10me209 - Design Data

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1 | P a g e D E S I G N L A B - 1 0 M E 2 0 9 . D A T A B A N K - S O U R C E O F D A T A .

SCHOOL OF MECHANICAL SCIENCES

DESIGN DATA BANK

10ME209 DESIGN LAB

CREDIT 0:0:1

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TABLE OF CONTENTS

Exercise

No.Name of the Exercise

Page

No.

1.DESIGN OF A CLOSED COIL HELICAL

COMPRESSION SPRING USING C/C++PROGRAM3

2.DESIGN OF A SINGLE DRY PLATE CLUTCH

USING C/C++PROGRAM5

3.DESIGN OF A BRAKE DRUM USING C/C++

PROGRAM9

4.DESIGN OF A JOURNAL BEARING USING C/C++

PROGRAM11

5.DESIGN OF A GEAR DRIVE USING C/C++

PROGRAM

14

6.DESIGN OF A HELICAL GEAR USING C/C++

PROGRAM19

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DESIGN OF A CLOSED COIL HELICAL COMPRISSION SPRING

PARAMETERS:

W

max

Maximum service load, N (Newton)

W

min

Minimum service load, N (Newton)

range

Axial deflection for the given range of load, mm (millimeters)

max

Maximum deflection of the spring, mm (millimeters)

C

Spring index ( ), dimensionless

max

Permissible shear stress intensity, N/mm

2

(Newton per millimeters squared)

G Modulus of Rigidity, N/mm

2


K

s

Wahl stress factor, dimensionless

W

range

The load for which the spring deflects to the given magnitude (W

max

W

min

), N (Newton)

d Diameter of the spring wire, mm (millimeters)

D

Mean diameter of the spring, mm (millimeters)

D

i

Inner diameter of the spring, mm (millimeters)

D

o

Outer diameter of the spring, mm (millimeters)

n Number of turns of the spring wire, turns

n

f

- Number of turns of the spring wire with squared and ground ends, turns

L

f

Free length of the spring, mm (millimeters)

p Pitch of the coils, mm (millimeters)

ALGORITHM:

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DESIGN OF A SINGLE PLATE FRICTION CLUTCH

PARAMETERS:

P Power to be transmitted, kW (kilo Watts)

N Speed of rotation of the clutch plate, rpm (revolutions per minute)

D

shaft

Diameter of the shaft onto which the clutch plate is coupled, mm (millimeters)

R Mean radius of the clutch plate * +, mm (millimeters)b

Face width of the clutch plate , mm (millimeters)r

Ratio between mean radius and face width respectively , dimensionlessR

i

- Inner radius of the clutch plate, mm (millimeters)

R

o

- Outer radius of the clutch plate, mm (millimeters)

n

springs

Number of helical compression springs, numbers

C Spring index ( ), dimensionlessT Torque, N-m (Newton meter)

shaft

Permissible shear stress intensity of the shaft, N/mm

2


spring

Permissible shear stress intensity of the spring, N/mm

2


p

disc

Permissible pressure intensity on the friction disc, N/mm

2


- Coefficient of friction of the friction disc, dimensionlessn

contact

Number of pairs of friction or contact surfaces, numbers

W

axial

Axial load acting on the faces of the friction disc (clutch plate), N (Newton)

W

total

Axial load acting on the faces of the friction disc (clutch plate) by considering 25 overload

for the transmission of maximum engine torque, N (Newton)

W

spring

Load acting on each spring, N (Newton)

max

Maximum deflection of the spring, mm (millimeters)

G Modulus of Rigidity, N/mm

2


K

s

Wahl stress factor, dimensionless

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W

range

The load for which the spring deflects to the given magnitude (W

max

W

min

), N (Newton)

d Diameter of the spring wire, mm (millimeters)

D Mean diameter of the spring, mm (millimeters)

D

i

Inner diameter of the spring, mm (millimeters)

D

o

Outer diameter of the spring, mm (millimeters)

n Number of turns of the spring wire, turns

n

f

- Number of turns of the spring wire with squared and ground ends, turns

L

f

Free length of the spring, mm (millimeters)

p

Pitch of the coils, mm (millimeters)

ALGORITHM:

Permissible shear stress intensity of the shaft,

shaft

= 40 N/mm

2

(assume)

Permissible pressure intensity on the friction disc, p

disc

= 0.07 N/mm

2

(assume)

Coefficient of friction of the friction disc = 0.25 (assume) Number of pairs of friction or contact surfaces

n

contact

= 2 (for single plate clutch)

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Number of turns of the spring wire, n = 4 (assume)

Modulus of Rigidity, G = 84000 N/mm

2

(assume)

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DESIGN OF A BRAKE DRUM

PARAMETERS:

D

rd

Diameter of the rope drum, mm (millimeters)

R

rd

Radius of the rope drum , mm (millimeters)D

bd

Diameter of the brake drum, mm (millimeters)

