster angle outside wheel will camber in a negative direction while the inside wheel cambers positive direction.

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c)Oversteer Tendency The car is said to oversteer when the rear wheels do not track behind the front wheels but instead slide out toward the outside of the turn. The tendency of a vehicle to oversteer is affected by several factors such as mechanical traction, aerodynamics and suspension, and driver control. The basic condition is that rear slip angle is greater than front slip angle. Limit oversteer happens when the rear tires exceed the limits of their lateral traction during a cornering situation before the front tires do. Rear wheel drive cars are generally more prone to oversteer, in particular when applying power in a tight corner. This occurs because the rear tires must handle both the lateral cornering force and engine torque. The car’s tendency toward oversteer is generally increased by softening the front suspension or stiffening the rear suspension. Camber angles, ride height, and tire pressures can also be used to tune the balance of the car. In view of the above factors, the Mini Baja vehicle is expected to oversteer as the center of gravity is behind the centerline of the vehicle. Hence, the centrifugal force while turning would tend to form a couple (rotating force).

• Factors Affecting Understeer Or Oversteer: 1) Critical Speed:

Oversteering vehicles have an associated instability mode, called the critical speed. As this speed is approached the steering becomes progressively more sensitive. At the critical speed the yaw velocity gain becomes infinite, that is, the car will continue to turn with the wheel held straight ahead. Above the critical speed a simple analysis shows that the steer angle must be reversed (counter steering).Understeering vehicles do not suffer from this, which is one of the reasons why high speed cars tend to be set up to understeer. But in the case of

Vcrit = √ (-57.3 L g/K) K is called oversteer gradient. Its unit is deg/g =-2.14 deg/g (From Milliken and Milliken) L=68 inches=1.7272 m V=√ (-57.3 x 1.7272x 9.81 /-2.14) = 73 Kmph K is negative in value, thus the whole equation having a positive value. The critical speed is dependent on wheel base of the vehicle; for a given level of oversteer, long wheel base vehicles have a higher critical speed than short wheel based. An oversteer vehicle can be driven at speeds less than critical speed, but becomes directionally unstable at and above it. There is an apparent gain in yaw rate and lateral acceleration, making it unstable.

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2) Yaw Velocity Gain:

A second reason for steering a vehicle is to change the heading angle by developing a yaw velocity (yaw rate). The yaw velocity, r is the rate of rotation in heading angle and is given by: r = 57.3 V/R (deg/sec) Substituting this in the lateral acceleration expression, r/δ = (V/L)/(1 + [KV²/57.3 L g]) Let the speed be around 50kmph or 13.88 m/s which is less than the critical speed. Hence, for this speed, the ratio of yaw velocity to steering angle r/δ = (V/L)/(1 + [KV²/57.3 L g]) = (13.88/1.7272)/ (1+ [0.24x13.882 /57.3x1.7272x9.81]) = 7.67 sec-1 This ratio represents a “gain” which is proportional to the velocity in the case of a neutral steer vehicle, which can be seen in figure. In over steer, the yaw velocity gain becomes infinite when the speed reaches the critical speed. Thus the characteristic speed is that which gives maximum yaw response.

Fig.56.Variation of yaw velocity gain with speed

3) Sideslip Angle:

When Lateral acceleration is negligible, the rear wheel tracks inboard of the front wheel. But as lateral acceleration increases, the rear of the vehicle must drift outboard to develop the necessary slip angles on the rear tires. At any point on the vehicle a sideslip angle may be defined as the angle between the longitudinal axis and the local direction of the travel. In general, the sideslip angle will be

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different at every point on the car during cornering. For any speed the side slip angle,β, at the CG will be:

β = 57.3 c/R – αr = 57.3 c/R – Wr V²/(Cαr g R) From the model, the center of the gravity is at a distance of 27 inches, hence c=27 inches Here, Wr = Weight on the rear axle = 60% of the total weight= 2058 N Cαr = Cornering Stiffness = Experimental data

4) Static Margin:

It is the distance the neutral steer point falls behind the CG, normalized by the wheel base = e/L. When the point is behind the CG, it is positive. This is determined by the point on the vehicle where a side force will produce no steady-state yaw velocity (neutral steer point). The neutral steer line is the locus of the points in the x-z plane along which external lateral forces produce no steady state yaw velocity.

