Cylinder and Cylinder Head

7/30/2019 Cylinder and Cylinder Head

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Cylinder and Cylinder Head

Cylinder of an internal combustion engine should be designed to withstand

the high pressure and temperature conditions ; it should be able to transfer the

unused heat efficiently so that metal temperature does not approach the dangerous

limit, and it should be economical to repair it in the event of wear and tear. For

this reason, it is usual to use cylinder liners or sleeves in all the big engines

became of the following advantages :

(i) These are more economical because of ease of replacement after wear

and tear.

(ii) Instead of making the whole of the cylinder of best grade of material,

only the liner can be made of better grade, wear resistant cast iron and the jacket

made of cheaper grade.

(iii) It's, use also allows for longitudinal expansion.

In big engines the various parts, viz., cylinder, water jacket, frame etc. are

manufactured separately, whereas in small engines these are all made as one

piece.

The cylinder liner should be made of such material which is strong enough

to withstand high gas pressure and at the same time sufficiently hard enough to

resist wear due to piston movement. It should also be corrosion resistant and

produce good bearing surface to guide the piston movement. It should also be

capable of resisting thermal stresses due to heat flow through the liner wall. The

various materials commonly used and satisfying the above requirements in the

order of preference are: Grey cast iron with homogeneous and close grained

structure (pearlitic cast iron); which is usually cast centrifugally; nickel cast iron

and nickel; chromium cast iron ; nickel-chromium cast steel (with molybdenum,

in some cases). The aeroplane engine cylinders are made of forged alloy steel.

The cylinders are usually made of cast steel. The inner surface of the liners is

usually heat-treated properly in order to obtain hard surface to reduce wear.

Sometimes it is chromium electroplated to obtain very hard and porous surface

such that an oil film is formed and retained thereby reducing th wear appreciably.


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Two types of liners (wet, a ad dry type) are commonly used. The dry liner

is press fit into tha cylinder and the cooling water

(to remove the unused heat) does not come in contact with the liner but contacts

only the cylinder outer surface. This arrangement is adopted for smaller cylinder

below 125 mm bore, above which wet liners are used in which case the cooling

water in the jacket comes in direct contact with outside layers of the liner. In the

case of wet liner, a heavy flange has to be provided at the top which means that

centre distance between two cylinders in case of multi cylinder engine will be

more and the cooling of the top of liner will not be proper. However it permits

easy supporting of heavy internal cores of cylinder bore. The dry liner is" easier to

replace and the wet liner is difficult. The heat flow through dry liner is poor, but is

uniform throughout including the top. In the case of dry liner there is no

possibility of water leakage into crank case or combustion chamber ; whereas in

wet liner such a risk exists if the liner casting is defective.

The stresses in the cylinder liner are of two typespressure stress which is

tensile throughout, and thermal stresses which result In compressive stress in the

inner fibres and tensile stress in the outer fibres. From past experience it is

observed that cylinder liners rarely fail due to pressure and thermal stresses, but

the failure is generally due to distortion and wear.

The cylinder wall is subjected to gas pressure and the piston side thrust.

Piston side thurst tends to bend the wall but the stress in the wall due to side thrust

is very small and can be neglected. The gas pressure produces two types of

stresses : longitudinal and circumferential which act at right angle to each other

and so the net stress in each direction is reduced.

s l = longitudinal stress

=Area

Force

=22

0

max

2

)(4

4

DD

XpD

Where D = cylinder inside dia,


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D0 = cylinder outside dia.

and pmax = max. gas pressure

sc = circumferential stress

=t

xDP

2max

Nets l =s l -m

so

and Netso

= so

-m

sl

m1 = Poisson's ratio.

Thus t = ks

Dp

o

+

2

.max

Where t = wall thickness in cm.

so

= maximum hoop stress. It varies from 350 (for small bore) to

1000 kg/cm2

for larger bores) depending on the size an material.k = reboring factor.

