design, development and evaluation of dual belleville...

International Journal of Mechanical & Mechatronics Engineering IJMME-IJENS Vol:15 No:01 79

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152301-8484-IJMME-IJENS © February 2015 IJENS I J E N S

Design, Development and Evaluation of Dual

Belleville Clutch in Front End Loader Tractor for

Clutch Life Enhancement

1Dr.P.S.Senthil Kumar., Professor, Department of Mechanical Engineering.,VEL TECH UNIVERSITY, Avadi, Chennai-

600062,Tamil nadu,India

Email: [email protected] 2P.Srinivasan,Member R&D.,Tractor & Farm Equipments Limited,TAFE.,Sembium, Chennai,Tamil Nadu,India.

Email:[email protected] 3Kishore Kumar Palati., Assistant Professor.,. Department of Mechanical Engineering.,VEL TECH UNIVERSITY,

Avadi,Chennai-600 062,Tamil nadu,India

Email: [email protected] 4Gandhi Mallela., Assistant Professor., Department of Mechanical Engineering. VEL TECH UNIVERSITY, Avadi,Chennai-

600 062,Tamil nadu,India

Email: [email protected]

Abstract-- In tractors, during front end loader, dozer, dry and

wet puddling applications there will be higher frequency of

operation of clutch because of increased need of quick forward

and reverse transmissions. This leads to the slippage of dry clutch

because of the usage of coil type clutch where there is reduction

in clamp load within minimum hours of time because of inability

of the coil springs to withstand high temperature. This would

result in excessive scoring marks in pressure plate and flywheel,

and glazing of friction material, effecting in slip of the clutch

leading to replacement of dry clutch assembly as a whole. In this

experimental work, it is proposed to modify the design and

replace the coil springs used in the dry clutch with belleville

spring which has the ability to withstand high temperature

leading to enhanced clutch life in high torque applications of the

tractor. This project comprises of calculations and analysis made

in support to the usage of belleville spring in dry clutch for front

end loader and other high torque applications. The result is a

theoretical comparison of existing coil springs with the belleville

spring in the clutch to prove the enhanced heat carrying capacity

of the belleville spring and thus resulting in improved reliability

of dry clutch in tractors.

Index Term-- Belleville clutch, operating temperature,

slippage, excessive scoring, slip

1. INTRODUCTION

Agricultural tractor attached with the front end loader

is commonly used in construction site because of its cheaper

price as compared to earth moving machine. The tractor used

for this application need to move forward and backward

repeatedly because of this clutch need to operate very

frequently which lead to increase the temperature of clutch. If

during single clutch engagements or in many repeated

applications the fiction surface temperature generated

becomes coefficient of fiction between the mating surfaces

fluctuates appreciably (generally µ decreases with increase in

temperature); steel contacting bodies distort and develop

surface cracks, metallic plates weld together and non-metallic

clutch plates wear at a greater rate than at lower temperatures.

The most failures and damages in the friction clutch occur due

to excessive frictional heat and heat fluctuations. These

situations lead to generate high thermal stresses, which causes

cracks and deformation for the friction material of clutch.

Finally, these disadvantages lead to reduce the lifecycle of the

friction material. After repeated clutch engagement the

temperature on the clutch facing could attain very high values,

around 300-350°C and, above 350-400°C the friction system

starts to suffer permanent damage.

In Heavy Duty front end loader application the

frequency of operation even reach to 7 to 8 application of

clutch per minute. As a result the temperature of the clutch

assembly rises beyond their temperature limit thereby

decreasing torque transmitting capacity of the clutch due to

fading or glazing of clutch friction material. In normal

application of tractor, the expected life of clutch is 500 h of

operation but when the tractor is used for this application the

life of clutch is reduced to 300 h of operation. The coil starts

to lose its load carrying capacity because of increased heat at

the clutch. In other words, the clamping load reduces for the

clutch, resulting in clutch slippage at frequent intervals. In

order to compensate the above loss, it is proposed to replace

the coils in the dry clutch with modification in belleville

spring to avoid failure.

2. LITERATURE REVIEW

2.1 H.K.Dubey and Dr. D.V. Bhope (2012), Many

researchers have carried out stress and deflection analysis of a

Belleville spring. the stress and deflection analysis to prepare

a CAD method for the checkout and design of the Belleville

springs. The method eliminates the need to resort to

conventional trial-and-error techniques. In a matter of seconds,

it rapidly and accurately checks out and designs Belleville

springs, outputting the load deflection characteristics in

graphic and table formats and can generate a dimensioned

drawing. the stress and deflection analysis of a slotted


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Belleville spring to develop a analytical relationship for

deflection and stress of a slotted conical spring.