• Torques About The Vehicle C.G • Understeer Torque: The conditions under which the understeer torque is observed:

1. Lateral load transfer between the front wheels 2. Longitudinal load transfer to rear wheels. 3. Cornering drag on front tires. 4. Increased rear or decreased front aerodynamic down force. 5. Unfavourable front tire camber angles. 6. Bottoming of the front suspension. 7. Pulling the inside front tire off the road while it is in a partially laden

condition. 8. Increasing the relative front brake ratio. 9. Locking of the front brakes.

• Oversteer Torque: The conditions under which the oversteer torque is observed: 1. Lateral load transfer between the rear wheels. 2. Excessive accelerative thrust on the rear tires. 3. Unfavourable rear tire camber angles. 4. Decreased rear or increased front aerodynamic down-force. 5. Longitudinal Load transfer to the front tires. 6. Bottoming the rear suspension. 7. Pulling the inside rear off the road in droop. 8. Increase in rear braking effort. 9. Locking of the rear brakes.

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10.Effects Of Different Parameters On Vehicle Dynamics And Stability

1. The effect of CG location on straight line braking - When the brakes are applied, load is transferred from the rear tires to the front tires. The higher the CG the more is transferred from any given deceleration. Tires are load sensitive, thus theoretical best braking occurs with the tires evenly loaded. For the typical RWD or FWD road car this says that to improve the absolute level of braking, the CG should be as low as possible and moved towards the rear of the car.

2. The effect of Tires and Rim sizes on straight line braking – To the extent that changing the tires/rims change tier/road friction properties, the braking performance will be affected. The end with the stickiest tires can generate relatively more braking force for the load on it.

3. The effect of Brake Balance on straight line braking – Brake balance is the name given to the proportioning of brake force or braking torque to the front and rear tires. The brake balance to give “correct” proportions of braking to the front and rear varies with deceleration rate. The harder the stop, the more heavily loaded will be the front wheels and the more braking effort they can support. Similarly the rear tires are unloaded as the deceleration increases and they must have less braking force. This is accomplished in several ways; either the brake balance is fixed and is biased heavily towards the front, which means the rears don’t do their share on relatively gentle stops, or various types of proportioning values are used to limit hydraulic pressure to the rear brakes to prevent premature locking. Finally, an anti-lock system may be fitted. If the rear locks first, the car will tend to spin, based on the destabilizing side force that the still rolling front tires provide. If the front locks up first, steering control will be lost and the car will go straight or slide down the camber of the road. Because the desirable brake balance varies with the deceleration, different brake balance is required on different coefficient of friction surfaces.

4. The effect of CG location when braking and cornering - The result of combined braking and cornering on turn entry is to load the outside front wheel very heavily, as much as one half the weight of the car. At the same time, the inside rear wheel is lightly loaded. The higher the CG, the more load will be transferred. This situation will be exaggerated if the CG is forward. The heavily loaded outside front tire will be operating at a relatively low coefficient because of load sensitivity. A low and rearward CG position is going to keep the tires most evenly loaded on turn entry and this will give the best performance.

5. The effect of roll center location when braking and cornering - Roll center heights front and rear partially determine the way the roll moment on the car from lateral force is distributed. Lowering the roll center on one end will lower the roll moment resisted by that tend; the wheels on that end will be more evenly loaded in cornering compared to the other end of the car. Too high a roll center leads to jacking.

6. The effect of Brake balance on combined braking and cornering - The correct brake balance for straight line stopping may not be appropriate on turn entry. The outside front tire is very heavily loaded and generates relatively more lateral force than the rear leading to spin. This is true even though the front is operating at a lower coefficient. When there is no rear brake bias or braking while on the throttle in a front wheel drive there is a compensating effect. The engine power reduces the braking on the front wheels while the rears are still receiving a normal amount of braking force. The result is that the rear tires are saturated and the car begins to spin; carefully controlled, this can be used to get the tail out on entry to tight corners.