Cylinder bore

mm

75 100 150 200 250 300 350 400 450 500

k, mm 1.5 2.25 3.75 6 7.5 9 10.5 12 12 1

The thickness of the cylinder walls usually varies from aboi 4.5 mm to 25

mm or more depending upon the cylinder size. For liners of oil engines, thickness

of liner and that of cylinder is greater than fifteenth part of cylinder bore.

The thickness of the dry liner is given as

t'= 0'0 30D to 0035 D

The thickness of the inner walls of the automobile engine cylinders is

usually given empirically as


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t= 0'045D + 0'16 cm.

The thickness of jacket wall is given as 4

3t for smaller cylinder to 3

1t

for bigger cylinder, or

= 0.032D + 0'16 cm.

The water space between the outer cylinder wall and the inn jacket wall is

9 mm for a 75 mm cylinder to about 75 mm f 750 mm cylinder, or

=0'08 D+0'65 cm.

' The cylinder is usually attached to the upper half of the crai 0 ase with the help of

flanges, studs and nuts.

The cylinder flange is made thicker than the wall of tl cylinder, the usual value of

flange thickness being taken as 1'20 1-4 t, or

= 1-25 to 1'50 dtZ=bolt diameter.

The distance of the end of the flange from the centre of the stud or bolt should not

be less than d+6 mm, and not more than 1-5 d.

The diameter of the bolt or stud can be calculated by equating gas load to the area

of all the studs at the root of the threads multiplied by the allowable fibre stress.

t.e.

n=no. of studs and lies toO'2D+4

where

between O'l D+4

s(=allowable fibre stress, 300 to 600 kg/cma for nickel steel bolts.


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to accommodate the ports for air and gas passages, and to accommodate valves

and also accommodate the atomiser at the centre of the cover in the case of diesel

engines.

If the cylinder head is approximately a flat circular plate, then its thickness-can'be

determined by the relation :

st

where

^

(7=constant, in this case equal to O'l. s=allowable stress, (300 to 500 kg/cm2)

Problem 36'1. A four stroke 1.0. engine has the following

specifications :

BHP=10 EPM=UOO

Indicated mean effective pressure

p=3'5 kg/cm2Max. gas pressure pmax.

=.35 kg/cm2.

968

Net

and net

\

MACHINE DESK

9'85-53'25

43'40 kg/cma (compressive)

213-2-46 210-54 kg/cma.

37 Pistons

37 '1. Introduction

Piston is an important part of an I.C. engine which receives impulse from the

expanding gases in the cylinder and transmits the energy to the crankshaft through

the connecting rod. It also disperses a large amount of heat from the combustion

chamber to the cylinder walls. I.C. engines employ trunk type pistons which are


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open at one end and consist of (i) head or crown to carry the cylinder pressure, (it)

skirt to act as a bearing for connecting rod side thrust, (iii) piston pin to connect

the piston to the connecting rod, and (iv)piston rings to seal the cylinder.

The various important design considerations for a piston are :

The piston is subjected to highly rigorous conditions and must therefore have

enormous strength and heat resistant properties to withstand high gas pressures.

Its construction should be rigid enough to withstand thermal and mechanical

distortion. As high speeds up to 15 mpm may be attained in high speed engines

the weight of piston should be minimum possible to minimise the inertia forces.

To maintain the piston temperature within limits, the heat from the crown of

piston must be dissipated quickly and efficiently to-the rings and bearing area and

then to the cylinder walls. , The profile of piston head is dependent on the design

of combuston chamber. The thickness of piston head is determined by the

criterion of strength and the heat to be dissipated. From strength considerations, it

may become necessary to, use different material for head, like-cast steel.

The bearing area of piston should be sufficient to prevent undue wear and it

should form an effective seal to avoid gases from leaking to oil side or oil to gas

side. The number and type of piston rings is influenced by many factors including

the balancing weight of crank.