2.2 Abdullah and Schlattmann (2012) developed a

two dimensional model to obtain the numerical simulation for

band contact of disc clutch during slipping. In this study, three

type of pressure application was used viz. constant pressure,

linearly increasing pressure and parabolically increasing

pressure. Finite element method was used to calculate the heat

generated on the surfaces of friction clutch and temperature

distribution for case of bands contact between flywheel and

clutch disc, and between the clutch disc and pressure plate.

They found that both slipping time and contact area ratio are

intensely effect disc clutch temperature fields in the domain of

time. Temperature distribution for constant pressure type is

higher than the other types of pressure, because of the total

quantity of thermal load is applied in short time with

compared to other types of thermal loads. In case of repeated

engagement, the linearly increasing pressure developed

maximum temperature. The damaged or incorrectly machined

flywheel causes many of problems one of them is focusing the

pressure on small regions of nominal frictional interface (e.g.

bands and spots).

2.3 Cappetti, et. al. (2012) observed the influence of

temperature on the cushion spring behavior in terms of the its

load-deflection curve and in terms of the axial thermal

expansion. They developed FE model for cushion disk and it

was validated by measuring actual load deflection of the

cushion spring at temperature of 20°C. Afterward, FE model

was set to different temperature and observed the change in

load-deflection curve as an effect of temperature. In their

study, they had found that at higher temperatures the curves

start deflecting before of the reference curve at 20°C start

deflecting. This effect is due to thermal expansion which

produces axial size increase and consequently a change of the

kiss point position.

2.4 Yevtushenko e al (1999) applied one-

dimensional transient heat conductivity to study the contact

problem of a sliding of two semi-spaces, which induces effects

of friction, heat generation and water during braking. In the

present temperature analysis the capacity of the frictional

source on the contact plane dependent on the time of braking.

The problem solved exactly using the Laplace transform

technique. Numerical results for the temperature are obtained

for the different values of the input parameter, which

characterize the duration of the increase of the contact

pressure during braking from zero to the maximum value. An

analytical formula for the abrasive wear of the contact plane is

obtained in the assumption, that the wear coefficient is the

linear function of the contact temperature

3. PROBLEM DESCRIPTION

The major cause of clutch failure can be

summarized with two words, “EXCESSIVE HEAT”

extreme operating temperatures (excessive heat) can cause the

clutch to fail because the heat generated between the

flywheels. Driven discs, coil spring and pressure plate are high

enough to cause the metal to flow and the friction material to

be destroyed. During high torque applications like front end

loader, dozer, dry and wet puddling, there will be higher

frequency of operation of the clutch as forward and reverse

gears will be used regularly and the application would

continue for around eight hours together. Since the clutch

engagement frequency is around seven to eight times per

minute and it is persistent for very long duration, during the

above mentioned applications, there will be increased heat

generation in the clutch.The coil starts to lose its load carrying

capacity because of increased heat at the clutch. In other

words, the clamping load reduces for the clutch, resulting in

clutch slippage at frequent intervals. In order to compensate

the above loss, it is proposed to replace the coils in the dry

clutch with belleville spring to avoid failure due to facing

wear.

Fig. 1. Front end loader

Fig. 2. Dry puddling

Fig. 3. Dozer


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Fig. 4. Wet puddling

Figure 1 details about the use of front end loader and

figure 2 describes the dry puddling application.In figure 3 the

dozer is used for special purpose in fields.Tractor is used for

wet puddling in figure 4.

4. PROPOSED WORK

In this project, the study and analysis is made on a 55kw

tractor subjected to front end loading and back hoe

applications. This duty cycles were chosen since the rate of

failure is higher with the existing dry clutch tractors (Coil

type).

The design approach adapted to this work is as

follows

Evaluation of different concepts to derive the optimal

design

Detailed design and analysis of finalized concept

Mathematical model to derive

Disc belleville load calculation.

Clutch clamp load Vs Travel

Clutch release load Vs Travel`

Torque capacity of clutch

Development and assembly of proto samples in the

base tractor, testing and experimental evaluation

Comparison of results from mathematical analysis

with the experimental measurement.

