86415132 Ansys Cae Paper

SUPRA SAEINDIA 2011 ANSYS CAE PAPER

Team Registration ID: 607736 (Customer ID)

Author: Tejas Ulavi Co-Author: Nipun Kuzhikattil

ABSTRACT

The Supra SAE design competition

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

the job of the engineer easier, given he

provides appropriate inputs.

Tools like ANSYS, MATLAB, MSC Adams,I-

DEAS, 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

into the CAE used by team OCTANE

RACING for the design stage of the

competition in three sections.

1. ROLLCAGE -

1.1 Introduction

The static finite element analysis of the

rollcage was done in ANSYS, and several real-

world situations and loading conditions were

simulated as representative of worst-case

scenarios. The ultimate aim was to ensure a

fully functional, weight-effective and sturdy

vehicle that can survive harsh test conditions.

1.2 Problem Description

The aim of the analysis 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 (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

The various cases for the static simulation and

analysis of the chassis or rollcage are as

follows-

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

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.

2. Rear impact In this case, another car is

considered to collide head-on with the rear of

the car at maximum speed with an impact

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)

4. Rollover impact In this case, overturning

or rollover of the chassis is considered and the

effect of self weight is considered as an

impact load.

5. Front wheel bump In this case, a front

wheel is considered to go into full bump with

all other wheels fixed.

6. Rear wheel bump In this case, a rear

wheel is considered to go into full bump with

all other wheels fixed.

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.

1.3 Simulation Methodology and

parameters

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

the entire rollcage, with real constant as the

thickness of the pipes. This was more

convenient than the pipe element owing to

the incorporation of a number of pipes of

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.

The material properties were specified from

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

Ex = 210000MPa, nuxy = 0.30

1.4 Force estimation for loading

conditions

Estimation of Impact force

By the laws of motion,

v = u + a.t

For impact analysis, consider u = max. speed

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

v = 0 (after impact, perfectly inelastic

collision)

t = time of impact

From the above equation, calculate the value

of a which is the Gs of acceleration

witnessed by the rollcage during the impact.

Now, F = m.a, gives the impact force to be

applied to the members.

Calculating thus,

Front/rear impact F = 15000N

Side impact F = 5000N

Rollover impact 6000 N ((weight of car

+driver) X 2)

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.

Hence, 2F = m1 * g

F = m1 * g

F= *300 * 10

F = 1500 N

Hence, designing for F = 1500N (approx).

Similarly, F = 2500 N for the rear wheel bump

condition.

1.5 Boundary conditions

The various loading conditions and the

boundary conditions assumed for each of the

fore-mentioned analyses have been carefully

formulated. For many of the analyses pseudo

boundary conditions are assumed to constrain

the model so that a realistic simulation result

is obtained. The assumptions are enlisted in

the Table 1.1.




1.6 Simulation Results Preliminary results and conclusions The preliminary design was analyzed for the

above tests and the results have been

tabulated in Table 1.2. Also, the various

contour plots for the displacement and Von-

misses 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

created a rigid frame by including structural

members in key locations. The torsional

rigidity analysis involved fixing the rear of the

frame, applying a torque to the front of the

frame, and measuring the deflection. Our

frame was found to have a torsional rigidity

of 5240 N-m/degree without any sign of yield

with a factor of strength of 1.63.

Iterative Design & Re-validation -

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.

The results after modification are tabulated in Table2.3.

1.7 Conclusions & Recommendations 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

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. 2. SUSPENSION AND OTHER STRUCTURAL COMPONENTS 2.1 Introduction

The structural integrity of the vehicle and co-

existence 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

The front and rear knuckles were both

designed specifically for the application of the

competition. Owing to the low weight budget

and the high strength requirement, Al6061

was chosen as the material for the knuckles.

The aim of the analysis was to determine the

best profile of the knuckle to satisfy the

strength requirements. Primary calculations

were done using a 3G vertical, 2G lateral and

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

properties -

Syt = 276 MPa

Sut = 310 MPa

The different analyses carried out in ANSYS

included

1. Static structural analysis




2. Shape optimization analysis

3. Fatigue analysis

2.2.1 Simulation methodology and

parameters

The knuckles were designed in Pro-E and

imported into ANSYS. Minor modifications

whenever required were made in the ANSYS

modeler.

The Solid45 element was used for meshing

the knuckle as a three-dimensional entity.

