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SUPRA SAEINDIA 2011 – ANSYS CAE PAPER Team Registration ID: 607736 (Customer ID) Author: Tejas Ulavi

Co-Author: Nipun Kuzhikattil

The various cases for the static simulation and

ABSTRACT

analysis of the chassis or rollcage are as

The

Supra

SAE

design

competition

follows-

provides a unique challenge for designing a formula type racing car and test it in the real-world situation. While simulating real world situations is difficult and can be obtained

by

complex

analytical

formulations, the advent of CAE has made

– In this case, the front of the 1. Front impact –

car, disregarding the impact attenuator is considered to collide with a stationary object in a head-on collision at maximum speed with an impact time of 0.3 sec.

the job of the engineer easier, given he

2. Rear impact – impact – In this case, another car is

provides appropriate inputs.

considered to collide head-on with the rear of the car at maximum speed with an impact

Tools like ANSYS, MATLAB, MSC Adams,IDEAS, etc help simulate real-life situations and loading conditions and provides a way to validate results. An attempt has been made to provide a comprehensive insight

time of 0.3 sec. 3. Side impact – In this case, a sideways

impact into an obstruction is considered at the maximum speed with an impact time of 1.2 sec. (This is a safe case of side rollover)

into the CAE used by team OCTANE RACING for the design stage of the

4. Rollover impact – In this case, overturning

competition in three sections.

or rollover of the chassis is considered and the effect of self weight is considered as an

1. ROLLCAGE -

impact load.

1.1 Introduction

5. Front wheel bump – In this case, a front

wheel is considered to go into full bump with The static finite element analysis of the

all other wheels fixed.

rollcage was done in ANSYS, and several realworld situations and loading conditions were

6. Rear wheel bump – In this case, a rear

simulated as representative of worst-case

wheel is considered to go into full bump with

scenarios. The ultimate aim was to ensure a

all other wheels fixed.

fully functional, weight-effective and sturdy vehicle that can survive harsh test conditions.

1.2 Problem Description

7. Torsional rigidity - The torsional rigidity of

the frame is determined by applying an equal and opposite bending moment on the chassis and quantifying the angular displacement.

The aim of the analysis is to carry out a design check of the given Mini Baja chassis under

1.3 Simulation Methodology and

estimated loading conditions and to minimize

parameters

the weight of the frame (limit it to 35 kg) keeping a Safety Factor of 1.5. Material of the tubes is AISI 1020, Hot Rolled with properties Sut = 394.7 MPa Syt = 294.8 MPa

A geometric model of the rollcage was constructed in Pro-E and was imported into ANSYS Workbench in IGES format. ANSYS was used to create a finite element formulation of the problem for static structural analysis.



The Shell81 element was used for meshing

Calculating thus,

the entire rollcage, with real constant as the

Front/rear impact – F = 15000N

thickness of the pipes. This was more

Side impact – F = 5000N

convenient than the pipe element owing to

Rollover impact – 6000 N ((weight of car

the incorporation of a number of pipes of

+driver) X 2)

different diameters, and cross-sections, and the presence of square and rectangular pipes.

The meshing was done globally with a size of 3mm, with local mesh size at the area of interest as low as 1mm. Smooth transition in the mesh size was ensured. The local variation of the mesh size enabled us to achieve good convergence with minimum strain energy

error (less than 7% in the area of interest) without

compromising

seriously

on

the

computational speed and size.

within the existing library in Workbench. The properties of Structural steel were modified for AISI 1020 steel , with the most important linear, isotropic properties being

estimation

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

These two point loads will be equal to the weight of the chassis. Hence, 2F = m1 * g F = ½ m1 * g F= ½ *300 * 10 F = 1500 N

Ex = 210000MPa, nuxy = 0.30

Force

An assumption is made that when the vehicle

Chassis to fail in case the suspension fails.

The material properties were specified from

1.4

Estimation of wheel bump forces -

for

loading

conditions

Hence, designing for F = 1500N (approx). Similarly, F = 2500 N for the rear wheel bump condition.