4.1 OBJECTIVE

Based on the above brief discussion, the following objectives

have been decided.

1. Study and Analysis of exiting clutch system

and parameter affecting clutch system.

2. Design and fabrication of belleville system

of clutch.

3. Performance evaluation of the Designed

clutch system

5. METHODOLOGY

5.1 DISC BELLEVILLE LOAD CALCULATIONS

The Nomenclature section describes the spring

parameters used in calculations. Load in lbs at a given

deflection and flat are the formulas. When calculating loads

take care to pay special attention to the four factors mentioned

below.

Nomenclature of belleville

5.2 LOAD FOR A GIVEN DEFLECTION.

In flattened condition, the deflection f is equal to the

conical height h and the equation becomes. If Pf

5.3 DYNAMIC LOADING AND FATIGUE LIFE

5.3.1 DYNAMIC LOADING

Dynamic loading of disc springs occurs when the

load continuously changes from preload to final load. The

"stress-time" curve of such disc springs which pulsate

uniformly is sinusoidal. This is not true in cases of impact

loading and therefore it is difficult to predict their life and

behavior.


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Disc spring "life" may be differentiated into 2

categories:

1. Limited life: where cycles vary without failure

between 40,000 and 2,000,00 cycles.

2. "Unlimited life": cycles in excess of 2·106 without

failure. For virtually indefinite life, the table below indicates

the appropriate values required given in percent of travel,

relating preload to final load and considering the disc spring

thickness Table I

Maximum deflection

Max Deflection

in % of h

Preload

in % of

h

Disc

Thickness <=.039

Disc

Thickness

b>=.157

15 50 44

25 56 49

50 67 64

5.3.2 FATIGUE LIFE

Fatigue life for a disc spring is defined by the

effective number of stress cycles that can be sustained prior to

failure under certain conditions. This depends on the minimum

stress, maximum stress and stress range.

The diagrams presented here are for evaluating

fatigue life of single disc springs or series stacks not more than

6 springs. There are three basic groups, depending on

thickness (see legend under each diagram).The horizontal

border line enclosing the top portion of the graph (zone)

represents the yield strength of the spring steel material.

Intersection points of min/max stress limits which fall outside

the graph/zone boundaries are to be avoided as they indicate

spring failure is likely at an early stage.

The graphs were developed based on empirical test

data. The test loads were sinusoidally executed.

Graph 1: Final stress Vs Preload stress



5.4 LOAD AND STRESS CALCULATIONS.


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5.2 CLUTCH CLAMP LOAD / RELEASE LOAD VS

TRAVEL

As well as enabling a compact design, transmission

disc springs offer an additional functional advantage. As

opposed to coil spring packs, which always have a linear load

curve, slotted disc springs can achieve a mainly horizontal

load curve if desired. The spring load can be kept constant

during the stroke ( working range ) , i.e. between E.H (

engaged height ) and spring height at closed clutch including

wear F.L.H. ( fully loaded height ). In spite of the somewhat

higher manufacturing tolerances required for disc springs (

standard wise +/- 10% compared to +/- 6% for coil springs ),

this results in a significantly lower load difference between the

two working points E.H and F.L.H . In addition, tighter load

tolerances can be kept for specific working points if desired.

Graph 4: Clamp load Vs Travel (Comparison of coil

Vs Diaphragm)

This leads to a big advantage with respect to gear

change, particularly with the control of the hydraulic system.

The spring load is constant, regardless of the actual position of

the piston in the engaging range, and the oil pressure can be

controlled within a much smaller range, as less hydraulic

pressure is required due to the lower load difference between

E.H and F.L.H. In this study, the hydraulic pressure required

for torque transmission could be reduced by 60Kpa, leading to

an improvement of more than 0.2% in total efficiency at the

F.L.H. point.

5.2.1 Radial Rigidity at High Revolutions

A further benefit offered by disc springs as opposed

to coil springs is the high shape retention and stiffness against

external forces. This is particularly noticeable at high clutch

revolutions, which lead to high radial centrifugal forces. The

single coil springs have a tendency here to “bulge out “radially

which has a negative impact on performance. belleville disc

springs are not affected due to high revolutions

Fig. 6. Stiffness at high revolutions

5.3.1 Torque capacity of the clutch A simplified clutch diagram is shown in

figure given below

Fig. 7. Line Diagram of Clutch

In this figure

F = Clamping force of the clutch, N.

