darshan institute of engg. & tech. · 1. two shafts at right to each other are connected by a...

DARSHAN INSTITUTE OF ENGG. & TECH.

Department of Mechanical Engineering

B.E. Semester – VII

Machine Design (2171909)

Batch: ____________ Roll No.: ___________

List of Assignments

Sr.

No. Title Start Date End Date Sign Remark

1. Design of Spur Gears

2. Design of Helical Gears

3. Design of Bevel Gears

4. Design of Worm Gears

5. Design of Gear Boxes

6. Design of Journal Bearing

7. Selection of Rolling Contact bearing

8. Design of I. C. Engine components

9. Design of Crane Components


B. E. Semester – VII Department of Mechanical Engineering Darshan Institute of Engineering and Technology, Rajkot 1

ASSIGNMENT – 1 Design of Spur Gears

Theory

1. Explain the different causes of gear tooth failures and suggest possible remedies to

avoid such failures. (V. B. Bhandari – 17.12)

2. What do you mean by interference and undercutting of gear? How it can be avoided?

(V. B. Bhandari – 17.9)

3. Explain standard system of gear tooth. Also discuss advantages and disadvantages of

14.5° and 20° involutes system. (V. B. Bhandari – 17.7)

4. Explain law of gearing. Prove that “The common normal to the tooth profile at the

point of contact should always pass through a fixed point, in order to obtain constant

velocity ratio”.

5. Differentiate between involute and cycloidal profile of the gears. (R. S. Khurmi –

28.10)

6. Give broad classification of gears along with their applications.

Examples

1. Design a spur gear pair from the following given data:

Power to be transmitted = 22.5 kW,

Pinion speed = 1450 rpm,

Speed reduction = 2.5,

No. of teeth on pinion = 20,

Service factor = 1.5,

b= 10 m, (m = module),

Pitch line velocity = 5 m/sec (For initial calculation of module),

Maximum permissible error in gear tooth profile = 0.025 mm,

k = A factor depending upon the form of teeth = 0.111,

Velocity factor =3

3 V, where V is the pitch line velocity in m/s.

Take endurance surface hardness = 600 MPa.

Lewis form factor =0.912

0.154 – No. of teeth

for 20⁰ pressure angle involute tooth

system.

The materials and stresses are as under:

Material [σb] Elasticity Modulus Hardness Pinion (Fe 410) 135 N/mm2 2.1 x 105 N/mm2 260 BHN Gear (FG 200) 65 N/mm2 1.1 x 105 N/mm2 250 BHN



2. Design a spur gear pair to transmit 15 kW power from an electric motor shaft

running at 1500 rpm to a machine shaft from the following specifications:

Tooth system = 20⁰ pressure angle full depth involute,

Number of teeth on pinion = 25,

Speed reduction ratio = 3:1,

Service factor = 1.25,

Material of pinion and gear = FG 200,

Design bending stress of material = 60 MPa,

Surface hardness of pinion and gear = 200 BHN,

Endurance strength of the material = 84 MPa,

Dynamic load factor = 178 N/mm,

Modulus of elasticity = 1.1 x 105 MPa.

Assume pitch line velocity as 7.5 m/sec for module calculation.

Velocity factor, 3

3VC

V

Lewis form factor, 0.912

0.154yZ

for 20⁰ pressure angle full depth involute

tooth system

Dynamic load equation, 21 ( )

21

td t

t

v f C FF F

v f C F

Wear load, w pF d KW Qf

3. A pair of mating carefully cut spur gears has 20° full depth of 4 mm module.

The number of teeth on pinion and gears are 38 and 115, respectively.

The face width is 40 mm.

If the pinion and gear are made of steel with fb static = 233 MPa and surface hardness

of 300 BHN, calculate the safe power that can be transmitted when the pinion is run

at 1200 rpm.

4. A spur gear pair made of plain carbon steel 55C8 (σut = 720 N/mm2 and E = 210

GPa) is required to transmit 7.5 kW power from an electric motor running at 1440

rpm to a machine running at 370 rpm.

The tooth system is 20⁰ full depth involute and number of teeth on pinion is as

minimum as possible.

The service factor and load concentration factor are 1.25 and 1.2 respectively.

The factor of safety required is 1.25 to 1.5.

The face width is twelve times the module. The gears are to be machined to meet the

specifications of grade 7.