Estimation of Impact force – By the laws of motion,

1.5 Boundary conditions

v = u + a.t

The various loading conditions and the

For impact analysis, consider u = max. speed

boundary conditions assumed for each of the

(150 kmph i.e. 41.67 m/s)

fore-mentioned analyses have been carefully

v = 0 (after impact, perfectly inelastic

formulated. For many of the analyses pseudo

collision)

boundary conditions are assumed to constrain

t = time of impact

the model so that a realistic simulation result is obtained. The assumptions are enlisted in

From the above equation, calculate the value of “a” which is the G’s of acceleration

witnessed by the rollcage during the impact. Now, F = m.a, gives the impact force to be applied to the members.

the Table 1.1.



1.6 Simulation Results Preliminary results and conclusions –

required in the rear and side impact members. In future, alternative materials for the rollcage will be looked at as a viable solution for weight reduction.

The preliminary design was analyzed for the above tests and the results have been tabulated in Table 1.2. Also, the various

2. SUSPENSION AND OTHER STRUCTURAL COMPONENTS

contour plots for the displacement and Vonmisses stress have been shown in Fig. 1.1 to Fig. 1.6 in Appendix 1.

The results showed that the design would fail for the side impact and rear bump conditions with FOS 0.90 and 0.73 respectively. Also, the rear impact condition had a low FOS of 1.18, which is less than the aimed minimum FOS of 1.5. Torsional rigidity is important to prevent excessive frame flexure during operation. We

2.1 Introduction The structural integrity of the vehicle and coexistence of all the subsystems depends upon

proper material selection and appropriate strength of the components individually and as a unified entity. ANSYS was used to validate the design and to analyze pivotal structural components such as the knuckle, bellcrank, rotor assembly and the brake assembly.

2.2 Front/rear knuckles

created a rigid frame by including structural members in key locations. The torsional

The front and rear knuckles were both

rigidity analysis involved fixing the rear of the

designed specifically for the application of the

frame, applying a torque to the front of the

competition. Owing to the low weight budget

frame, and measuring the deflection. Our

and the high strength requirement, Al6061

frame was found to have a torsional rigidity

was chosen as the material for the knuckles.

of 5240 N-m/degree without any sign of yield with a factor of strength of 1.63.

The aim of the analysis was to determine the

best profile of the knuckle to satisfy the Iterative Design & Re-validation -

strength requirements. Primary calculations were done using a 3G vertical, 2G lateral and

Considering the results obtained by preliminary analysis, certain changes were implemented in the design of the rollcage. The major identified areas for the proposed changes and the modified rollcage are shown in Fig 1.7.

1G longitudinal force template

on the

knuckle. Thus forces on both the front and rear knuckle were found – The factor of safety requirement was again fixed at 1.5 as earlier. Also, Al6061 as the material offers excellent material strength

The results after modification are tabulated in Table2.3.

properties -

1.7 Conclusions & Recommendations

Sut = 310 MPa

Syt = 276 MPa The different analyses carried out in ANSYS

The final achieved weight of the rollcage is about 40 kg, 5 kg more than the proposed 35 kg. This was due to the strengthening

included – 1. Static structural analysis



2. Shape optimization analysis

Front knuckle –

3. Fatigue analysis

Lower ball joint, Fx = 1300N, Fy = 0N, Fz = -

2.2.1

Simulation

methodology

and

parameters

5813N Upper ball joint, Fx = Fy = 0N , Fz = -1700N Spindle, Fz = 1500, Mz = 228600 N.m

The knuckles were designed in Pro-E and

Fixed support at the inner race of the knuckle

imported into ANSYS. Minor modifications

where the spindle rests.

whenever required were made in the ANSYS modeler.

For the rear knuckle, the drive shaft goes through the knuckle and is a moving part.