P = Pressure on the contact surface, N/ m2.

μ = co-efficient of friction.

Ri = Inner radius of the friction disk, m.

Ro = Outer radius of the friction disk, m.

Consider an elemental ring of radius ‘r’ having thickness of

‘dr’ as shown in above figure.

Force acting on the elemental ring, dF = 2πrpdr

Total force , -----------------

-------------- (1)

Torque transmitted by the elemental ring, dT =

2πrpdr ×μ×r

Total torque transmitted, -----

----------- (2)

There are two case

Case 1: Uniform pressure condition, (p =

constant)

From equn (1), we have,


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--(3)

From equn (2) , we have,

By using equ

n. (3), we have,

---------(4)

Case 2: Wear rate is constant

The rate at which surfaces wear depends not only on

the pressure but also on the velocity of sliding between the

surfaces.

rate of wear = f(p, v)

rate of wear = a1pv [ where a1 = constant]

rate of wear = a2pr [v = 2πrω and a2 =

2πωa1]

If the rate of wear is constant,

pr = constant = a3

The equn. (1) becomes

-----------(5)

And equation (2) becomes

By using equation (4),

-------------- (6)

In our calculation we are going to use equn. (6)

because the torque capacity given by the equn. (6) is less as

compared to the torque given by equn. (4) on top of this when

the clutch is used for longer period the concept of uniform

pressure is no longer valid rather it is a constant wear.

5.3.2 Variation of coefficient of friction with

sliding velocity

Generally, during the clutch engagement the

transienttemperature causes a variation in μwhich may be

consideredto be included in the change in friction with

slidingvelocity( T. P Newcomb. 1961).

Graph 5: variation of coefficient of friction with the

sliding velocity

The graphs show that the coefficient of friction

decreases withrelative sliding velocity. For many materials,

this variation may be approximated to by the following

relationship.

------------ (7)

where,

= Co-efficient of friction at high velocity

= Static coefficient of friction

μ = Co-efficient of friction value at sliding speed v

a = Constant

The tangential relative velocity at radius r is

v = ωrr

Using above relation in equn.(7), we get

--(8)

Using equn.(8) in equ

n.(2) and adopting constant rate

of wear condition

Solving the above equation by using ILATE, we get


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--

------------------------------------------------------- (9)

5.3.4 Rate of energy dissipation

Consider the basic power transmission

system shown in following diagram.

Fig. 8. Basic power transmission

Let two bodies having inertias II and I2initially

rotating at unequal angular velocities Ωland Ω2, and at any

instant ‘t’ rotating at different angular velocities ω1 and ω2

respectivelythroughout the clutch engagement. Il belongs to

the input shaft being driven by a torque T1 and I2belongs tothe

output shaft having a resistive load torque T2. Duringslipping

the torque capacity of the clutch varies as a functionof time (t).

It is assumed that the torques Tl and T2are constant since any

variation in these values is likely tobe small compared to the

uncertainty of their measuredvalues.

When the clutch is start engaging (ω1>ω2 )

---------(10)

------------- (11)

Where initially at t = 0, ω1= Ωland ω2 = Ω2

Integrating equn. (10) and (11) w.r.t time ‘t’ within interval

(0,t) , we get,

---

-----(12)

-----(13)

The relative angular velocity is given by

Using equ

n. (12) and (13) in above equation, we get

Where,

When the system consists purely of two inertias Iland

I2rotatingat different speeds with no external torques, then M

= 0. The above equation becomes

--------------(14)

The slipping period ts, is determined by putting ωr, = 0.

---------

----- (15)

For the peddle operated clutch, torque transmitted by the

clutch during engagement is increasing linearly (T.P

Newcomb, 1961)

--------- (16)

Where,

To = Engine torque.

Using equn.(16) in equ

n.(15) and solving for ts we get,

----------------------------- (17)

Total thermal energy dissipation is given by

------------- (18)

Solving above equation by using equn. (14) and equ

n. (16), we

get

-------------------

--------------- (19)

5.3.5 Distribution of heat between rubbing

surfaces

Fig. 9. Sketch showing distribution of heat between the rubbing

surfaces The above figure shows a single plate clutch under

the action of and axial clamping force, F. Outer two plates are

Flywheel and pressure plate while the middle plates are

friction plated riveted on the clutch disk. Heat is generated at

the rubbing interfaces i.e. between flywheel-friction material


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and friction material-pressure plate. Generated heat is

dissipated between the rubbing bodies depending upon the

properties of the materials. A heat distribution model given by

T.P. Newcomb in 1960 is given below.