It should have least friction and have noiseless operation. Material of the piston

must possess good wearing qualities, so that the piston is able to maintain the

surface hardness up to the operating temperatures and there should be little or no

tendency towards corrosion. The most commonly used materials for the pistons of

.internal combustion engines are : cast iron, cast aluminium, forged

969

Aluminiurn , cast steel, and forged steel. Cast iron pistons ma used for moderately

rated engines with piston speeds below 6J and aluminium alloy pistons are used

for highly rated engines running at higher piston speeds, ||

37'2. Thickness of Piston Head ' |


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The thickness of piston head can be calculated by assui the head to be a flat plate

of uniform thickness and fixed at edges and assuming the gas load to be uniformly

distributed, f

where


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Higher calorific value for diesel fuel may be taken as 10,500 teal/kg.

BHP=brake horse power of engine per

cylinder.

(7 constant (usually =0 '05), represents the portion of the heat supplied to the

engine which is absorbed by the piston.

37'3. Piston Rings

These impart the necessary radial pressure to maintain the seal between the piston

and the cylinder bore. The piston rings inserted at the top function as compression

rings or pressure rings, and may be 3, 4, 5, 6 and 7 in number. These also transfer

heat from piston to cylinder liner and absorb part of piston fluctuation due to side

thrust. The rings inserted at bottom serve as oil scraper or oil control rings. These

provide proper lubrication by allowing sufficient oil to move up during upward

stroke and at the same time also minimise oil flow to combustion chamber.

In the oil rings, either the bottom outer edge is stepped or upper edge bevelled, or

slot cut in the centre of the ring all around the periphery and the lower edge of the

groove of the piston is bevelled and small holes drilled towards the inside of the

piston so that the excess oil scrapped by the ring flows through these holes into

{he piston and falls into the sump.

The compression rings are usually made of rectangular cross-section and the

diameter of the rings is made slightly larger than the cylinder bore. A part of the

ring is cut-off in order to permit it go into cylinder against the liner wall. This also

produces pressure on rings,. The gap between the ends should be sufficiently large

so that even at the highet temperature the ends will not touch each other,

otherwise there might be buckling of the ring, The square cut ends are most

commonly used. Sometimes angular cut, or square step cut ends are also used.

The ring joints of various rings should be spaced equally round the piston and

should not come one below the other. The rings are sometimes, therefore, located

by pins, so that all the gaps do not come in one line due to the rotation of the

rings.


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Piston rings are usually made of cast iron and alloy cast iron due to their good

wearing qualities and also their retaining the spring characteristics even at high

temperatures. To avoid the Wear, the rings are chrome plated.

The radial width (wr) of the ring is selected so as to limit the the wall pressure to

0'25 to 0'42 kg/cm2.

'

where w, =radial width of ring, cm. #to=wall pressure, kg/cm2. st =allowable

stress in bending, kg/cm2.},.. =850- 1100 for G.I. rings.

The axial thickness of the rings may he taken more.

The thickness between the ring grooves i.e. the land ,i$ taken as equal to or

slightly less than the axial thickness -:W

ring. ' . . ;.||

The width of the top land is made larger than the otlierj! lands (about 0'2 to 0'3

Z>) to protect the top ring from;. thel| temperature conditions existing at the top of

the piston. ||

37'4. Piston Skirt . .

The portion of the piston barrel below the ring sectiojp the open end is known as

"skirt" and it takes the side thrust', { connecting rod. Its length should be such

that the side pressure does not exceed 2'5 kg/cma for low speed engines kg/cm2

for high speed engines.

Side thrust y.

75

T

'.*

/Lt=co-efficient of friction, between skrit (0'03 to O'lO).

Z>=piston diameter, cm ;

p=gaspressure Z=skirt length, cm,

ptside thrust pressure, kg/cm2 .*. Length of piston, L*=l+length of ring

section+toprlaj It usually varies betweenD and 1*5 D.


