4041 final presentation

Georgia Institute of TechnologyME 4041

Selectable Locking Differential

Dillon VoegeliWeeriya Meesook


Presentation Outline (1 Slide)

• Introduction and Objectives• Modeling• Assembly• FEA• Hand Validation• Conclusion


Introduction and Objectives

• Almost every vehicle sold today has an open differential design

• Selectable locking differentials exist, however they are very expensive– Expensive kit– Expensive professional installation


Introduction and Objectives

• Our group designed a selectable locking differential that can be installed by the average user– Installs without modifying crucial internal

components


• The carrier was the most complex component modelled for this project

• The general shape was created from a revolved profile, and multiple subtractions, fillets, and chamfers were used to model it as accurately as possible based on dimensions collected by hand

Modeling


• The basic shape of this 36-tooth gear was created using SolidWorks’ Gear Studio

• This was further modified in NX using subtractions to generate the desired geometry

Modeling


• The side gears were created using SolidWorks’ Gear Studio.

• Our calculations showed that these needed to be 16 tooth to handle the desired loads

• This constraint meant multiple redesigns to fit as many load-bearing gears as possible into the tight space of the carrier

Modeling


Modeling• The gear carriers were

relatively simple models, extrusions and subtractions

• Initially modeled as one component, split into two identical halves to facilitate installation in the limited space inside the carrier


Modeling• The center lock was created

using extrusions and subtractions

• Initially modeled as one component, split into the lock and shoe halves to facilitate installation in the confined space of the carrier


Assembly• Final assembly of our

design• 2 subassemblies• 13 parts total

• Two spur gear subassemblies constrained by concentric and touch/align mates


Assembly

• Planetary Gear Subassembly– Two identical assemblies– 1 Spur gear carrier, 3 spur gears

• Spur gears mated to carrier with concentric mates

• Subassemblies mated together with concentric and touch/align mates


Assembly

• Pin constrained by concentric and distance mates

• Large side gears constrained by touch/align and distance mates


FEA• All FEA analysis was performed as 3D static structural analysis under

conditions that would result in the failure of other critical components– Drive axles fail at 4200 ft-lb, want our design to survive this

• Analysis centered around material choice and cost vs performance optimization

– 1040 High-Carbon Steel, Yield Strength 374 MPa– 4140 Chrome-Moly Steel, Yield Strength 415 MPa– 9310 High-Strength Stainless Steel, Yield Strength 986 MPa

• Data on these showed that all have elastic and shear moduli of around 200 and 80 GPa respectively, so this made analysis easier as results could be compared to yield strength of all three options


FEA• Inside of gear constrained with a no

translation constraint

• Loading represents forces from three side gears

– 2 close together on one side and one opposite them

• Loaded to force at which the axle will fail

– 4200 ft-lb = 700 ft-lb per tooth– Want to ensure our design isn’t the weak link

in the drive train


FEA • Mesh convergence yielded a maximum stress

of 388 MPa at a mesh size of 1.5mm

• Max deflection of 0.0126 mm

• Originally intended to use AISI 4140 steel for this part, with a yield strength 415 MPa

– This analysis demonstrated the potential need to use a higher strength steel for improved safety factor

• While this component did not fail (yield), it highlighted the battle between cost and safety

– Higher strength alloys significantly more expensive than 4140 steel.


FEA• Inside of gear constrained with a no

translation constraint

• Loading represents forces from six locking teeth

– Arranged 60 degrees apart and normal to the surface of each tooth

• No translation constraints on outer teeth

• Loaded to force at which the axle will fail

– 4200 ft-lb = 700 ft-lb per tooth


FEA• Mesh convergence yielded a maximum stress

of 429.21 MPa at a mesh size of 2.5 mm

• Max deflection of 4.505 x 10-3 mm

• This component failed based on our original material selection and design parameters, and highlighted the need for stronger material choice or redesign to improve the safety factor of our design.


FEA• Mesh convergence yielded a maximum stress

of 1334.47 MPa at a mesh size of 1.5 mm

• This highlights an obvious weak link not initially captured in our hand calculations, and is the result of a design change

• This analysis highlighted the need of a design change going forward


Hand calculation Forces on side gear : where T is torque, D is diameter, and Wt is the tangential load

D = 36 teeth / 20 teeth/inch = 1.8 inT = 4500 lb-ft from initial conditions


Saftey Factor of gears Stress = (Wt*P)/(3*F*Y)/1000 where P is dimetral pitch, F is face width of tooth, and Y is the Lewis form factor

Stress = (60 ksi * 20)/(3*0.8*0.377)/1000Stress = 132 ksi

S = 162 / 132 = 1.23

Hand calculation


Animations (1 slide)


Concluding Remarks

• This project highlighted design issues that would need to be dealt with in order to improve the safety factor and overall robustness of our product

• It was also a valuable exposure to FEA analysis and its strengths and limitations, something that will prove invaluable in future applications.

4041 final presentation

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