Design the gear pair by using the velocity factor3

3VK

V

and Buckingham’s

equation for dynamic load.

Suggest the case hardness for gear pair.

Use the following relations.

2.87

0.484

YZ

max

max

21 ( )

21

ta

t

v b C FF N

v b C F

0.111 /p g

p g

E Ec e N mm

E E

2

20.16 /100

BHNK N mm

For grade 7, 11 0.9( 0.25 )e m d m

5. A spur gear having 22 teeth to be made of plain carbon steel 40C8 (Sut=580 N/mm2)

is to be mesh with a gear having 88 teeth to be made of grey cast iron FG260

(Sut=260 N/mm2). The pinion shaft is connected to 12 KW, 1440 rpm electric motor.

The starting torque of the motor is approximately twice the rated torque. The tooth

system is 20° full depth involute. The face width is 10 times module for which the

load distribution factor is 1.4. The gears are to be machined to meet the

specifications of grade 7 for which deformation factor is 240 N/mm.

(i) If factor of safety required against bending failure 1.0, design the gear pair by

using velocity factor 6

6VK

V

and Buckingham’s equation for dynamic load.

(ii) If the factor of safety required against pitting failure is 1.5, specify surface

hardness.

2.870.484

YZ

max

max

21 ( )

21

td t

t

v b C FF F

v b C F

,

2

0.18100

BHNK

for steel pinion and cast iron

Standard modules are: 4, 5, 6, 8, 10, 12 and 16.

Service factor = 2, Load concentration factor = 1.4



ASSIGNMENT – 2 Design of Helical Gears

Theory

1. Explain the following terms used in helical gears:

(a) Helix angle; (b) Normal pitch; (c) Axial pitch; (d) Normal Pressure angle

2. What are the advantages of helical gears over spur gears? (Dr. Sadhu Singh – 19.2)

3. What is formative number of teeth in helical gear? Derive the expression for

formative number of teeth in helical gears. (V. B. Bhandari – 18.3)

4. What is herringbone gear? State advantages of herringbone gear.

Examples

1. A pair of parallel helical gears consists of a 20 teeth pinion meshing with a 100 teeth

gear. The pinion rotates at 720 rpm. The normal pressure angle is 20°, while the

helix angle 25°. The face width is 40 mm and the normal module is 4 mm. The pinion

as well as the gear are made of steel 40C8 having ultimate tensile strength of 600

N/mm2 and heat treated to a surface hardness of 300 BHN. The service factor and

the factor of safety are 1.5 and 2 respectively. Assume that the velocity factor

accounts for the dynamic load and calculate the power transmitting capacity of

gears.

2. Determine the power capacity of a pair of helical turning gears having a

transmission ratio of 10:1. The teeth are 20° full depth involute – 6 mm module. The

pinion has 25 teeth and rotates at 5000 r.p.m. The active face width is 76 mm and

material is C-40 steel untreated.

3. A pair of parallel helical gears consists of 24 teeth pinion rotating at 5000 rpm and

supplying 12 kW power to a gear. The speed reduction is 4:1. The normal pressure

angle and helix angle are 20° and 23° respectively. Both gears are made of hardened

steel (Sut = 600 N/mm2). The service factor and factor of safety are 1.5 and 3

respectively. Calculate:

(i) Module and face width of gears.

(ii) Surface hardness for the gears assuming a factor of safety of 1.5 for wear

consideration.

4. A helical gear speed reducer is to be designed. The rated power of the speed reducer

is 75 kW at a pinion speed of 1200 rpm. The speed ratio is 3:1 for medium shock

condition and 24 hours operation. Determine module, face width, no. of teeth in each

gear. Specify material & heat treatment. The teeth are 20⁰ full depth in the normal

plane.



5. A pair of helical gears having a transmission ratio 8:3, with a steady load condition,

used for turbine. The maximum speed is 2400 r.p.m. The pinion is to have 27 teeth

and a face width of 100 mm. The circular module is 6 mm. The material used for

gears is heat treated steel with 250 BHN and have static stress of 210 MPa. The gears

are carefully cut. Calculate value of dynamic load and wear load.

6. The following data is given for a pair of helical gears made of steel:

Normal module = 5 mm, Face Width = 50 mm, No. of Pinion Teeth = 30,

No. of Gear Teeth = 60, Centre distance = 245 mm, Normal Pressure angle =20°,

Pinion speed = 1000 r.p.m, surface hardness = 300 BHN, FOS = 2,

Service Factor = 1.5, Grade of Machining = 8, Tooth form factor (Y) = 0.385,

Permissible σb for pinion and gear material=150 N/mm2.