This was convenient given the complex

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.

Meshing was done with a size of 1 mm due

to the small size of the knuckle. Initially, local

size was not tampered with for the first run.

After the first run results were obtained, local

element sizing was enhanced for greater

convergence. In the final iteration, the mesh

size locally was as low as 0.7 mm.

The material chosen from the library was

Al6061 alloy. The properties were changed as

per requirements. The typical properties of

Al6061 are

Ex = 69000MPa, nuxy = 0.33

2.2.2 Boundary conditions

For the front knuckle the spindle is stationary

and the wheel rotates with the help of a

bearing which has its inner race on the spindle

and outer race on the hub of the wheel. The

forces from the tire are transmitted to the

knuckle through the moment arm of the

spindle, thus creating a torque. The forces on

the tires were thus transferred to the knuckle

in addition to the torque produced.

Front knuckle

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

5813N

Upper ball joint, Fx = Fy = 0N , Fz = -1700N

Spindle, Fz = 1500, Mz = 228600 N.m

Fixed support at the inner race of the knuckle

where the spindle rests.

For the rear knuckle, the drive shaft goes

through the knuckle and is a moving part.

Hence, a bearing is fitted in the hub of the

knuckle. The forces for the rear knuckle are

determined as were for the front.

Rear knuckle -

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

6310N

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

Spindle, Fz = 2500N

Pseudo fixed support at the inner race of

the knuckle where the drive shaft rests.

2.2.3 Simulation results and iterative

process

Static structural analysis

The results for the static structural analysis of

the front and the rear knuckles are shown in

Fig 1.8 and Fig 1.9. The following Table 2.4

shows a tabulated result of the analysis

Shape optimization analysis

Initially, the knuckle was considered to be a

solid block of the outermost dimensions.

These dimensions were calculated by

preliminary calculations with a FOS of 2 and

considering a uniform beam section under

bending. (Fig 2.2)

The block thus obtained was analyzed for

optimization of material. The resulting shape




plot obtained was as shown in Fig. 2.3. This

was used as a template for further weight

reduction in the knuckle. The reduced knuckle

was iterated till the material reduction and

strength obtained were optimal.

The final design is shown in Appendix 1 and

mentioned under the static structural analysis

earlier.

Fatigue analysis

The knuckle is one of the most important

suspension components and is the medium

that joins the wheel to the chassis. The

knuckle thus is thus subjected to constant

fluctuating loads and fatigue failure is an

important criteria. Fatigue is also responsible

for failure of 50-60 % of the components.

For fatigue analysis, the fatigue tool was

added 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 106 cycles)

The minimum life of the rear knuckle was

determined as 70500 cycles and the factor of

strength 0.39.

2.3 Bellcranks

Inboard suspension system in the front

facilitates the use of bellcranks to pivot the

spring and pushrod assembly. The entire road

forces are transferred to the spring through

the bellcranks, hence appropriate design of

the bellcranks is necessary for fatigue.

The aim of the analysis was to design a

functional bellcrank in the least amount of

weight possible, yet sturdy enough to support

the tire forces.

Material used is again Al6061 and the FOS

requirement is 2.


parameters

Same as for the front and rear knuckles

2.3.2 Boundary conditions

The bellccrank has a central pivot with either

end supporting the spring one side and the

pushrod on the other.

The bellcrank bolts will be loaded in shear as

the pushrod actuates the spring and

experiences the opposite reaction.

Fixed support the inner race of the central

pivot.

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

has a higher FOS than required, even after

high weight reduction of the cut-outs through

a process identical to the knuckle iterative

loop.

2.4 Wishbones or A-arms

The wishbones or the A-arms of a typical

double wishbone suspension have been

analyzed for strength. The results for the front

lower wishbone have been showed as an

example to represent the design process of

the suspension A-arms.

The factor of safety for the A-arms is chosen

to be 1.5 and the material as AISI 4130 steel.




Syt = 360 MPa

Sut = 560 MPa


parameters

The Pipe18 element was chosen for the

representation of the A-arms. This facilitated

the construction of a simple line figure in

ANSYS, hence allowing for a flexible design

which could be reiterated or changed easily.