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Solved Problems :"

Problem. 37'1. What are the purposes piston. '

Solution. Generally 4 to 6 ribs of 0'3 to 0-5 times the$ ness of piston head are

provided radially in the head of t>-~ ~ and these in general provide the

stiffening effect and th make the head rigid and ' capable oi withstanding gas

and thej loads. These also assist in transferring heat from the piston hej the

piston rings and then to the cylinder liner. . .- ^

A stiffening rib which is provided at the centre line of * ! " and extends around the

skirt, helps in properly transmitting ^utt; thrust from the connecting rod to the

skirt through the pist


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Aluminium pistons are very light and due to high thermal Conductivity these

dissipate heat in a better way. At high tempera-:ures it loses strength, (50% loss

above 320C). It has high expansion about 2'5 times that of G.I.). It has low

abrasion resistance at high temperature.

' All these disadvantages are easily overcome by alloying alumi-lium with other

materials. Cast aluminium alloy has nearly the iame strength as C.I.

fir Proble-in. 37'4. Why clearance is provided between the piston ind cylinder

liner, and how much ?

p Solution. The clearance between the piston and the cylinder ijiier is provided to

account for the thermal expansion of piston and stortion under load. Its magnitude

depends upon the engine igEign and the piston size. It may be mentioned that

whereas the gcessive clearance will lead to piston slap, the less clearance would

[fad to seizure of piston. The top dia. of piston is kept as 0'974 to 96 D and

bottom dia. as 0 9993 to 0'9996D.

llfr To account for thermal expansion, the piston on top portion is |||de tapered, the

amount of taper being dependent upon the opera-pjjg temperature of the piston,

the relative expansion between the |ljton, and liner, and the running clearance

adopted.

$P;' Problem 37'5. Design a cast iron piston for a four-stroke, A cylinder,

semidiesel engine running at 600 r.p.m. The maximum ^osion pressure on the

cylinder head is to be near about 40 kg/sq. cm, |B;mean effective pressure is

about 7'5 kg/sq cm. The fuel consump. p is 0'20 kg per B H.P. per hour.

Diameter and stroke of the ||Q/I is 25 cm and 30 cm respectively. The

connecting rod length is The piston is to have at least 3 sealing rings and 2 oil

rings. pressure of the rings should be between 0'35 to 0^42 kg/sq Permissible

bending stress for the piston ring materiat is to be

974

MACHINE D

about 800 kg/sg cm< Heat conducted through the piston cro approximately 4\ to

5% of the total heat produced. Beat coneLui of G.I. is about 0'4 kcal/cm hr G.


12/23

Temperatures at the centre and edges of the piston face may be assumed

approximately 320C 150Gjespectivdy. BCV of fuel may be taken as 10,000

kcal/kg.

t.,=25

All dimensions in mm I*

Fig. 37-1 . ,,;

Solution,

Piston bead. The thickness of the piston head 'is prii

found from its heat dissipation capacity. Heat flow throng

headfper hr in kcal. =(

.Cf=a constant, representing portion of heat,-

mitted by the crown of the piston ; .'.%

=0-045 to^'05 J|

HGV=higber calorificwalue of the fuel :| ^=10,000 kcal/kg v 'Vj

TF=weight of fuel used in kg per B.H.P. p

Now, B.H.P. of the engine ;'pm.L.A.N . : - 4500

SSTONS

975

lere,

#m=mean effective pressure, kg per sq. cm i=length of the stroke, metres Aarea,

of cylinder bore, sq. cm 2Vr=namber of explosions per minute r.p.m.

BH.P.=

7'5 X 0-30 X 07854 X 25 X 25 X 600 X 0'8

4500X2

(taking >j=0'8)

=59

#=10,000x0-05x0-20x59 = 5900 kcal/hr.