Determine:

(i) Helix angle,

(ii) Beam strength,

(iii) Maximum static load that gear can transmit and

(iv) Power transmitting capacity



ASSIGNMENT – 3 Design of Bevel Gears

Theory

1. Explain terminology of bevel gear.

2. What is formative number of teeth in bevel gear? Derive the expression for

formative number of teeth in bevel gears. (V. B. Bhandari – 19.2)

Examples 1. Two shafts at right to each other are connected by a bevel pair having full depth

involute teeth. The pinion having 20 teeth transmits 40 kW at 750 rpm to gear shaft

running at 375 rpm. Take allowable static stress for pinion and gear materials 100

N/mm2 and 70 N/mm2 respectively. Determine module, pitch diameters and face

width from strength considerations.

2. It is required to design a pair of bevel gears, which are mounted on shafts

intersecting at right angles. The pinion receives 20 kW power through its shaft and

rotates at 720 rpm. The number of teeth on pinion and gear are 30 & 45

respectively. The pressure angle is 20° full depth teeth form. The gears are made of

plain carbon steel with permissible bending stress as 200 MPa. The gears are case

hardened and the surface hardness is 300 BHN. Take service factor = 1.25.

3. Design a pair of bevel gears for two shafts whose axes are at right angles. Speed of

pinion shaft is 240 rpm and that of gear shaft is 120 rpm. Power at gear shaft is 75

kW.

4. A pair of straight bevel gears, manufactured by generation, consists of 14 teeth

pinion meshing with 85 teeth gear. The module at large end is 5.5 mm while the face

width is limited to 0.25 times the slant height. The pinion is made of steel having

ultimate tensile strength is 750 N/mm2 and surface hardness of 210 BHN. The gear

is made of cast iron having ultimate tensile strength 260 N/mm2 and surface

hardness of 210 BHN. The shaft angle is 90°. If the pinion rotates at 750 rpm,

estimate the power that gear pair can transmit.



ASSIGNMENT – 4 Design of Worm Gears

Theory

1. Explain the importance of thermal considerations in worm and worm gear design.

(V. B. Bhandari – 20.9)

2. Explain about efficiency of worm gearing.

3. Explain the designation “z1/z2/q/m” used for the pair of worm and worm gear.

4. Give advantages and drawbacks of worm gear. (V. B. Bhandari – 20.1)

Examples

1. A triple threaded worm rotating at 1200 r.p.m. drives a worm gear having 36 teeth

and transmits 15 KW power. The teeth are of 20° full depth involute profile. The

axial pitch of the worm is 30 mm and pitch diameter of 60 mm. The co-efficient of

friction is 0.03. Calculate:

(i) Helix angle of worm, (ii) Speed ratio, (iii) Centre distance between two shafts,

(iv) Apparent stress in the worm gear and (v) Efficiency of drive.

2. A speed reducer unit is to be designed for an input of 1.1 kW with transmission ratio

27. The speed of hardened steel worm is 1440 rpm. The worm wheel is to be made

from phosphor bronze. The tooth form is to be 20° involute. Take center distance

between worm and worm wheel = x =100 mm.

Pitch circle diameter of worm =0.875

1.416

x , worm is double start.

6

6vC

v

, Form factor

0.9120.154y

T

Allowable stress for phosphor bronze = 84 MPa, flexural endurance limit for

phosphor bronze = 168 MPa, load stress factor = k = 0.55. Check for

(i) Tangential load- power transmitted due to tangential load, (ii) Dynamic load,

(iii) Static load or endurance strength, (iv) Wear load, and (v) Heat dissipation.

3. The speed reducer unit is to be designed for an input of 2 kW at 1600 rpm. The

velocity ratio is 25. The worm is to be made of hardened steel and the gear of

phosphor bronze having a static stress of 70 MPa. The approximate distance

between two shafts is 120 mm. Take velocity factor 6

6vC

v

and tooth form factor

0.9120.154

g

yZ

and wear factor of 0.7. Find: (i) Standard module of gear,

(ii) Face Width of the gear & length of worm and (iii) Check the design for wear load.