Also, this would reduce the computational

time and endow a simple geometry. Meshing

size was chosen as 1mm uniform, which gave

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

2.4.2. Boundary conditions

To find the optimal loading condition for the

front geometry

Total Weight of vehicle + human = 3000 N

Number of Suspension Arms = 4 *2 = 8

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

Fx = 0.3 * 600

Fx = 180 N

In addition to this, the pushrod mounting on

the A-arm will experience a maximum force

equivalent to the weight of the front end,

about 800 N.

Hence, the loading and boundary conditions

are,

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

Fz = 800 N at the pushrod mount

Fixed support at the rod end bearing and

chassis mounts.

2.4.3 Simulation results

The contours have been shown in Fig 1.13.

The FOS was iterated by changing the cross-

section of the Pipe18 elements. An acceptable

design is obtained with a hollow circular pipe

of dimensions 20 X 2 mm.

2.4.4 Conclusions

Weight of front/rear knuckle = 1.023/1.25 kg

Weight of bellcrank = 0.77 kg

The wishbones were also chosen of optimal

cross-section. Hence, targets were achieved.




2.6 Additional CAE for suspension and

dynamic analysis of the vehicle

MSC Adams

MSC Adams is a mechanical systems analysis

software which enabled us to study the

dynamics of our sub-systems, the interactions

of various sub-systems and thus optimize

their design and performance. It helped us to

eliminate the need to actually build and test

our designs.

2.6.1 Front Suspension and Steering

MSC Adams was used to create a template of

the suspension and steering system of the

car into a front assembly. This was done by

modifying the hardpoints of an available FSAE

template for inboard suspension.

The resulting assembly was analyzed for

parallel wheel travel and toe change was

measured against wheel travel. The height of

the steering rack was iterated for minimum

toe change during the suspension travel.

Hence Bump Steer was eliminated. An

optimized graph for bum steer is shown in Fig

2.5.

2.6.2 Full-vehicle assembly

We have presently constructed a full-vehicle

assembly in MSC Adams incorporating the

suspension, steering, chassis, brake,

powertrain and engine into a single assembly.

Results with this assembly for dynamic driving

conditions are being pursued.

MATLAB

MATLAB is a powerful mathematical tool with

multiple applications and a user-friendly

interface. We used MATLAB to simulate the

longitudinal vehicle dynamics of the entire

vehicle by considering the vehicle to be a two

DOF system.


parameters

The entire vehicle was modeled as a 2 degree

of freedom (DOF) spring-damper system in

MATLAB. Differential equations for the model

were derived from first principles, and

modeled for the gross vehicle parameters.

2.6.4 Loading conditions and results

Sinusoidal excitation of 10mm amplitude and

20 rad/s frequency was given as input to

obtain the frequency of front and rear setups,

and the overall vehicle pitch and bounce for

the said excitation. The pitch and bounce

values originally received depended upon the

damper gain in the SimuLink model.

Increasing this gain, we were able to achieve

reduction in pitch from 0.28 deg to 0.20 deg

and the bounce from 15 mm to 10 mm.

The results have been shown in Appendix 2,

Fig 2.6.

2.6.5 Future work

Presently, we are working on a MATLAB

model for optimizing the lateral dynamics of

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

Intake Restrictor Analysis The intake Restrictor is to be fitted in the air

intake pipe in order to restrict the air flow

into the engine in order to limit the speeds

attainable by the engine. The commercially

available software Fluent was used for

simulating the flow through the nozzle.

Aerodynamics In the Octane Racing vehicle, we used a front

wing, a rear wing and a nose cone to

optimize downforce and drag across the

vehicle. We have used the software FOILSIM

available at the NASA website for selecting

an aerofoil that satisfies our downforce

requirements. The results have been validated

by referring a book The Theory of Wing

Sections by Ira Abott.

3.1 Problem Description Intake restrictor

The restrictor has to be fit in intake manifold

of the engine. So the diameter of manifold as

actually measured is 35mm. The maximum

diameter allowed in rulebook for restrictor is

20mm, that is going to be the throat

diameter.

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

Reduction in the throat area will only

serve to reduce the mass flow

The maximum possible area is defined

by the rulebook i.e. maximum

diameter of 20mm

Thus, the only design variable is to check the

restrictor geometry for flow separation in the

exit section i.e. the divergent portion of the

nozzle. That is done on ANSYS 12 FLUENT

software.

In FLUENT analysis, we used different outlet

pressure values and studied the flow pattern.