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7. The effect of roll stiffness distribution when braking and cornering – The roll stiffness distribution is the other way of changing the loads on the wheels under lateral force. Roll stiffness can be raised by increasing anti-roll bar stiffness or increasing spring stiffness. If the car is loose on turn entry, higher roll stiffness will help to resist more of the body roll moment on the front.

8. The effect of CG location on steady state cornering – A neutral car is best for steady state cornering. By turning, cars with a range of CG positions near the center of the car can be made to be neutral. If the CG is forward, the lightly loaded rear end will be degraded to bring the car back to neutral. The best use of equal-sized tires in steady state cornering is made with the CG near the center of the car.

9. The effect of roll center location on steady state cornering – If the car is forward weight biased, a rear roll center higher than the front will tend to make it neutral. If both roll centers are so low that the car has a large amount of body roll, absolute cornering performance may be affected through adverse tire camber. It is necessary to strike a compromise here because too high a roll center leads to jacking and undesirable characteristics.

10. The effect of camber on steady state cornering – It is desirable to have a small amount of negative camber on outside wheels. This produces the maximum lateral force from the two outside tires. For race course which have same direction of turn, positive camber in the inner wheels also helps. Camber can also be used to balance the car.

11. The effect of tire and rim size in cornering – Cornering stiffness is often a function of tire/rim size, aspect ratio, and width. A higher cornering stiffness tire requires lower slip angles to produce a given amount of lateral force. A lower slip angle means lower scrub and less speed loss in cornering. The optimum rim width may also play a part in maximizing the total grip available from a given tire.

12. The effect of roll stiffness distribution on steady state cornering – For a symmetrical car, the roll stiffness would ideally be the same on front and rear, if steady cornering were to be optimized. On the other case the drive degrades the lateral force capability at that end and the roll stiffness is biased towards the un driven end. Thus while making the anti-roll bars calculations must be done carefully so that they optimize their double duty.

13. The effect of Roll Stiffness Distribution on acceleration Out of a Corner – Roll Stiffness is the easy way to change lateral load transfer distribution. For rear wheel drive the tendency is to spin the inside rear; more roll stiffness on the front (less on rear) will help this. Whereas, acceleration from low speed can reduce the front tire load so much that the car plows.

14. The Effect of CG Location on Straight Line Acceleration- The CG location determines the point of wheel spin. As the CG moves further rearward, in a RWD, the traction available increases. Traction is a problem at low speeds (low gear=high torque). The CG should be as low as possible to avoid Weight transfer. The CG should be toward the center of the car.

15. The Effect of Anti-Pitch Characteristic on Straight Line Acceleration- RWD Cars especially those for drag racing, may profit from rear lift (anti-squat). This raises the CG and increases weight transfer to the rear wheels on acceleration. The lift effect is created by choosing the rear suspension attachment point to give a high pitch center. The torque reaction from the driving wheels lifts the car.

16. The Effect of Brake Balance on Combined Braking and cornering at high Speed- The best braking balance for high speed turn for front heavy cars is is putting more brake on the front as the car corners harder. Too much braking on the lightly loaded rear during cornering may result in spin.

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17. The effect of CG location while braking in a turn - When the brakes are first applied, a large amount of load shifts from the rear to the front axle. This changes the tire operation loads and side forces and the car tucks in. lowering the CG reduces the load change on braking.

18. The effect of brake balance while braking in a turn – The large transient effect of brake application is to transfer load forward and change the loads on the tire. Brake balance also can affect this transient through the friction ellipse effect. If the rear of the car is coming around too much on brake application, shifting the brake balance forward will reduce the side force available from the front tires and effectively increase the side force at the rear. Lockout or proportioning valves can change this behavior but require adjustment for different track friction coefficient.