ASSIGNMENT – 5 Design of Gear Boxes

Theory

1. Explain the procedure of designing multi speed gear box.

2. Why is it necessary to provide multispeed drive for a machine tool? What are the

basic considerations in design of multi speed gearbox?

3. Explain what is structural diagram and method of drawing structural diagram of

gear box.

4. Explain difference between structure diagram and speed diagram giving suitable

example.

5. Give step by step procedure for the design of 8 speed drive for a lathe giving

governing design equations.

6. Draw the ray and speed diagram for a nine speed gear box. State the necessary

assumptions taken.

Examples

1. Draw speed ray diagram and layout for a 6 speed gear box. The output speeds are

160 r.p.m. minimum and 1000 r.p.m. maximum. The motor speed is 1440 r.p.m.

2. Draw the structure and speed diagram for a gear box having operating speed range

from 56 rpm to 1000 rpm. Use R4 series, with standard spindle speed. The gear box

is connected to a motor driven by a pair of pulleys. Assume the motor speed to be

1440 rpm. Draw the gear box layout diagram.

3. Design a suitable speed gear box for a head stock of a lathe that has a variation of

speed from 105 rpm to 690 rpm in 9 steps. The power is supplied by an electric

motor of 10 kW capacity running at 1000 rpm and driving the input shaft through a

V belt drive having a speed reduction of 2:1. Draw the structural diagram, speed

chart and determine the number of teeth on each gear.

4. A speed gearbox for a head stock of a lathe machine is to give speed variation from

125 rpm to 1500 rpm in 12 steps. The power is supplied to the input shaft by an

electric motor of 5 kW running at 1500 rpm, through a belt drive, giving speed

reduction 1.2:1. Find the speed steps arranged in geometric progression. Draw the

structural diagram, Ray diagram and speed charts.



5. A radial drill machine using a gear box is required to give 8 stepped speeds. The

motor power is 4 kW at 1440 rpm. The power from motor to the input shaft of gear

box is transmitted by a V belt drive giving a speed reduction of 1:6. The minimum

and maximum speeds are 70 rpm and 1800 rpm respectively. Make layout diagram

of gear box. Draw ray diagram and speed chart.

6. A 2 x 2 drive is required to be designed for transmitting speeds starting from 400

rpm with a geometric progression of 1.4. Draw a suitable structure and speed

diagram. Also draw the layout of the gearbox and determine the number of teeth on

each gear.

7. A 3 stage gear box with 12 speeds is to be designed based on R10 series with

minimum spindle speed of 125 rpm. The second stage consists of 3 speed steps. The

electric motor is connected to the gear box through a belt drive and runs at 1440

rpm and transmits of 5 kW. Using standard spindle speeds,

a) Draw the structure and speed diagram for the arrangement,

b) Determine the ratio of the belt pulley diameters,

c) Draw the gear box layout and

d) Determine the number of teeth on each gear of the gear box

MACHINE DESIGN (2171909)


ASSIGNMENT – 6 DESIGN OF JOURNAL BEARING

Theory

1. Explain lubricant and properties of lubricants for sliding contact bearing.

2. Give classification of bearing.

3. Explain hydrodynamic bearing & bearing modulus with neat sketch. OR

Explain the performance of a hydrodynamic bearing with the curve of µ versus Z.N / p.

4. Explain following terms for Journal Bearing

Bearing Characteristic Number, Viscocity Index, Somerfied Number, Bearing Modulus.

5. Explain Petroff’s equation for hydrodynamic journal bearings.

6. Explain the design procedure for hydrodynamic journal bearings.

7. Explain the thermal consideration in journal bearing design.

8. Explain the effects of the L/D ratio and C/D ratio parameters on the performance of

journal bearing.

Examples

1. The following data is given for a 3600 hydrodynamic bearing.

Radial load = 3.2 kN, Journal speed = 1490 rpm, journal diameter = 50 mm, bearing

length = 50 mm, radial clearance = 0.05 mm, viscosity of lubricant = 25 cP. Assuming

that the total heat generated in the bearing is carried by the total oil flow in the bearing,

calculate (i) Coefficient of friction (ii) Power lost in friction (iii) Minimum oil film

thickness (iv) Flow requirement in liter/minute. (v) Temperature rise.