In analysis, the turbulence model used is k-

omega SST (shear stress transport) for low to

medium turbulence intensity. The flow

separation was not shown by any result.

Hence the restrictor design in free from

separation losses.

Aerodynamics

The downforce required in the Octane Racing

vehicle was approximated as one-third of the

downforce available on a Formula 1 vehicle.

The downforce generated by the rear wing of

a Formula1 race car is approximately 450 kg at

300 kmph. Considering the top speed of the

Octane racing car as 100 kmph, the

downforce required is approximately 170 N

(as downforce is directly proportional to

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

by IRA ABOTT for validation.

We decided not to use the front wing for

obtaining downforce as the requirement for

the Octane Racing vehicle is minimal. Hence,

we chose the NACA0008 profile for the front

wing; it only serves the purpose of supporting

the end plates.

3.2 Simulation results

The simulation images from ANSYS FLUENT 12

are shown in Fig. 1.14 and Fig. 1.15 in the

appendix.

3.3 Future work

We hope to get concrete results in the flow

simulation for the validation of the chosen

profiles for the rear and front wings. Also, we

propose to conduct simulations for nozzle

flow analysis using actual pressure values for

the restrictor.

FINAL CONCLUSION

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 for weight and stiffness,

validating the design of key structural

components like the knuckle, bellcrank, brake

assembly, etc. ANSYS Workbench provides a

simple interface which offers options for

online modification of the design and re-

evaluation.

ANSYS also offers various modules such as

static structural, transient structural, modal,

thermal, etc. which can be effectively used to

incorporate the requirements of the various

sub-systems in the present application.

Apart from ANSYS, ANSYS FLUENT has been

used to design and validate the intake

restrictor, side-pod and the wings. MSC

Adams has been used for dynamic simulation

of steering and suspension systems for the

elimination of bump steer. MATLAB and C++

have been used for the optimizing various

vehicle dynamics parameters, such as the

longitudinal, lateral and vertical behavior of

the vehicle.

REFERENCES

1. Supra SAE rulebook, 2011 (Version 2)

2. Theory of wing sections Ira Abott

3. Race Car Vehicle Dynamics Milliken and

Milliken

4. Octane Racing Preliminary Design Report,

Supra SAE 2011

5. Fundamentals of Vehicle Dynamics

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

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 1.6 Rear bump

Fig 1.4 Rollover impact

Fig 1.5 Front bump

12 | P a g e

Fig 1.8 Front knuckle deformation, stress, fatigue life

Fig 1.9 Rear knuckle

Fig 1.7 Modified rollcage




Page | 13

Fig 1.13 Front wishbones

Fig 1.10 Bellcrank

Fig 1.11 MSC Adams suspension, steering assembly Fig 1.12 MSC Adams full vehicle assembly




Page | 14

Fig 1.15 FLUENT simulations

Fig 1.14 NACA models - Ira Abott




Page | 15

APPENDIX 2 Tables and graphs

No Type of analysis Load value Boundary conditions

1 Frontal impact 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 Roll over impact 2G Base All DOF=0

5 Front wheel bump 1500N 1 front+2 rear wheels All DOF=0

6 Rear wheel bump 2500N 2 front+1 rear wheels All DOF=0

7 Torsional rigidity 1320N-m Rear roll hoop All DOF=0

Table 2.1 Rollcage boundary conditions

No Type of analysis

Displacem

ent (mm)

Stress

( MPa) FOS

1 Frontal impact 0.02 77.283 3.80

2 Rear impact 4.64 250.07 1.18

3 Side impact 0.53 325.57 0.90

4 Roll over impact 0.71 74.219 3.96

5 Front wheel bump 0.43 49.554 5.93

6 Rear wheel bump 27.01 401.46 0.73

7 Torsional rigidity 0.78 105.55 2.79

Table 2.2 Rollcage analysis results old design

Type of analysis Stress (old)

Stress (modified)

New FOS

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 | 16

Table 2.4 Result for suspension components

Fig. 2.5 Bump steer, MSC Adams

Component Deformation(mm) Stress (Mpa) FOS

Front knuckle 0.46855 180.87 1.53

Rear knuckle 0.06475 106.53 2.59

Bellcrank 0.01412 22.127 12.47

Fig. 2.6 MATLAB suspension, longitudinal dynamics

86415132 Ansys Cae Paper

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