Minimum thickness tof the crown or head necessary to dissi-fte the heat may

be calculated from the equation :

t=-


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H

K

T

.' T1=ten)perature at the centre

=320C Ta=temperature at the edge

, =150G '

&=heat conductivity factor ' =0-4 kcal/cm hrG I,, . ^

______5900______ 16XO'4[320-150]07854 =6'9 cm, say 7 cm

The strength of the piston head may now be checked by ing it as a circular plate

carrying a uniformly distributed load xed at the edges. Grashof's formula for

maximum stress for i case is

at==fibre stress in kg/sq. cm

=200 kg/cm2

p=maximum gas pressure in kg/sq. cm =40

D=diameter of piston face =25 cm

976 Hence

MAC

-v-

3X40X25X25 16XZOC

=4'85cm .*. 7 cm may be taken as thickness of piston head,^

Radial ribs may be 4 in nos. and their thickness

7 7 '' 0'33 to 0'5 t. Here we will allow -75- to =2'5cm, sa

3

If the piston be cooled by oil, then the valug|| be reduced considerably but

not less than the ' calculat^ of4'85cm.

Piston Rings

The radial width of a G.I. cut ring can be calculated

I-3 "*|

following empirical relation, wr=D A/ ; where, ^i


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width ; pw=radial wall pressure : st=permissible ben|| in the ringspurposely

taken very high from 770 toj cm. to avoid an excessive stress when slipping

[-the :ri| piston. ' . '

Hence, wv=$

Axial thickness of the ring may be taken between;(||||| or may be based on the

number of-rings required ; i e. >lf"; '*

''iC Minimum thickness W

_D_

~~\Qn , '.

where w=no. of rings.

.*. Minimum thickness

25 ';-:


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um thickness t^of the cylindrical portion may be determined jrom the equation,


16/23

=:33*5 cm "m

The longer the piston is, the better is the bearing svrF"' affording quieter running

of the engine, but unnecessary long, L will increase its own weight and thus the

inertia forces in1 3

cation. ' ?$$$

^;4i Gudgeon Pin. In order to obtain a uniform distributi

side pressure between the piston and the cylinder wall, the giic pin should be

placed near about from the edge of the open ,e

the piston. In fact, the skirt can be treated as a crosshead; ^ in order to reduce the

weight of the gudgeon pin it can be;S hollow. The length l^of the pin bearing in

the small r connecting rod is limited by the piston diameter. Thereimx, diameter

of the gudgeon pin has to be decided according t allowable bearing pressure on

the projected area available. "

For oil engines of the type in question, a maximum be pressure of 125 to 155

kg./sq. cm. is possible with a ratio of 1-5 to 2.

Assuming li=l'5do (outside dia. of the pin) and p=140 kg./sq. cm., 19,600 =

r5daXd0X140 ;

or

19.600

=9'74 cm, say 9'8 cm

The dimensions, especially the diameter of the on the higher side for

accommodating in a piston diameter of% and should therefore be reduced

somehowsay 'by increasii intensity of bearing pressure to 200 kg/sq. cm. or

nearaBp,i improving the heat-treatment and material. Si

Try a 8'0-cm. diameter X 12'0 cm. long solid pin, Then, 19,600=8-0x12-0^

19 600 P==8-Oxl2-Q =20 kS-/S(l' cm,.say*l8

or

The bending stress in the pin is determined from the cpni tion that the pin is a

beam uniformly loaded for a distance || freely supported at the bosses and making

an effective span of?

25+12-0


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=18-5

SSTONS

979

Then, max. B.M. on the pin.

,, 19,600x25 ,.. , -B.M. = =61,300 kg. cm.

8

Section modulus of the pin

71

= Xdia. of pin3

O"i

=i50 61,300

8t=-

50

= 1225 kg./sq. cm. approx.

I As the stress induced in the pin is on the higher side, the mate-| of the pin may

be taken as 0'35 to 0'4% carbon steel.