l

d S oh

c

ohf

c

s

Q

rcn l

1 0.264 0.6 5.79 3.99

0.121 0.4 3.22 4.33

2. The following data is given for a 360° hydrodynamic bearing:

Radial load = 10 kN, Journal speed = 1440 rpm, Unit bearing pressure = 1000 kPa,

Clearance ratio (r/c) = 800, viscosity of lubricant = 30 mPa.s. Assuming that the total

heat generated in the bearing is carried by the total oil flow in the bearing, calculate (i)

Coefficient of friction (ii) dimensions of bearing (iii) power lost in friction (iv) total flow

of oil (v) side leakage.

r

f 9.55c

Q3.78

rcnl

sQ0.38

Q



3. The following data is given for a 3600 hydrodynamic bearings:

Radial load : 3.1 kN

Journal diameter : 50 mm

Bearing length : 50 mm

Journal speed : 1440 rpm

Radial clearance : 50 microns

Viscosity of lubricant : 25 cP

Density of lubricant : 860 kg / m3

Sp. Heat of lubricant : 1.76 kJ / kg 0C

Assuming that the total heat generated in the bearing is carried by the total oil flow in

the bearing. Calculate:

i) Sommerfeld Number

ii) Minimum oil-film thickness

iii) The coefficient of friction

iv) The power lost in friction

v) The total flow rate of lubricant in liter /minute

vi) Side leakage

4. The thrust of propeller shaft in a marine engine is taken up by a number of collars

integral with the shaft which is 300 mm in diameter. The thrust on the shaft is 200 KN

and speed is 75 rpm. Find: (i) Numbers of collars required (ii) Power lost in friction and

(iii) Heat generated at the bearing in KJ/min. Take μ = 0.05 and bearing pressure = 0.3

N/mm2.

5. A foot step bearing supports a shaft of 100 mm diameter which is counter-bored at the

end with a hole diameter of 50 mm. The bearing pressure is limited to 1 N/mm2. The

speed of shaft is 100 rpm. Assume µ = 0.015. Find: (i) the load to be supported (ii) the

power lost in friction (iii) the heat generated in bearing.

6. Following data is given for a 360° hydrodynamic bearing: Length to diameter ratio = 1,

Journal speed = 1350 rpm, Journal diameter = 100 mm, Diametral clearance = 100 µm,

External load = 9 kN, The value of minimum film thickness variable = 0.3, S = 0.0828,

Find viscosity of oil that need be used.

7. The following data refer to 3600 hydrodynamic journal bearing:

Journal speed = 900 rpm, End leakage factor = 0.002, Journal diameter = 50 mm,

Bearing length = 100 mm, Clearance ratio (c/d) = 0.001, bearing pressure = 1.4 N/mm2,

Absolute Viscosity of lubricant = 0.011 kg/m-sec at 750C operating temperature, Room

temperature = 350C, Inlet temperature of the oil = 100C, Specific heat of the oil = 1850

J/k/0C, Heat dissipation Coefficient = 280W/m2/0C, Calculate: (i) the amount of artificial

cooling required (ii) the mass of the lubricating oil required.

8. A 80 mm long journal bearing supports a load 2800 N on a 50 mm diameter shaft. The

bearing has a radial clearance of 0.05 mm and the viscosity of the oil is 0.021 kg/m-s at



the operating temperature. If the bearing is capable of dissipating 80 J/s, determine the

maximum safe speed.

9. Design a journal bearing from the following data:

Radial load = 20 kN, diameter of journal = 100 mm, Speed of journal = 900 r.p.m., oil

SAE 10 with viscosity at 550 C = 0.017 kg/m-sec, ambient temperature =15.50 C,

maximum bearing pressure = 1.5 MPa, permissible rise in oil temperature = 100 C, heat

dissipation coefficient = 1232 W/m2/0C, L/D ratio = 1.6, Design parameter µN/p = 28,

clearance ratio (c/d) = 0.0013, specific heat of oil = 1900 J/kg/0C.

10. The load on the journal bearing is 150 kN due to turbine shaft of 300mm diameter

running at 1800 r.p.m. determine the following:

i. Length of the bearing if the allowable bearing pressure is 1.6 N/mm2, and

ii. Amount of heat to be removed by the lubricant per minute if the bearing

temperature is 600C and viscosity of the oil at 600C is 0.02 kg/m-s and bearing

clearance is 0.25 mm.