The Solid45 element was used for meshing

Hence, a bearing is fitted in the hub of the

the knuckle as a three-dimensional entity.

knuckle. The forces for the rear knuckle are

This was convenient given the complex

determined as were for the front.

geometry of the knuckle and a thickness of almost 2 inches. Hence, the plate or shell element could not have been effectively used to represent the geometry.

Rear knuckle -

Lower ball joint, Fx = 1500N, Fy = 0N, Fz = 6310N

Meshing was done with a size of 1 mm due

Upper ball joint, Fx = Fy = 0N, Fz = -1600N

to the small size of the knuckle. Initially, local

Spindle, Fz = 2500N

size was not tampered with for the first run.

“Pseudo” fixed support at the inner race of

After the first run results were obtained, local

the knuckle where the drive shaft rests.

element sizing was enhanced for greater convergence. In the final iteration, the mesh

2.2.3 Simulation results and iterative

size locally was as low as 0.7 mm.

process

The material chosen from the library was

Static structural analysis –

Al6061 alloy. The properties were changed as per requirements. The typical properties of

The results for the static structural analysis of

Al6061 are

the front and the rear knuckles are shown in Fig 1.8 and Fig 1.9. The following Table 2.4

Ex = 69000MPa, nuxy = 0.33

shows a tabulated result of the analysis –

2.2.2 Boundary conditions

Shape optimization analysis –

For the front knuckle the spindle is stationary

Initially, the knuckle was considered to be a

and the wheel rotates with the help of a

solid block of the outermost dimensions.

bearing which has its inner race on the spindle

These

and outer race on the hub of the wheel. The

preliminary calculations with a FOS of 2 and

forces from the tire are transmitted to the

considering a uniform beam section under

knuckle through the moment arm of the

bending. (Fig 2.2)

dimensions

were

calculated

by

spindle, thus creating a torque. The forces on the tires were thus transferred to the knuckle

The block thus obtained was analyzed for

in addition to the torque produced.

optimization of material. The resulting shape



plot obtained was as shown in Fig. 2.3. This

weight possible, yet sturdy enough to support

was used as a template for further weight

the tire forces.

reduction in the knuckle. The reduced knuckle was iterated till the material reduction and

Material used is again Al6061 and the FOS

strength obtained were optimal.

requirement is 2.

The final design is shown in Appendix 1 and

2.3.1

mentioned under the static structural analysis

parameters

Simulation

methodology

and

earlier. Same as for the front and rear knuckles Fatigue analysis –

2.3.2 Boundary conditions

The knuckle is one of the most important suspension components and is the medium

The bellccrank has a central pivot with either

that joins the wheel to the chassis. The

end supporting the spring one side and the

knuckle thus is thus subjected to constant

pushrod on the other.

fluctuating loads and fatigue failure is an important criteria. Fatigue is also responsible for failure of 50-60 % of the components.

The bellcrank bolts will be loaded in shear as the

pushrod

actuates

the

spring

and

experiences the opposite reaction.

For fatigue analysis, the fatigue tool was

Fixed support – the inner race of the central

added

pivot.

to

the

solution

in

the

ANSYS

Workbench environment. Analysis for life in cycles and factor of strength were calculated. The minimum life of the front knuckle was determined as 98806 cycles and the least

factor of strength as 0.45 (for 10 6 cycles)

2.3.3 Simulation results and weight reduction The results have been tabulated in the Table 2.4. Also the contour plots for the bellcranks

are given in Fig 1.10 . The result obtained still

The minimum life of the rear knuckle was

has a higher FOS than required, even after

determined as 70500 cycles and the factor of

high weight reduction of the cut-outs through

strength 0.39.

a process identical to the knuckle iterative loop.

2.3 Bellcranks

2.4 Wishbones or A-arms

Inboard suspension system in the front

The wishbones or the A-arms of a typical

facilitates the use of bellcranks to pivot the

double wishbone suspension have been

spring and pushrod assembly. The entire road

analyzed for strength. The results for the front

forces are transferred to the spring through

lower wishbone have been showed as an

the bellcranks, hence appropriate design of

example to represent the design process of

the bellcranks is necessary for fatigue.

the suspension A-arms.