I If the pin be made a nice push fit in the piston, then the |ue ofstwill be reduced

to some extent. As a matter of fact |"ding theory does not rigidly apply to cases

where the ratio of |;span to beam .depth does not exceed 2, as in this case.

Empirically, maximum diameter of gudgeon pin varies from fi&o 0'4D for G.I.

pistons of 20 cm. to 40 cm. diameter.

According to this erntpirical rule, the maximum diameter of |geon pin will

vary fromO'3x25 to 0-4x25=7'5 cm. to 10 cm.

I We have provided 8'0 cm. (O.K.).

|v}The pin must be located yyays in the piston to pre-'ijfepontact with the cylinder

' |t There are several methods

he use of bronze buttons into the pins is unwise,

pre is a tendency towards

|fing 'the cylinder walls.


18/23

"

plate of circular shape j||ecessed in the piston wall dot touching the cylinder

may be employed as |in Fig. 37'2.

^Piston Clearances. Clearances are necessary due to thermal jjsion or due to

distortion under load. If excessive clearance is

Button Type (not recommended)

Fig. 37-2.

Keep Plate Type (recommended) for pistons upto 25 cm. bore cylinder.

980

MACHINE 1

allowed, the phenomenon known as "piston slap" will result; phenomenon is not

only an objectionable noise, but resu damage. Too little a clearance will cause

seizing of the] For C.I. pistons, as in this case, clearance near the head is 0'87 near

the bottom-most oil ring groove is 0'2 mm. approx., and. i bottom open end is 0'15

mm. approx.

Thermal expansion may be counteracted Ly making the more or less tapered. The

amount of the manufactured taper re will depend on(t) the operating

temperature of the piston, (ii relative expansion between the piston and the

cylinder wall () the running clearance adopted.

Distortion due to load is caused by the bending of the gi: pin and bosses. This is

aggravated by local heating due to fricl the top (small) end of the connecting rod.

It is counteracted by grinding an additional clearance over an arc of 90 on eacl of

the piston on the pin centre linewor by grinding the piston ,c shape as shown in

Fig. 37'3.

*-BEAR!NG

ARA >

O

O


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0

o

o

o

0 3 D 0

0

3 3

0 3 3

p

ADDITIONAL |

CLEARANCEofo^sjiM

Fit. 37-3 ;;;|

Unsolved Problems r =1

37'1. What are the functions of a trunk type of piston ? What be kept in view in

the design of I.C. engine pistons ? What is the dil between trunk piston and barrel

piston ? 1: ||

37'2. How would you proceed to design an I.C. engine pisfnn fi data supplied ?

How would you calculate the maximum explosion p the size of the cylinder bore

if this data is not supplied ?

37'3. What do you understand by composite pistons and wl... "advantages? . ~m

37'4. Name two suitable metals for I.C. engine pistons. Of WhaM piston rings be

made and why ? Piston rings are sometimes seen plate some metal or alloy. Why

it is so and what is the name of the platedM alloy? . ' .-''/Si!

37'5. It is always desired by the designers of I.C. engine, to;jj the finished weight

of the piston and the connecting rod. Why it is so;?''||

37'6. Explain the terms piston slap; piston seizure; scuffing.

PISTONS

981


20/23

37'7. How many minimum compression spring rings and how many oil scraping

rings be there in an I.C. engine ? What is the use of allowing ribs inside the piston

j'


21/23

37'19. Design and draw to scale a trunk type of piston for a single voider, four

stroke cycle, diesel engine from the following data :

B.H.P.-^ 10 Speed 1000 r.p.m.