11. A Petroff’s sleeve bearing consists of a sleeve having a bore diameter of 100.1 mm and a

length of 100 mm. A shaft having 100 mm diameter supports a load of 4000 N. A shaft

runs at 2880 r.p.m in the sleeve. If the frictional torque on the shaft is 10 N-m, find

i) The absolute viscosity of lubrication

ii) The bearing pressure

iii) The coefficient of friction and

iv) The power lost in bearing.



ASSIGNMENT – 7 SELECTION OF ROLLING CONTACT BEARING

Theory

1. Explain the difference between sliding contact bearing and rolling contact bearing.

2. Explain the static load capacity, dynamic load capacity and equivalent dynamic load

capacity.

3. Discuss causes of failure of antifriction bearing.

4. Explain the factors affecting selection of antifriction bearings.

5. Explain in detail the selection procedure of rolling contact bearing from manufacturer's

catalogue.

6. Discuss bearing seals & lubrication of rolling contact bearings.

7. Explain the bearing materials in detail.

Examples

1. Design a self-aligning ball bearing for a radial load of 7000 N and a thrust load of 2100

N. The desired life of the bearing is 160 mr at 300 rpm. Assume uniform and steady

load. The value of X and Y factors are 0.65 and 3.5 respectively. The outer ring rotates.

2. It is required to select a ball bearing suitable for 50 mm diameter shaft rotating at 1500

rpm. The radial and thrust loads at the bearing are 4500 N and 1600 N respectively. The

value of X and Y factors are 0.56 and 1.2 respectively. Select a proper ball bearing from

following table for rotating life of 22500 hr. Inner ring rotates and service factor is 1.

Bearing No. 6010 6210 6310 6410

C (N) 21600 35100 61800 87100

3. A single - row deep groove ball bearing No. 6002 is subjected to an axial thrust of 1000

N and a radial load of 2200 N. Find the expected life that 50% of the bearings will

complete under this condition. [Static load capacity CO: 2500 N, Dynamic Load Capacity

C: 5590N].

4. Single row deep groove ball bearing 6010 is subjected to an axial trust of 1200N and

radial load 2400 N. Find the expected life that 50% of the bearing will complete under

this condition. CO= 13200 N, C = 21600 N.

Fa/ CO Fa/Fr > e e

Fa/Fr < e

X Y X Y

0.07 0.56 1.6 0.27 1 0

0.13 0.56 1.4 0.31 1 0



5. For SKF 6207 bearing is to operate on following work cycle.

Radial load of 6307 N at 200 rpm for 25% of time



The inner ring rotates. The loads are steady. Find expected average life of this

bearing in hours if C = 25500 N.

6. A ball bearing is operating on a work cycle consisting of three parts:

a radial load of 3000 N at 1440 rpm for one quarter cycle,

a radial load of 5000 N at 720 rpm for one half cycle and

a radial load of 2500 N at 1440 rpm for remaining cycle.

The expected life of the bearing is 10000 hours. Calculate the dynamic load carrying

capacity of the bearing.

7. The following data refers to ball bearing work cycle:

Sr. no

Radial load

(N)

Axial load

(N)

Radial factor

Thrust factor

% time

Service factor

Speed (r.p.m.)

1 4000 800 1 0 30 % 1.25 900 2 8000 3000 0.56 2 40 % 1 600 3 - - - - 30 % - 600

Calculate the dynamic load rating of the bearing, if the expected bearing life is 10000

hrs with reliability of 95 %.

8. A ball bearing, subjected to a radial load of 5 KN, is expected to have a life of 8000 hrs at

1450 rpm with a reliability of 99%. Calculate the dynamic load capacity of bearing, so

that it can be selected from manufacturer’s catalogue based on a reliability of 90%.

9. A single row deep groove ball bearing is subjected to a radial load of 8000 N and a

thrust load of 3000 N. The values of X and Y factors are 0.56 and 1.5 respectively. The

shaft speed is 1200 rpm and diameter of shaft is 75 mm. The bearing selected for this

application is No.6315 (C = 112000 N). Find the life of the bearing with 90 % reliability

and estimate the reliability for 20000 hr life.