The aim of the analysis was to design a

The factor of safety for the A-arms is chosen

functional bellcrank in the least amount of

to be 1.5 and the material as AISI 4130 steel.



Syt = 360 MPa

Fx = 0.3 * 600

Sut = 560 MPa

Fx = 180 N

2.4.1

Simulation

methodology

and

parameters

In addition to this, the pushrod mounting on the A-arm will experience a maximum force equivalent to the weight of the front end,

The Pipe18 element was chosen for the

about 800 N.

representation of the A-arms. This facilitated the construction of a simple line figure in

Hence, the loading and boundary conditions

ANSYS, hence allowing for a flexible design

are,

which could be reiterated or changed easily.

Fx = 180, Fz = 600 at the ball joint/rod end

Also, this would reduce the computational

Fz = 800 N at the pushrod mount

time and endow a simple geometry. Meshing

Fixed support at the rod end bearing and

size was chosen as 1mm uniform, which gave

chassis mounts.

sufficiently good results for the analysis. Material properties for 4130 alloy steel were entered after creating a new material model. The major properties were –

Ex = 320000 Mpa, nuxy = 0.3

2.4.3 Simulation results The contours have been shown in Fig 1.13. The FOS was iterated by changing the crosssection of the Pipe18 elements. An acceptable design is obtained with a hollow circular pipe

of dimensions 20 X 2 mm.

2.4.2. Boundary conditions 2.4.4 Conclusions To find the optimal loading condition for the front geometry –

Weight of front/rear knuckle = 1.023/1.25 kg Weight of bellcrank = 0.77 kg

Total Weight of vehicle + human = 3000 N

The wishbones were also chosen of optimal

Number of Suspension Arms = 4 *2 = 8

cross-section. Hence, targets were achieved.

Assuming 40% force distribution on front arms, Static Upward Force on each arm, Fzs= 3000*0.4/4 = 300 N Since the suspension may be subjected to dynamic loads which can have a maximum value equal to twice the static load, Hence, Fz = 2 * Fzs=2 *300 = 600 N Also, due to rolling motion and friction there will be a load in the direction of motion, which was estimated as 0.3 times the normal load where 0.3 is the estimated co-efficient of friction. Fx = 0.3 * Fz



2.6 Additional CAE for suspension and

MATLAB

dynamic analysis of the vehicle MATLAB is a powerful mathematical tool with multiple applications and a user-friendly

MSC Adams

interface. We used MATLAB to simulate the MSC Adams is a mechanical systems analysis

longitudinal vehicle dynamics of the entire

software which enabled us to study the

vehicle by considering the vehicle to be a two

dynamics of our sub-systems, the interactions

DOF system.

of various sub-systems and thus optimize their design and performance. It helped us to

2.6.3

Simulation

eliminate the need to actually build and test

parameters

methodology

and

our designs.

The entire vehicle was modeled as a 2 degree of freedom (DOF) spring-damper system in

2.6.1 Front Suspension and Steering

MATLAB. Differential equations for the model MSC Adams was used to create a template of

were derived from first principles, and

the suspension and steering system of the

modeled for the gross vehicle parameters.

car into a front assembly. This was done by modifying the hardpoints of an available FSAE template for inboard suspension.

2.6.4 Loading conditions and results Sinusoidal excitation of 10mm amplitude and

The resulting assembly was analyzed for

20 rad/s frequency was given as input to

parallel wheel travel and toe change was

obtain the frequency of front and rear setups,

measured against wheel travel . The height of

and the overall vehicle pitch and bounce for

the steering rack was iterated for minimum

the said excitation. The pitch and bounce

toe change during the suspension travel.

values originally received depended upon the

Hence “Bump Steer” was eliminated. An

damper

optimized graph for bum steer is shown in Fig

Increasing this gain, we were able to achieve

2.5.

reduction in pitch from 0.28 deg to 0.20 deg

gain

in

the

SimuLink

model.

and the bounce from 15 mm to 10 mm.