R

bd

Radius of the brake drum , mm (millimeters)n - Number of brake shoes, numbers

2ngle subtended by each of the brake shoes : (degrees)

m

Mass of the elevator when loaded, kg (kilograms)

v Linear velocity of the elevator when loaded, m/s (meters per second)

u

Initial velocity of the elevator when it starts from rest, m/s (meters per second)

a

Acceleration of the elevation, m/s

2

(meters per second squared)

h Linear distance moved by the elevator after the application of the brake, m (meters)

- Coefficient of friction between the brake drum and brake shoes, dimensionlessp

b

Allowable pressure on brake shoes, N/mm

2


Q Heat generated in stopping the elevator, J (Joules)

F

a

Accelerating Force, N (Newton)

F Total load acting on the rope while moving, N (Newton)

T Torque acting on the rope drum (shaft), N-mm (Newton millimeters)

F

tb

Tangential force acting on the brake drum, N (Newton)

F

t

Tangential force acting on each brake shoe, N (Newton)

R

N

Normal load on each shoe, N (Newton)

A

b

Projected area of the bearing on each brake shoe, mm

2

(millimeters squared)

w Width of the brake shoe, mm (millimeters)

g Acceleration due to gravity (9.80665 m/s

2

), m/s

2

(meters per second squared)

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ALGORITHM:

From the third equation of motion, we have

Total load acting on the rope while moving = Load on the elevator in Newton + Accelerating force

The brake drum is provided with four number of cast iron shoes, therefore the tangential force acting

on each shoe is,

If the angle of contact () of each shoe is 45: the equivalent coefficient of friction () need not be

calculated.

We know that,

The heat generated in stopping the elevator = Total energy absorbed by the brake

= kinetic energy + potential energy

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DESIGN OF A JOURNAL BEARING

PARAMETERS:

W Load on the journal, N (Newton)

N Speed of rotation of the clutch plate, rpm (revolutions per minute)

Z Absolute viscosity of SAE 10 oil, kg/m-s (kilograms per meter second)

t

oilFilm

Oil film temperature i.e. the temperature at which viscosity of the oil is considered,:C (degree

Celsius)

t

ambient

mbient temperature :C (degree Celsius)

p

max

Maximum bearing pressure for pump, N/mm

2


t

oil

Maximum rise in temperature of oil,

:

C (degree Celsius)

h Heat dissipation coefficient, W/m

2

-:C (Watts per meter squared degree Celsius)

l Length of the journal bearing, mm (millimeters)

d

Diameter of the journal bearing, mm (millimeters)

p

Pressure acting on the journal bearing as per the load acting on it, N/mm

2

(Newton per

millimeters squared)

- Coefficient of friction between the journal and the journal bearing, dimensionless

v

Linear velocity of the journal, m/s (meters per second)

Q

generated

Heat generated per second, W (Watts)

Q

dissipated

Heat dissipated per second, W (Watts)

Q

removed

Amount of heat that must be removed per second, W (Watts)

- Mass flow rate of the oil, kg/s (kilograms per second)C

p

Specific heat capacity of the oil at constant pressure, J/kg-:C (Joules per kilogram degree Celsius)

l/d

Length to the diameter ratio of journal bearing, dimensionless

c/d Clearance to the diameter ratio of the journal bearing, dimensionless

K Bearing modulus (the point at which the coefficient of friction is minimum and at which the

bearing must not be operated because a slight decrease in speed or increase in pressure might break

the oil film and make the journal operate with metal to metal contact), dimensionless

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k

Factor to correct for end leakage. It depends upon the ratio of length to the diameter of the

bearing (l/d) (k = 0.002, for l/d ratios of 0.75 to 2.8), dimensionless

ALGORITHM:

Table: 1

Machinery Bearing Maximum

bearingpressure,

N/mm2

Operating values

Absolute

viscosity(Z), kg/m-s

Np Z, kg/m-s

p, N/mm2

cd ld

Generators,

motors ,

and

centrifugal

pumps

Rotor 0.7 1.4 0.025 28 0.0013 1 - 2

First of all let us find the length of the journal bearing (l). Assume that the diameter of the journal

bearing (d) as 100 mm. From Table: 1 we find that the ratio of l/d for centrifugal pumps is in the

range between 1 and 2. Let us take l/d = 1.6

If the pressure induced in the journal bearing by the load acting on the journal is less than the

maximum bearing pressure of the pump, then the design is safe.

In other words, if p < p

max

, then the design is safe.

Otherwise change the dimensions appropriately.

From Table: 1 we find that the operating value of the bearing characteristic number,

The value of bearing characteristic number at which the friction is at its minimum is called thebearing modulus and is denoted by the letter K The bearing should not be operated at this value ofbearing modulus because it might break the oil film and cause metal to metal contact to occur

between the journal and the journal bearing, therefore the bearing should be designed for a value of at least three times the minimum value of bearing modulus.

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Substitute the

value obtained

from Table: 1

The

computed

value

The computed

value

, assumedfor convenience ( )

* +

9 (Assume)

84

Therefore the minimum value of bearing modulus at which the oil film will break can be represented

mathematically as,

Check whether From Table: 1, we find that the clearance ratio for centrifugal pumps is 0.0013.

( ) * +

If , cooling of the journal bearing must be facilitated artificially.

