cylinder and cylinder head
TRANSCRIPT
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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.
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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,
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=: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.
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"
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
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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'
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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
-
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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
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