`

------ (20)

Where,

q1(t) = heat flux moving toward body I.

q2(t) = heat flux moving toward body

q(t) = heat flux generated at the interface.

k1 = thermal conductivity of body I.

𝜌1= density of the body I.

c1 = specific heat of body I.

k2 = thermal conductivity of body II.

𝜌2= density of the body II.

c2 = specific heat of body II.

Inthe normal operation of a friction clutch repeated

engagements are made and the average temperature of the

assemblyduring operation under these conditions must

bedetermined. In a single engagement the friction

surfacetemperature is highly transient in nature since the

slippingperiod is usually less than a second. The heat

developedthen flows into the components of the assembly

until allare at a uniform temperature. This occurs within

afewseconds and afterwards the heat transfer is mainly by

convectionand radiation. The average temperature rise in a

single engagement is given by

------------------------------ (21)

Where,

Q = total heat entering to the body.

Δθ = average temperature raise in the body.

m = mass of the body.

c = Specific heat of the body.

In a well-ventilated clutch, the principle method of

heat transfer is by force convection to the atmosphere. If the

thermal conductivity of the material are high and operating

time are fair long, Newton law of cooling may be used.

According to the Newton law of Cooling,

when a solid body of massm, specific heat c, and exposed

surface area A cools slowly from an initial temperature θ", the

temperature θat any subsequent time t is given by

---------- (22)

Where,

h = coefficient of heat transfer

θo= ambient temperature

5.3.6 Stability state temperature determination

Fig. 10. Simplified heat transfer diagram

The figure which is shown in above consists of clutch

assembly, casing and outer atmosphere. Heat is generated in

the clutch assembly and it is transfer to the casing by

convection and radiation. Then finally from casing it was

transfer to the atmosphere. When the system reaches the

stability state, heat transfer rate from clutch assembly to

casing is same as the heat transfer rate from casing to the

outside atmosphere.The average rate of flow of heat,

--------------------------- (23)

Where,

Q = total amount of heat generated in time‘t’.

Heat flux from clutch assembly to casing, -------

-------------- (24)

Heat flux from casing to atmosphere, ------

--------------- (25)

Where,

AI= heat dissipating area of clutch assembly.

AII = heat dissipating area of casing.

By combing radiation heat transfer and convection heat

transfer, we get

If temperature difference ( ) is not large, then an

approximation of the above equation can be given as

----------------------- (26)

The heat flux from casing to atmosphere can be given as

----------------- (27)

Where,

ϵI= Emissivity of clutch assembly.


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ϵII = Emissivity of casing.

θI = Clutch assembly temperature.

θII = Casing temperature.

θIII = Outside atmosphere temperature.

ho= heat transfer coefficient outside the casing.

hi= heat transfer coefficient inside the casing.

σ = Stefan-Boltzmann constant

In above equations, the temperature of the

components can be determined if heat transfer coefficient is

known. For determining heat transfer coefficient, Reynolds

number needs to determine.

Reynolds number,

Where,

= velocity of air with respect to surface

X = Critical length of the surface

υ = Kinematic viscosity of air.

Critical value of Reynolds number = 5 ×105

If the value of Reynolds number is less than critical

value then the mode of flow is laminar and corresponding heat

transfer coefficient is given by

-----------------

---------------------- (28)

If Reynolds number is more than critical value then

the mode of flow is turbulent and corresponding heat transfer

coefficient is given by

-------

----------------------- (29)

Where,

ka= thermal conductivity of air

Pr= Prandtl number =

cp = Specific heat of the air

μ = dynamic viscosity of air

5.3.7 Reduced Moment of inertia

Reduced Moment of inertia of all driven

component at the input shaft of gear box is the inertia that is

required at input shaft to accelerate all the driven components.

Its name is reduced because when calculating equivalent

moment of inertia at the input shaft, moment of inertia of all

the driven components except input shaft are reduced by a

factor of square of train value of the component with respect

to input shaft.

Fig. 11. A Simple gear box

Let,

I1 = Moment of inertia of shaft 1 including all components on

shaft 1, kg m2.


shaft 2, kg m2.


shaft 3, kg m2.

