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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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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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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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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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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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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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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)
CARRIER PLATE DESIGN
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CARRIER PLATE DESIGN
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