June 12 Th Presentation

8/13/2019 June 12 Th Presentation

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DESIGN & DEVELOPMENT

OF

7.5 K.W STARTER MOTOR

CLIENT : LUCAS – TVS

AU-FRG INSTITUTE FOR CAD/CAM


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To enter into world market

To design a starter motor for Earth movers / heavyequipment

Compact in size (outer dia. < 114 mm)

Need and Objectives

Low speed ,high torque electric motor is bulky

Process – sintering ,extrusion, rolling

Combined use of existing imported parts

Challenges


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High torquelow speed

compactness

Low torque

High speed

compactness

Gear

reduction

Gear reduction starter motor assembly


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Gear reduction

? Simple gear train

? Planetary gear train

Compactness

Low backlash

High Torque/weight(Volume)

Input * Output shaft are coaxial


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Detailed design and Validation

Kinematic

arrangement of

Planetary

Gear Design

Profile shift&

Backlash Pin,output shaft

Design…

Casing design

3D modeling

FE Analysis


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KINEMATIC ARRANGEMENT The Design Process … N1 = 3060N2 = N1/i

Zs = Ds/m

Za = ((N1 - N2)/N2)*ZsZp = (Za - Zs)/2

Ns = N1Nc = N2Np = N2 -((Zs/Zp)*(N1-N2))

Na = N2 -((Zs/Za)*(N1-N2))

Ts = (P*60*7)/(2*22*N1)

Tc = -Ts*(Ns/Nc)

Ta = -(Ts+Tc)

N1 = 3060N2 = N1/i

Za = Da/m

Zs = (N2 * Za)/ (N1 – N2)

Zp = (Za - Zs)/2

Ns = N1

Nc = N2Np = N2 -((Zs/Zp)*(N1-N2))

Na = N2 -((Zs/Za)*(N1-N2))

Ts = (P*60*7)/(2*22*N1)

Tc = -Ts*(Ns/Nc)

Ta = -(Ts+Tc)

N1 = Input Speed N2 = Output Speed Zs = Sun Teeth Zp = Planet Teeth

Za = Annulus Teeth Ts = Sun Torque Tc = Carrier Torque Ta = Annulus Torque


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KINEMATIC ARRANGEMENT OF

GEARS

Determination of Number of teeth on Sun, Planet &

Annulus based on various conditions,

Various Conditions considered for this arrangement are1. Diameter of the Annulus Gear (Ds)

[ 90, 94 ]

2. Module (m)

[ 1, 1.25, 1.5, 2 ]

3. Reduction Ratio ( i )

[ 3, 3.25, 3.5 ]


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Kinematic arrangement of Planetary

Determination of Number of teeth on Sun(Zs), Planet(Z

p) &

Annulus(Za)

Various Conditions considered

1. PCD of the Sun Gear (Ds)

[ 31, 32, 33, 34, 35 ]2. Module (m)

[ 1.5, 1.75 ]

3. Speed Reduction Ratio ( i )

[ 3 - 3.6 ]

Checking for Planetary Assembly

( Zs + Za ) / No of Planets = an integer

Combinations not satisfying this condition were

ruled out.


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Sun PCD 31mm Sun PCD 32mm

Sun PCD 33mm Sun PCD 34mm

Various combinations obtained from “C” program


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Sun PCD 35mm

Output for gear design from Kinematic arrangement

No of teeth in Sun gear, Planet gear, Annulus gear

(Zs, Zp & Za)

These 66combinations

considered for

Gear design.


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Kinematic

arrangement of

Planetary

Gear Design

Profile shift


Design…

Casing design

3D modeling

FE Analysis


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GEAR DESIGN

For each Set of Kinematic Arrangement the following

Gear Parameters have been determined,

1. Module

2. Centre Distance

3. Face Width

4. Pitch Diameter

Considering the materials mentioned below,

1. EN 8 ( C 45 )

2. EN 325 ( 15 Ni 2Cr 1 Mo15 )

3. EN 24 ( 40Ni 2Cr 1 Mo28 )


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VARIOUS COMBINATIONS USED IN DESIGN OF

KINEMATIC ARRANGEMENT

SPEED

RATIO

(I)

MOD - 1.25 MOD - 1.5 MOD - 1.75 MOD – 2

Zs Zp Za Zs Zp Za Zs Zp Za Zs Zp Za

3 36 18 72 30 15 60 25 13 51 23 11 45

3.25 32 20 72 28 16 60 23 14 51 21 12 45

3.5 28 22 72 24 18 60 21 15 51 27 14 45

DIAMETER OF ANNULUS GEAR (Ds) 90 mm

Combinations considered forNo interference criteria Combinations failed for Assembly Condition


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The materials considered for Gear Design are

1. EN 8 ( C 45 )

2. EN 325

3. EN 24

4. EN353

Gear Design

Center distance

Module

Based on Torque & Compressive strength

Based on Torque & Bending strength


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Checking for Plastic deformation

2.5 times the lock Torque.