19. The effect of steering axis geometry on poor road behavior – Kingpin inclination and kingpin lateral position determine the scrub radius measured at the ground. Negative scrub radius tends to stabilize the car in straight running when the two wheels are on different coefficient surfaces under braking or traction. For poor road straight running, this is a good thing. Un-driven front axles ideally have a small scrub radius. This reduces steering torques due to one-wheel bumps. Large brakes and suspension links often conflict with centering the tire print on the kingpin. So in these cases the steering system must be designed to accept the shock loads. The caster angle and longitudinal kingpin location determine the trail.

20. The effect of ride or roll steer on poor road behavior -Ride steer is geometric effect which results in the wheel steering with ride motion. Ride/roll steer is often built to influence low lateral acceleration handling. Small changes in the wheel steer angles will have little effect on the limit handling because the tires are nearly saturated. Ride steer steers the car with bump travel when traveling straight. Ride steer and roll steer are closely related but they load the steering system in different directions depending on the detailed geometry. If there were no compliance in the steering system or suspension, ride steer and roll steer would be just a function of the wheel ride position to the chassis.

11.Gear Ratio Calculation

The gear ratios are calculated keeping two things in mind:

a) Maximum torque required b) Maximum power required

Each of them is related to the tractive force. This tractive force is calculated by:

1) The first term on the right hand side is the engine torque multiplied by the overall gear ratio and efficiency of the drive system, then divided by the tire radius. This term represents the steady state tractive force available at the ground to overcome the road load forces of aerodynamics and rolling resistance, to accelerate, or to climb a grade.

2) The second term on the right hand side represents the loss of tractive forces due to the inertia of the engine and drive train components. The term in brackets

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indicates that the equivalent inertia of each component is amplified by the square of the numerical gear ratio between the component and the wheel.

Hence, by back substitution of the tractive force according to the given requirement of either the maximum power or maximum toque, the gear ratio can be calculated and the gear ratio required comes in the range of around 8:1 to around 24:1 for the CVT.

12.Hub assembly The hub assembly is one of the most important components as the entire weight of the vehicle is transferred to the hub assembly. Here, while designing, as always, the worst case scenario was thought off. This happens when the suspension fails and the entire load acts on the assembly. This case is simulated here in ANSYS. As is clearly demonstrated by the graphics, the von misses stresses are under control. • Model Used: Rear Hub Mounting Bracket

• Loading: F= 6500 N • Boundary Conditions: All DOF =0 at the center sleeve • Element Type- Solid 45

Fig.57.Deformed Shape while analysis

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Fig.58.Variation of Von-Misses Stresses

13.Manufacturing Strategy

The following steps can briefly summarize the manufacturing strategy:

Order Materials/Parts – As per the design presented will, the group will need to order the necessary materials for fabrications and standardized parts. A thorough market survey of the concerned parts has already been completed with a list of probable vendors and the approximate price of the product.

Machine Parts/ Welding of frame - The design will need the fabrication of several parts. Prominent among them are the hubs of the vehicle which include typical operations like:

1) Turning/ Facing on the lathe 2) Press Fitting of the bearings 3) Milling of splines 4) Hardening of the drive shafts 5) TIG welding of the frame

Assembly - Once all the ordered parts and machined parts are ready, the components will be assembling onto the Baja. Those safety features needing maintenance will be addressed and the overall aesthetics of the vehicle will be improved. This includes sheet metal work (Tinkering) and subsequent body painting.

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Fig.59.Making prototype hub sleeve on centre lathe

Fig.60. Boring of the hub sleeve

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• Manufacturing Simulation

The parts to be manufactured majorly consist of those in the hub assembly. Before finalizing the strategy for manufacturing, a thorough knowledge of the processes is necessary. The simulation can be done on SOLIDCAM. The advantage of this simulation is that the process time is estimated and the suitable process and sequence is selected.

The basic terminology followed in SOLIDCAM is:

i) Profile/ Pocket Milling - Profile milling is a milling process through which outside contours can be machined. The pocket milling is the process through which the inner slots and profiles are made.

ii) Down Step- This is the thickness of the workpiece machined per cycle.