B.E. Semester VII Department of Mechanical Engineering Darshan Institute of Engineering and Technology, Rajkot 1

ASSIGNMENT - 8 Design OF I .C. Engine Components

Theory

1. What is the function of a valve gear mechanism? Explain with neat sketches the

valve gear mechanism in case of a vertical and horizontal engine. (V. B. Bhandari –

25.28)

2. Draw a neat diagram of the piston and explain the design requirements of piston in

engine. (V. B. Bhandari – 25.8)

3. Explain the piston materials. What are the advantages and disadvantages of

aluminium piston over cast iron piston? (V. B. Bhandari – 25.9)

4. What are the different types of piston rings? State its functions. Why is more number

of thin piston rings preferred over small number of thick rings? (V. B. Bhandari –

25.12)

5. What are the merits and demerits of wet and dry cylinder liners? (P. C. Sharma –

27.1)

6. Explain the design of cylinder of an I. C. engine.

7. Discuss about various section of connecting rod. Why ‘I’ section is chosen for the

connecting rod in the design of I. C. Engine? (V. B. Bhandari – 25.17)

8. Draw a neat sketch of the connecting rod and explain its constructional detail. (V. B.

Bhandari – 25.16)

Examples

1. The cylinder of a four stroke diesel engine has the following specifications: Brake power = 7.5 kW; Speed = 1400 rpm; Maximum gas pressure = 3.5 MPa; Indicated mean effective pressure = 0.35 MPa; Mechanical efficiency = 80 %. The cylinder liner and head are made of grey cast iron (Sut = 260 MPa and μ = 0.25). The factor of safety for all parts is 6. Calculate: a) Bore and length of the cylinder liner b) Thickness of the cylinder liner (Take, C= 3.2 mm) c) Thickness of the cylinder head

2. Following data refer to a 4-stroke petrol engine: Brake power=7.5 KW, Indicated mean effective pressure=0.45 N/mm2, Maximum explosion pressure=3.2 N/mm2, Mechanical efficiency = 80%, Allowable stress for C.I. cylinder = 40 MPa, Allowable stress for Ni-steel bolt = 70 N/mm2. Find:



a) Bore and Stroke of engine taking L/D=1.25, b) Thickness of cylinder wall and flange, c) Size and number of bolts required to join the cylinder head.

3. Determine the principle dimensions of cylinder for a vertical 4 stroke compression ignition engine from the following data: Brake power = 4.5 kW, Speed = 1200 rpm, Indicated mean effective pressure = 0.35 MPa, Mechanical efficiency = 80%.

4. Design a cast iron piston for a single acting four stroke diesel engine for following data: Cylinder bore = 100 mm, Stroke = 125 mm, Pmax = 5.8 N/mm2, Pmean = 0.8 N/mm2, ηm = 85 %, Speed = 1500 rpm, Fuel consumption = 0.16 kg/BP/hr, H.C.V. = 40 x 103 kJ/kg, Constant C = 0.05, K = 46.6 W/m/°K, σ t (C.I.) = 30 N/mm2 For piston: μ = 0.1, pb = 0.45 N/mm2 For piston rings: pw = 0.030 N/mm2, σ t = 80 N/mm2 For piston pin: pb = 20 N/mm2, σ b = 120 N/mm2, τ = 60 N/mm2.

5. Design a cast iron piston for single acting four stroke engine for following specification. Cylinder Bore = 110 mm, Stroke = 130 mm, Maximum gas pressure = 5 N/mm2, Brake mean effective pressure = 0.5 N/mm2, Fuel consumption = 0.2 kg/kw/hr, Speed = 2000 rev/min. For C. I. permissible tensile stress is 40 N/mm2, HCV=41870 KJ/kg, K for C.I. = 46.6. Permissible tensile stress for piston ring is 100 N/mm2 and permissible tensile stress for pin is 150 N/mm2.

6. Find the thickness of a piston crown based on thermal considerations for 4 stroke engine with following specifications:

i. Engine speed = 1500 rpm ii. Piston diameter = 87 mm

iii. Length of stroke = 96 mm iv. Brake mean effective pressure = 0.7 N/mm2 v. BSFC = 0.26 kg/kw-h

vi. l/r ratio = 04 vii. Heat conducted through crown = 10% of heat generated during combustion

viii. Calorific value of fuel = 42 MJ/kg ix. Assume that the piston is made of aluminum alloy with thermal conductivity of

175 w/m °C and allowable temperature difference of 111° C.