Gear parameters

Module

PCD

Face width

Center Distance for Sun, Planet & Annulus


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Only EN 353 satisfies the requirement of

2.5 * Lock torque transmission

small module value (1 to 2)

without bending & compressive failures

PCD= m Z

45 combinations finalized out of 66

combinations based uponsmall module

reduction ratio 3 to 3.5


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45 combinations narrowed down to16 combinations based upon

minimum no teeth > 12

preferred module : 1.5 ,1.75

reduction ratio around 3.3


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Finally one combination is

finalized by Client as the speedreduction ratio & no. of planetswell suited their requirements

Zs= 18 Z p = 12 Za =42

module 1.75


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Kinematic

arrangement of

Planetary

Gear Design

Profile shift&


Design…

Casing design

3D modeling

FE Analysis


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Gear correction factors were

determined including the backlashfor sun and planet then to Annulus

Final Gear parameters

No of teeth : Zs, Z p & Za

Module

Correction factor

PCD

Tip circle diameter

Root circle diameter

Profile shift & Backlash


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Kinematic

arrangement of

Planetary

Gear Design

Profile shift&

Backlash Pin,Output shaft

Design…

Casing design

3D modeling

FE Analysis


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Parametric 3D modeling

Automation

Gear Assy. Model gets

input parameters from

external program


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Carrier Pin Design

Output shaft design

Bearing design

Existing components

Std. parts…


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Validation & optimization

with FEA


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D-Bracket

End Bracket

Knowledge from

existing casings

Validation & optimization

with FEA


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Finalized design


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PROJECT DEFINITION

Design & Development of a Starter Motor of 7.5k.W power for the conditions below,

1. No Load Condition – 3200 RPM – 0.66 kgf-m

2. Loaded Condition - 1020 RPM – 8.55 kgf-m

3. Locking Condition - Zero RPM – 23.5 kgf-m


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Plan of the project proceedings is as follows,

DESIGN MODELING ANALYSISKINEMATIC ARRANGEMENT

GEAR DESIGN

BACKLASH

PIN DESIGN

CARRIER PLATE DESIGN

BEARING DESIGN

OUTPUT SHAFT DESIGN

CORRECTION ( PROFILE SHIFT )

ASSEMBLY

CLUTCH ASSEMBLY

SUN, PLANET &

ANNULUS

END BRACKET

CARRIER PLATE,

PIN & SHAFT

D-BRACKETOUTPUT SHAFT

END BRACKET

CARRIER PLATE &

PIN

ANNULUS ( Lugs)

CORRECTIONS ( P fil Shift)


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CORRECTIONS ( Profile Shift)

Corrections for the Gears have been arrived atkeeping the Centre Distance of Annulus and Sun same

Correction factors have been given for the followingtwo cases,

* High Contact Ratio

* High Load carrying Capacity

All the Gears are S+ and S- Gears

CORRECTION FACTORS BASED ON HIGH CONTACT RATIO


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S.

N

o

Zs Zp Za M Np Xs Xp Xa NEW CENTRE

DISTANCE Ds Dp Da Ta

1 21 13 47 1.75 4 0.18 0.13 0.42 30.248 36.8 22.8 82.3 88.1

2 20 14 48 1.75 4 0.17 0.13 0.43 30.244 35 24.5 84 89.9

3 22 14 50 1.75 4 0.19 0.15 0.49 32.06 38.5 24.5 87.5 93.6

4 22 12 46 1.75 4 0.16 0.15 0.46 30.263 38.5 21 80.5 86.5

5 24 14 52 1.5 4 0.16 0.14 0.44 28.929 36 21 78 83.1

6 23 15 53 1.5 4 0.15 0.15 0.45 28.929 34.5 22.5 79.5 84.6

7 23 17 57 1.5 4 0.12 0.1 0.32 30.316 34.5 25.5 85.5 90.2

8 18 12 42 1.75 5,6 0.22 0.18 0.58 26.894 31.5 21 73.5 79.9

9 22 13 48 1.5 5 0.17 0.14 0.45 26.689 33 19.5 72 77.1

10 23 12 47 1.75 5 0.18 0.13 0.45 26.69 33 19.5 72 77.1

11 24 16 56 1.5 5 0.12 0.1 0.32 30.319 36 24 84 88.7

12 20 15 50 1.5 5 0.15 0.16 0.47 26.689 30 22.5 75 80.2

13 21 15 51 1.5 6 0.15 0.13 0.41 27.3996 31.5 22.5 76.5 81.5

14 22 14 50 1.5 6 0.16 0.12 0.4 27.3996 33 21 75 79.9

CORRECTION FACTORS BASED ON HIGH CONTACT RATIO


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S.