2.6.2 Full-vehicle assembly The results have been shown in Appendix 2,

We have presently constructed a full-vehicle

Fig 2.6.

assembly in MSC Adams incorporating the suspension,

steering,

chassis,

brake,

2.6.5 Future work

powertrain and engine into a single assembly. Presently, we are working on a MATLAB Results with this assembly for dynamic driving

model for optimizing the lateral dynamics of

conditions are being pursued.

the vehicle. This can be controlled by the yaw rate, yaw acceleration, etc. The MATLAB model is ready, but results obtained are unrealistic at present.



3. AERODYNAMICS



Reduction in the throat area will only serve to reduce the mass flow

Intake Restrictor Analysis



The maximum possible area is defined by

The intake Restrictor is to be fitted in the air

the

rulebook

i.e.

maximum

diameter of 20mm

intake pipe in order to restrict the air flow

Thus, the only design variable is to check the

into the engine in order to limit the speeds

restrictor geometry for flow separation in the

attainable by the engine. The commercially

exit section i.e. the divergent portion of the

available software Fluent was used for

nozzle. That is done on ANSYS 12 FLUENT

simulating the flow through the nozzle.

software.

Aerodynamics

In FLUENT analysis, we used different outlet pressure values and studied the flow pattern.

In the Octane Racing vehicle, we used a front

In analysis, the turbulence model used is k-

wing, a rear wing and a nose cone to

omega SST (shear stress transport) for low to

optimize downforce and drag across the

medium

vehicle. We have used the software FOILSIM

separation was not shown by any result.

available at the NASA website for selecting

Hence the restrictor design in free from

an aerofoil that satisfies our downforce

separation losses.

turbulence

intensity.

The

flow

requirements. The results have been validated by referring a book The Theory of Wing Sections by Ira Abott.

Aerodynamics –

3.1 Problem Description

The downforce required in the Octane Racing vehicle was approximated as one-third of the

Intake restrictor –

downforce available on a Formula 1 vehicle .

The restrictor has to be fit in intake manifold

The downforce generated by the rear wing of

of the engine. So the diameter of manifold as

a Formula1 race car is approximately 450 kg at

actually measured is 35mm. The maximum

300 kmph. Considering the top speed of the

diameter allowed in rulebook for restrictor is

Octane

20mm, that is going to be the throat

downforce required is approximately 170 N

diameter.

(as downforce is directly proportional to

The Inlet pressure is approximately equal to the

atmospheric

pressure,

barring

the

pressure losses in the inlet filter. The outlet pressure will be decided by the engine intake pressure i.e. suction pressure. Due to fixed diameter values of inlet, throat, exit, inlet pressure and outlet pressure, the restrictor is pre-designed for a certain mass flow rate, since –

racing

car

as

100

kmph,

the

square of the velocity). The aerofoil section was chosen using FOILSIM software, and the profile

selected

was

NACA4412

which

provides 162 N downforce at zero degree angle of attack. The flow across the wing was analysed using Ansys 12 Fluent and CFX for validation of the downforce and drag values. Thus far we have been unsuccessful in obtaining realistic values for the drag and downforce. However, we



have used a book THEORY OF WING SECTIONS

incorporate the requirements of the various

by IRA ABOTT for validation.

sub-systems in the present application.

We decided not to use the front wing for

Apart from ANSYS, ANSYS FLUENT has been

obtaining downforce as the requirement for

used to design and validate the intake

the Octane Racing vehicle is minimal. Hence,

restrictor, side-pod and the wings. MSC

we chose the NACA0008 profile for the front

Adams has been used for dynamic simulation

wing; it only serves the purpose of supporting

of steering and suspension systems for the

the end plates.

elimination of bump steer. MATLAB and C++ have been used for the optimizing various

3.2 Simulation results

vehicle dynamics parameters, such as the

The simulation images from ANSYS FLUENT 12 are shown in Fig. 1.14 and Fig. 1.15 in the

longitudinal, lateral and vertical behavior of the vehicle.

appendix.