The methodology followed in doing the manufacturing simulation is:

i) Importing of model from Solidworks ii) Giving the required size of the blank iii) Specifying the coordinate axis system and the controller iv) Specifying the tool path and starting the simulation v) Generating reports and documentation.

Fig.61. Profile Milling of Rear Hub

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Fig.62. Pocket Milling of the Rear hub

Fig.63. Stock Model of the finished product

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Fig.64. Tool Path figure in Pocket Milling

Fig.65. Tool Path figure in Profile Milling

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• Report Name : HUB1 ------------ Program number : 5000 Subroutine number : 1 Jobs : 2 ---- Pocket : 1 Profile : 1 Total time : 6:58:27 Tool Diameter Corner rad. Tool type Number Length H D ------------------------------------------------------------------ 1 10.0000 0.0000 TOOL END MILL 2 50.0000 1 1 2 10.0000 0.0000 TOOL END MILL 2 50.0000 2 2 Job Calls --------- 1. F_top_T1 Time: 1:29:06 2. P_pocket_T2 Time: 5:29:21 Job name : F_top_T1 -------------------- Job type : Profile Geometry name : top Time : 1:29:06 Tool ---- Tool number : 1 Spin Feed ---- ---- Spin type : S Feed type : F Spin Rate : 1000.0000 Feed XY : 100.0000 Feed Z : 33.0000 Milling levels Define depth : Constant -------------- Clearance level : 50.0000 Safety distance : 2.0000 Job Upper level : 0.0000

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Profile depth : 47.0000 Down step : 5.0000 Descent type : FEED Rest material \ Chamfer ----------------------- Rest material \ Chamfer : None Offsets Clear offset ------- ------------ Offset on profile : 0.0000 Offset : 0.0000 Finish : None Side step : 0.0000 Profile direction : Default Tool side --------- Tool side : Left Compensation : YES Extension distance : 0.0000 Approach Retreat -------- ------- Approach : NONE Retreat : NONE Extra parameters Home number : 1 ---------------- Extra parameters : OPT1 DELY : 0.0000 FEAD : 0.0000 Job name : P_pocket_T2 ----------------------- Job type : Pocket Geometry name : pocket Time : 5:29:21 Tool ---- Tool number : 2 Spin Feed ---- ---- Spin type : S Feed type : F Spin Rate : 1000.0000 Feed XY : 100.0000 Feed Z : 33.0000 Pocket type Milling levels ----------- --------------

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Pocket type : CONTOUR Clearance level : 50.0000 Direction : Climb Safety distance : 2.0000 Connect islands : No Job Upper level : 0.0000 Exit material : No Pocket depth : 47.0000 Corner : NONE Down step : 5.0000 Offsets Overlap : 0.6500 -------------- Compensation : YES Wall offset : 0.0000 Island offset : 0.0000 Floor offset : 0.0000 Finish : None Retreat Approach ------- -------- Retreat : None Approach type : None Rest material \ Chamfer Extra parameters ----------------------- ---------------- Rest material \ Chamfer : None Extra parameters : OPT1 DELY : 0.0000 FEAD : 0.0000 Home number : 1

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Fig.66. Profile milling of hub base plate

Fig.67. Drilling of holes on hub base plate

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Fig.68. Pocket Milling on hub base plate

• Report Name : REARBASEPLATE-SOLDCAM ---------------------------- Program number : 5000 Subroutine number : 1 Jobs : 3 ---- Profile : 1 Drill : 2 Total time : 0:20:14

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Tool Diameter Corner rad. Tool type Number Length H D ------------------------------------------------------------------ 1 10.0000 0.0000 TOOL END MILL 2 50.0000 1 1 2 10.0000 118.0000 TOOL DRILL 2 50.0000 2 2 3 40.0000 118.0000 TOOL DRILL 3 200.0000 3 3 Job Calls --------- 1. F_1_T1 Time: 0:17:20 2. D_circle baseplate_T2 Time: 0:01:57 3. D_centre1_T3 Time: 0:00:58 Job name : F_1_T1 ------------------ Job type : Profile Geometry name : 1 Time : 0:17:20 Tool ---- Tool number : 1 Spin Feed ---- ---- Spin type : S Feed type : F Spin Rate : 1000.0000 Feed XY : 100.0000 Feed Z : 33.0000 Milling levels Define depth : Constant -------------- Clearance level : 50.0000 Safety distance : 2.0000 Job Upper level : 0.0000 Profile depth : 8.0000 Down step : 6.0000 Descent type : FEED Rest material \ Chamfer