N

o

Zs Zp Za M Np Xs Xp Xa NEW CENTRE

DISTANCE Ds Dp Da Ta

1 21 13 47 1.75 4 0.5 0.43 1.3589 31.159 36.8 22.8 82.3 91.38

2 20 14 48 1.75 4 0.49 0.44 1.36896 31.1608 35 24.5 84 93.166

3 22 14 50 1.75 4 0.47 0.42 1.3046 32.8576 38.5 24.5 87.5 96.44

4 22 12 46 1.75 4 0.51 0.41 1.3291 31.1474 38.5 21 80.5 89.526

5 24 14 52 1.5 4 0.5 0.42 1.3397 29.7125 36 21 78 85.769

6 23 15 53 1.5 4 0.49 0.43 1.3462 29.708 34.5 22.5 79.5 87.2888

7 23 17 57 1.5 4 0.47 0.42 1.3084 31.1811 34.5 25.5 85.5 93.175

8 18 12 42 1.75 5,6 0.48 0.43 1.3387 27.6145 31.5 21 73.5 82.56

9 22 13 48 1.5 5 0.46 0.41 1.2798 27.3936 33 19.5 72 79.589

10 23 12 47 1.75 5 0.47 0.4 1.2698 31.9592 33 19.5 72 91.0695

11 24 16 56 1.5 5 0.48 0.41 1.2984 31.1811 36 24 84 91.645

12 20 15 50 1.5 5 0.44 0.43 1.2998 27.3936 30 22.5 75 82.6495

13 21 15 51 1.5 6 0.45 0.43 1.308 28.1566 31.5 22.5 76.5 84.1741

14 22 14 50 1.5 6 0.46 0.41 1.2747 28.1448 33 21 75 82.5742

CORRECTION FACTORS BASED ON HIGH LOAD CARRYING CAPACITY


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PIN DESIGN

Pin Designed using the following equations,

Induced Bending stress = 32*M / (pi*d3

)

Induced Bearing stress = 2 * Ft / d*b

Total Length of Pin = b + (0.625*dpin)



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Carrier Plate Diameter = 2a’ + 2dpin

where

a’ = Working Centre Distance dpin = Diameter of Pin

Thickness (Tc) = 0.625 * dpin


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SHAFT DESIGN

Output Shaft is designed based onTorque Transmitted & Bending Moment due to Gear loads

Data Required from client side,

Pinion Detail - Weight, Position, Bore Dia,Correction Factor

Clutch Details

Traveling Distance

Any other Constraints


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BACKLASH

Arc Tooth Thickness at PCD for Sun

S1 = ( ( (pi*m)/2 ) + ( 2*Xs*m*tan(Alpha) ) );

Arc Tooth Thickness at PCD for Planet

S2 = ( ( (pi*m)/2 ) + ( 2*Xp*m*tan(Alpha) ) );

Arc Tooth Thickness at WCD for Sun

S1DASH = ( ((Dw1*S1)/D1) - (Dw1*(IWpa - .0149)) );

Arc Tooth Thickness at WCD for Planet

S2DASH = ( ((Dw2*S2)/D2) - (Dw2*(IWpa - .0149)) );

New Correction for Annulus after giving Backlash isXa = ( 1/(2*m*.36397) ) * ( (D3/Dw3)*(S2DASH+BACK+( Dw3*(IWpa - .0149)) - ((pi*m)/2 ) ) );


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5 PLANETS ASSEMBLY

A A


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5 PLANETS ASSEMBLY

MOUNTING BRACKET


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MOUNTING BRACKET

PLANETARY ASSEMBLY


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PLANETARY ASSEMBLY

PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY

FOUR PLANETS ASSEMBLY


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FOUR PLANETS ASSEMBLY


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PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY

PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY

PLANETERY ASSEMBLY


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PLANETERY ASSEMBLY

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June 12 Th Presentation

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