Allowable bearing pressure on gudjjeon pin 105 to 140 kg./cm1. Permissible stress

in gudgeon pin of carbon steel 560 kg./cm.1 Weight of reciprocating parts

per sq. cm. of piston 0'14 kg. Ratio of connecting rod to Crank

4

In the drawing show(a) piston and scraper rings in position and (b)j?pn pin

with method of its fixing.

|;\Assume suitable values of mechanical efficiency, ratios of compression

Mansion, indices of compression and expansion curves, explosion pressure, iin

factor, ratio of the stroke to bore, etc.,-$"-

'''>' 37'20. Design the cylinder size and the aluminium alloy piston

dimensions ;ix cylinder automobile petrol engine or diesel engine to deliver 125

B.H.P. iO r.p.m. The following particulars are furnished :

;;! ForPetrol Engine. Petrol air mixture has a heating value of 606'7 kilo-gligs per

cubic metre of suction displacement. Suction pressure is 0'945 kg./ "a| the

beginning of compression and 8'05 kg./cm2. at the end of compression.

- . MACHINE 1

the compression curve being Pp-S2=C. The stroke bore ratio is 1'2 ai mechanical

efficiency is 85%.

For jpiesel Engine, Suction pressure and temperature is 0'945 kg and 49C.

Pressure at the end of compression is 36'75 kg./em. z, the eq of .the compression

curve being PK1-85i=C. Fuel consumption is 0'2 k B.H.P. per hour ; the

calorific'value of the fuel being 10890 kilp-calprj kilogram of fuel. Mechanical

efficiency and bore ratio are as per the engine.

37'21. Design a trunk type cast iron piston for "an I.C. engine fri .following

data:

Diameter of the cylinder , 10cm.


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Stroke . 15 cm

Maximum explosion pressure 35 kg./sq. cm

Maximum permissible tension for cast iron for the design ( thickness is

300*kg./sq. cm. and the flexural stress for the pin may be take 900 to 1200 kg./sq.

cm. The gudgeon piri should be hardened and grpui should turn in phosphor

bronze bushing. Bearing pressure should belimi 200 kg./sq. cm.

Sketch the rpiston (cross-section) inserting important dimensions p istcn with

piston pin, piston rings and scraper ring in position.

Check the design from heat transfer view point.

38 Connecting Rods

Connecting rod is used to transmit motion from the reciprocating piston to the

rotating crank. It also conveys the lubricating oil from the crank pin to- the

piston pin and provides splash or jef cooling of the piston crown. In the most

usual form it consists of an eye at the small end for the piston pin bearing, a

long shank usually of I-section, and> big end opening which is usually split to

take the crankpin bearing shells. Low speed large engines usually employ

circular section with flattened sides, or rectangular section, tv'ith the larger

dimension being in the plane of rotation. High speed engines employ I-

section or H-section rods for lightness. I- sectiori is most common for high

speed engine connecting rods because lightness is essential in order to keep

the inertia forces as small as possible. I-section also provides ample

strength required to withstand the momentary high gas pressure in the

cylinder. Thus I-section fulfills the most desirable conditions for connecting

rod, i.e. the adequate strength and stiffness with mini-num weight.

The connecting rods of internal combustion engines are mostly nanufactured by

drop forging with outer surfaces left unfinished, rhese are usually made of carbon

steel with ultimate tensile strength 5500 to 6700 kg/cm2) for industrial engines, of

manganese alloy teels having a strength (7800 to 9400 kg/cm2) for transport

ingines, of nickel chrome steel having ultimate tensile strength 9400 to 13,500


23/23

kg/cm4) for aero engines, of duralumin and alumi-lium alloys for high speed

engines.

The length of connecting rod is usually made 4 to 5 times .crank radius.

The smaller length than this increases the Elarity of connecting rod

which increases the side thrust of Jpn against liner and thus the wear. More

length would me,an liter height of engine and thus a compromise is essential.

The lubrication of the two end bearings of the connecting rod {Every important;

Two methods commonly used are :

1. Splash lubrication.

2. Pressure feed lubrication.

';'- In splash lubrication, at the big end of the connecting rod is |tached a spout

which dips into the lubricating oil in the sump Siring downward motion of

connecting rod, and a splash of oil is

I' 983

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Cylinder and Cylinder Head