T1 = Number of teeth on gear 1.



Then,Grain train value between gear 1 and 2,

Grain train value between gear 3 and 4,

Total Reflected moment of inertia at input shaft

(31)

Assuming the whole tractor is accelerated by the two

rear wheel of tractor i.e. two wheel drive condition.

Let

m = Mass of tractor, kg.

v = linear velocity of tractor, m/s.

I = Equivalent Moment of Inertia, kg m2.

ω = Angular velocity of rear wheel, rad/s.

then,

[since v = ωr]

or, -------------------- (30)

where,

r = rolling radius of rear wheel, m.

6.1 TESTING / EXPERIMENTAL SET UP AND PLAN.

The tractor assembled with dual belleville clutch

system is connected with necessary sensor connections to

continuously monitor and log the speed response, clutch

release load and housing temperature during actuation then

after 500 hours field validation completed .To dismantling the

clutch cover assembly & friction plate to measure clutch

release load and wear measurement details. The experimental

setup is shown


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Field testing of Front-End Loader

Clutch release load Vs Travel Fig. 12. Experimental / Testing Setup

CONCLUSION AND RESULT

7.1 CONCLUSION

The formulated design has eliminated the need for

clutch pedal engagement during forward and reverse motion

and thereby results in reduced operator fatigue. During

operation dual belleville clutch was analyzed for its various

parameters like drag torque, engagement time response,

interfacial surface temperature, wear rate of friction disc

through set of mathematical models.

7.2 RESULT OUT COME These results on the performance parameters of developed

dual belleville clutch show smooth clutch actuation, and

improved reliability suits to perform in heavy duty front end

loading tractor applications.

7.1 FUTURE RESEARCH WORK

The future research work will be extended to

enhancement of the front end loader tractor to provide clutch

breather in clutch housing. This would enable to radiate heat

from one medium to the other and efficiency of the plate will

be more.

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The Higher petroleum Institute, tobruk (Libiya). Wear, 47 (1978)

[2] Ray shaver(Ed), Chrysler Corporation, SAE AE-17 manual

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[4] Robert L.Norton, “Machine design” Pearson Education, Inc,

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[6] Dieter Jentsch, “Hand book for disc Springs” Schnorr corporation,

Ann Arbo, MI48108.

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(IOSRJMCE)

[8] Abdullah, O. I. and J. Schlattmann. 2012. Finite Element Analysis for Grooved Dry Friction Clutch. Advances in Mechanical

Engineering and its Applications (AMEA). Vol. 2, No. 1: 121-133.

[9] Cappetti, N., Pisaturo, M. and A. Senatore. 2012. Cushion spring sensitivity to thetemperature rise in automotive dry clutch and

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[12] Bhandari, V.B. 2010. Design of machine element. 3rd edition. New Delhi: Tata McGraw hill education private limited.

[13] Newcomb, T. P. 1960. Temperatures reached in friction clutch

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[14] EL-Sherbiny, M. and T. P. Newcomb. 1976. Temperature

Distributions in Automotive Dry Clutches. Proceedings of the Institution of Mechanical Engineers. 190: 359.

[15] Sfarni, S., Bellenger, E., Fortin, J. and M. Malley. 2008. Finite

element analysis of automotive cushion discs. Thin-Walled Structures. Vol 47 : 474 - 483.

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commercial vehicle clutch system. Indian Academy of sciences. Vol - 35 (5).

[17] Barber. J. R. 1967. Distribution of heat between sliding surfaces.

Journal Mechanical Engineering Science. Vol. 9 (5) : 351-354. [18] Anderson, a. E. And R. A. Knapp. 1990. Hot spotting in

automotive friction systems. Wear Vol- 135, 319: 337.

[19] Abdullah, O. I. and J. Schlattmann. 2012. Effect of Band Contact on the Temperature Distribution for Dry Friction Clutch.

Engineering and Technology. Engineering and Technology. Vol-

69 : 167-177. [20] Watson, M., Byington, C.,Edwards, D., and S. Amin. 2005.

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First Author Detail Dr.P.S.Senthil Kumar has 14 years of experience in academics and presently is a Professor working at VEL TECH University.He has done his BE in

Mechanical Engineering from CIT,Coimbatore and ME in Product Design and Development from Sona College of Technology,Salem.His research area is

new product Development and had done his PhD from Anna

University,Chennai.

design, development and evaluation of dual belleville...

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