7. Design a connecting rod for a 4 – stroke petrol engine from the following data: Cylinder bore = 100 mm Stroke length = 140 mm Engine speed = 1500 rpm Possible over speed of engine = 2500 rpm Maximum explosion pressure = 2.5 MPa Weight of reciprocating parts = 18.5 N Length of connecting rod = 315 mm Yield strength of connecting rod material = 320 MPa Factor of safety = 5 Permissible bearing pressure for big end = 12.5 MPa Permissible bearing pressure for small end = 15 MPa

8. Design a connecting rod for a petrol engine from the following data: Diameter of piston = 110 mm; Mass of reciprocating parts = 2 kg; Length of connecting rod = 325 mm; Stroke = 150 mm; Speed = 1500 rpm with possible over speed of 1850 rpm; Compression ratio = 4:1; Factor of safety = 4; Maximum explosion pressure = 5.5 MPa. Select suitable material and permissible stresses for it.

9. Design a connecting rod for a high speed diesel engine from the following data: Cylinder bore = 100 mm, Stroke = 120 mm, Maximum speed = 1800 rpm, Compression ratio = 18, Max. Explosion pressure = 5 MPa, Mass of reciprocating parts = 3.5 Kg, Length of connecting rod = 240 mm. If the connecting rod is made of drop forged steel, determine the size of I-section, size of small end bearing, big end bearing and bolts. Assume suitable stresses.

10. Design a plain carbon steel centre crankshaft for a single acting four stroke, single cylinder engine for the following data: Piston diameter = 250 mm; Stroke = 400 mm; Maximum combustion pressure = 2.5 MPa; Weight of the flywheel = 5 kg; Total belt pull = 100 N; Length of connecting rod = 950 mm. The flywheel is used as a pulley. When the crank has turned through 30° from top dead centre, the pressure on the piston is 1 MPa and the torque on the crank is maximum. Any other data required for the design may be assumed.

11. Design the various components of a valve gear mechanism for a horizontal gas engine with the following data: Diameter of port is 70 mm, its weight is 5 N, and its lift is 25 mm.



The maximum combustion pressure is 4.5 MPa. The valve opens 33° before O.D.C. and closes 1° after I.D.C. and it is to open with constant acceleration and deceleration for each half of the lift. The gas pressure in cylinder when the exhaust valve start to opens is 0.34 N/mm2 , the pressure on the top side of the valve may be taken as 0.1 N/mm2 absolute and the greatest suction pressure is 0.035 N/mm2 below atmospheric. The engine runs at 350 rpm. The effective length of each arm of the rocker lever is 175 mm and the included angle is 140°.

12. Design a tubular type pushrod for operating an exhaust valve of 4 stroke I. C. engine using the following data:

i. Maximum force required to open the exhaust valve = 900 N ii. Ratio of lp/lw for rocker arm = 1.2

iii. Length of push rod = 110 mm iv. Ratio of inner diameter to outer diameter for tubular push rod = 0.75 v. Required factor of safety = 04

vi. Compressive yield strength for mild steel push rod = 350 N/mm2



ASSIGNMENT – 9 DESIGN OF CRANES

Theory

1. Explain design procedure of wire rope drum with necessary equations.

2. What do you understand by 6 x 37 ropes? Explain with neat sketch the different rope

section?

3. Give the detail classification of material handling equipments.

4. What are the basic principles in selecting the type of material handling equipments?

5. Why trapezoidal section is used in hook? Draw neat sketch of single hook and also

mention its critical section.

6. Write basic objectives of material handling system.

Examples

1. Design a single rope drum to transmit a torque of 8 kNm with a 32 mm rope. Assume

the height of the load to be raised as 2.7 meter and the ratio of the pulley system as 2.

The mean diameter of the drum is 576 mm. Assume the drum to be made of Grey cast

iron, grade 20 having allowable shear strength of 33 MPa.

2. Design a crane hook for lifting capacity of 5 tonnes. It is made from forged steel and has

triangular section. Take permissible tensile stress 80 N/mm2. Use data design book to

standardize the dimension of hook.

3. A single point hook is made from a 50 mm M.S. bar with 84 mm bed diameter. Calculate

the safe load that can be taken by this hook, if the design permissible stress is limited to

160 MPa. If the hook section is changed to trapezoidal section from triangular section

for the same bar what will be the change in load carrying capacity?

4. Select the ropes, pulleys and drum for an overhead travelling crane with a lifting

magnet. Lifting capacity = 4500 kg (mass), Weight of lifting magnet = 210 kg (mass),

Weight of lifting tackle = 110 kg (mass), Lifting height = 8.5 m, No. of rope parts = 4.

darshan institute of engg. & tech. · 1. two shafts at right to each other are connected by a...

Documents