REFERENCES

3.3 Future work

1. Supra SAE rulebook, 2011 (Version 2)

We hope to get concrete results in the flow

2. Theory of wing sections – Ira Abott

simulation for the validation of the chosen profiles for the rear and front wings. Also, we

3. Race Car Vehicle Dynamics – Milliken and

propose to conduct simulations for nozzle

Milliken

flow analysis using actual pressure values for 4. Octane Racing Preliminary Design Report,

the restrictor.

Supra SAE 2011

FINAL CONCLUSION

5. Fundamentals of Vehicle Dynamics –

CAE is a powerful tool and has been amply utilized by our team throughout the design process as an aid to design and a means for validation of the design. ANSYS has been our primary CAE software, which has been used for analyzing the chassis, optimizing

it

validating

the

for

weight

design

of

and

stiffness,

key

structural

Thomas Gillespie 6. ANSYS Help system (supported by full version ANSYS) 7. Race Car Aerodynamics – Gregor Seljak 8. Finite Element Procedures – K. J. Bathe 9. www.FSAE.com

components like the knuckle, bellcrank, brake assembly, etc. ANSYS Workbench provides a simple interface which offers options for online modification of the design and reevaluation. ANSYS also offers various modules such as static structural, transient structural, modal, thermal, etc. which can be effectively used to

10. www.wikipedia.com


APPENDIX 1 – List of Figures

Fig 1.1. Front impact

Fig 1.2 Rear impact

Fig 1.3 Side impact



Fig 1.4 Rollover impact

Fig 1.5 Front bump


Fig 1.7 Modified rollcage

Fig 1.8 Front knuckle – deformation, stress, fatigue life

Fig 1.9 Rear knuckle

12 | P a g e



Fig 1.10 Bellcrank

Fig 1.11 MSC Adams suspension, steering assembly

Fig 1.13 Front wishbones

Page | 13

Fig 1.12 MSC Adams full vehicle assembly



Fig 1.14 NACA models - Ira Abott

Fig 1.15 FLUENT simulations

Page | 14



APPENDIX 2 – Tables and graphs

No

Type of analysis

1

Frontal impact

Load value

Boundary conditions

5G

Suspension mounts Ux=Uy=0, Rear corner points All DOF=0

2

Rear impact

5G

Suspension mounts Ux=Uy=0, Front corner points All DOF=0

3

Side impact

3G

Right side frame All DOF=0

4

2G

Base All DOF=0

5

Roll over impact Front wheel bump

1500N

1 front+2 rear wheels All DOF=0

6

Rear wheel bump

2 front+1 rear wheels All DOF=0

7

Torsional rigidity

2500N 1320Nm

Rear roll hoop All DOF=0

Table 2.1 – Rollcage boundary conditions

No 1 2 3 4 5 6 7

Displacem Stress ent (mm) ( MPa) FOS

Type of analysis Frontal impact

0.02

77.283 3.80

Rear impact

4.64

250.07 1.18

Side impact

0.53

325.57 0.90

Roll over impact

0.71

74.219 3.96

Front wheel bump

0.43

49.554 5.93

Rear wheel bump

27.01

401.46 0.73

Torsional rigidity

0.78

105.55 2.79

Table 2.2 Rollcage analysis results – old design

Type of analysis

Stress (old)

Stress (modified)

Rear impact

250.07

182.34

1.62

Side impact

325.57

203.45

1.45

Rear wheel bump

401.46

256.01

1.15

Table 2.3 Results for the modified rollcage Page | 15

New FOS



Component

Deformation(mm)

Stress (Mpa)

Front knuckle

0.46855

180.87

1.53

Rear knuckle

0.06475

106.53

2.59

Bellcrank

0.01412

22.127

12.47

Table 2.4 Result for suspension components

Fig. 2.5 Bump steer, MSC Adams

Fig. 2.6 MATLAB suspension, longitudinal dynamics

Page | 16

FOS

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