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----------------------- Rest material \ Chamfer : None Offsets Clear offset ------- ------------ Offset on profile : 0.0000 Offset : 0.0000 Finish : None Side step : 0.0000 Profile direction : Default Tool side --------- Tool side : Right Compensation : YES Extension distance : 0.0000 Approach Retreat -------- ------- Approach : NONE Retreat : NONE Extra parameters ---------------- Extra parameters : OPT1 DELY : 0.0000 FEAD : 0.0000 Job name : D_circle baseplate_T2 --------------------------------- Job type : Drill Geometry name : circle baseplate Time : 0:01:57 Tool ---- Tool number : 2 Spin Feed ---- ---- Spin type : S Feed type : F Spin Rate : 1000.0000 Feed Z : 33.0000 Milling levels Drill Parameters -------------- ---------------- Clearance level : 50.0000 Drill cycle type : Drilling Safety distance : 2.0000 Job Upper level : 0.0000 Drill depth : 10.0000 Sort sequence by : DEFAULT Extra parameters

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---------------- Extra parameters : OPT1 DELY : 0.0000 FEAD : 0.0000 Job name : D_centre1_T3 ------------------------ Job type : Drill Geometry name : centre1 Time : 0:00:58 Tool ---- Tool number : 3 Spin Feed ---- ---- Spin type : S Feed type : F Spin Rate : 1000.0000 Feed Z : 33.0000 Milling levels Drill Parameters -------------- ---------------- Clearance level : 50.0000 Drill cycle type : Drilling Safety distance : 2.0000 Job Upper level : 0.0000 Drill depth : 20.0000 Sort sequence by : DEFAULT Extra parameters ---------------- Extra parameters : OPT1 DELY : 0.0000 FEAD : 0.0000

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Fig.69. Fabrication of die set for casting

Fig.70. Fabrication of die set for casting

An alternative method of fabricating the hub assembly can be casting of individual components in SOLIDWORKS/ MOLDFLOW EXPRESS . This type of computer modelling clearly helps in decision making.

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14.CONCLUSION

The Mini Baja that team Jabaaz intends to submit for entrance in the SAE Mini Baja Oregon competition was a collaborative design effort among students from several engineering disciplines be it mechanical or electrical. The team gained thorough knowledge in the field of designing through the virtual design contest. 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.

The team relied on individual member’s knowledge and experience with off-road vehicles as a tool for developing many of the initial subassembly designs for the prototype. Team members who attended the 2007/2008 SAE competition helped invaluably to gather ideas and information about what design choices were successful and how they could be incorporated into the prototype design.

Where applicable, selection of components for each subassembly of the prototype was based on engineering knowledge, through benchmarking. Reliance upon “engineering intuition” governed the selection of the remaining components. Computational design and analysis software like ADAMS/ANSYS were used to verify that each part of a subassembly design met or exceeded its stated objective. Use of these design tools also allowed the team to address and rectify conflicts between interfacing subassemblies before fabrication, saving both time and cost. Additionally, inventory of parts in DFMA software would be created so that parts could be readily duplicated.

Parts not machine able in-house would be out sourced to qualified professionals. In some cases after market parts common to certain off-road vehicles currently on the market would incorporated into the design for both convenience and because they would be readily accessible.

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. Employing an A-arm suspension provides a durable, less complicated system over other proposed alternatives. A modified rack and pinion steering system provides less road response after taking care of bump steer. The use of a gearbox in conjunction with a CVT allows for a broader range of gearing ratios. The diagonal braking system provides safety in case of brake failure. Each of these design features was incorporated into the designing stage in an attempt to produce a superior off-road recreational vehicle.