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DESIGN OF A GEAR DRIVE

PARAMETERS:

P Maximum power that can be transmitted, W (Watts)

i The gear ratio or velocity ratio , dimensionlessNPSpeed of the pinion, rpm (revolutions per minute)

NGSpeed of the gear, rpm (revolutions per minute)

DGPitch diameter of the gear, mm (millimeters)

DPPitch diameter of the pinion, mm (millimeters)

TGNumber of teeth on gear, numbers

TPNumber of teeth on pinion, numbers

L the center distance between the shafts, mm (millimeters)

og= opstatic stress of the material of which the gear and pinion is made of , N/mm2


ypLewis form factor or tooth form factor for pinion, dimensionless

b Face width of gear or pinion, mm (millimeters)

m Module of the gear or pinion* +, mm (millimeters)C Deformation or dynamic factor in Buckingham equation (a factor depending on

machining error)

K Material combination factor for wear

CvVelocity factor, dimensionless

v Linear velocity of the gears or pinions pitch circle, m/s (meters per second)

WTTangential tooth load (beam strength of the tooth), N (Newton)

WDDynamic load on tooth, N (Newton)

WIIncrement load due to machining error and the pitch line velocity , N(Newton)

WsStatic tooth load or endurance strength of the tooth, N (Newton)

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Q Ratio factor eFlexural endurance limit, N/mm2(Newton per millimeters squared)

wp Permissible working stress for pinion

[ ], N/mm2


pcpCircular pitch for pinion , mm (millimeters)WwMaximum or limiting load for wear, N (Newton)

r Ratio between face width and module * +(assumed for convenience),dimensionless

ALGORITHM:

Since both the gear and pinion will be made of the same material, the pinion that will besubject to much wear, tear, and rotations. Thus the design will be based on pinion.

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44 5 6

6

7575 75 5

4 684 4 54 9 75 84

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Table: 1 (Values of service factor (Cs))

Type of load Type of service

Intermittent or 3

hours per day

8-10 hours per day Continuous 24 hours

per day

steady 0.80 1.00 1.25

Light shock 1.00 1.25 1.54

Medium shock 1.25 1.54 1.80

Heavy shock 1.54 1.80 2.00

Assuming steady load conditions, and 8-10 hours of service per day, the service factor (Cs) as

given in Table: 1 be taken as 1.00.

( )

Solve the equation obtained which of the third order, and compare the value of the module

(m) against the standard values given in Table: 2, and then select the appropriate one.

Table: 2 (Standard module values in mm (millimeters))

Preferred values of module

(mm)

Option 1

(mm)

Option 2

(mm)

1

1.25 1.125

1.5 1.375

2 1.75

2.5 2.253 2.75 3.25

4 3.5

5 4.5 3.75

6 5.5

8 7 6.5

10 9

12 11

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16 14

20 18

Note: Choose the values of modules form the column that specifies the preferred values

A check for dynamic and wear load.

Table: 3

Material of pinion and gear Brinell Hardness Number

(B.H.N)

Flexural endurance limit (e) in

Mpa (N/mm2)

Grey cast iron 160 84

If, Ws>WD, and Ww>WDthen the design is safe.

Substitute the value

from Table: 3

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DESIGN OF HELICAL GEAR

PARAMETERS:

- Helix angle of the cast steel gear, :(degrees)

P Power that has to be transmitted, W (Watts)

N

G

of the gear, rpm (revolutions per minute)

T

G

-Number of teeth on gear, numbers

T

E

Equivalent number of teeth on gear* +, numbersDGPitch diameter of the gear, mm (millimeters)

b Face width of gear, mm (millimeters)

ostatic stress of the material of which the gear is made of , N/mm2(Newton per

millimeters squared)

pNNormal pitch of the gear, mm (millimeters)

r Ratio between face width and normal pitch * +(assumed forconvenience), dimensionless

y Lewis form factor or tooth form factor for gear, dimensionless

T Torque transmitted through the gear, N-mm (Newton millimeter)

v Linear velocity of the gears pitch circle, m/s (meters per second)

WTTangential tooth load (beam strength of the tooth), N (Newton)

CvVelocity factor, dimensionless

w Permissible working stress for gear , N/mm2(Newtonper millimeters squared)

pcCircular pitch for gear

, mm (millimeters)

pNNormal pitch of the gear , mm (millimeters)m Module of the gear , mm (millimeters)WATangential tooth load (beam strength of the tooth), N (Newton)

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4 684 4 54 9 75 84

66 5 5 5 7575 75 5

ALGORITHM:

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Assume that the peripheral velocity will be between 10 and 20 m/s.

Solve the equation obtained which of the third order, and compare the value of the module

(m) against the standard values given in Table: 1, and then select the appropriate one.

Table: 1 (Standard module values in mm (millimeters))

Preferred values of module(mm) Option 1(mm) Option 2(mm)

1

1.25 1.125

1.5 1.375

2 1.75

2.5 2.25

3 2.75 3.25

4 3.5

5 4.5 3.75

6 5.5

8 7 6.510 9

12 11

16 14

20 18

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10me209 - Design Data