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DESIGN, FABRICATION AND PERFORMANCE EVALUATION OF GARDEN TILLER

ABSTRACT
In order to meet the food requirements of the growing population and rapid industrialization, modernization of agriculture is inescapable. Mechanization enables the conservation of inputs through precision in metering ensuring better distribution, reducing quantity needed for better response and prevention of losses or wastage of inputs applied. Mechanization reduces unit cost of production through higher productivity and input conservation.
The problem arises from the fact that workers available to work in farmlands is insufficient. And as a ¦result most of the fields are left uncultivated. The middle class fanners cannot bear the high cost of power tillers available in the market. Most middle class fanners have small land holdings and to buy a tiller for serving their need is uneconomical. Cost of power tillers available in the market are in terms of lakhs. Also that the operation of the available power tillers is very complex, they cannot be operated without the help of skilled men. Thus it adds to the cost for the need of hiring a skilled labour.
Owing to the problems mentioned above our motive was to develop a tiller based on diesel. Diesel engines have more dead weight compared to petrol engines of same power. Thus the garden tiller basically is a tiller running on diesel which could be used effectively in cultivation of tapioca, ginger, pulses etc. And that the cost of the tiller should be affordable to the middle class fanners. Thus our objective was to design and fabricate a tiller based on diesel that is easy to operate and should come with an affordable price so that middle class fanners would not find any problem buying them.

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PROJECT REPORT Submitted by ANOOP PAUL ANUSH MOHAN ARUN JACOB P NIKHIL VIJAYAN ROMY

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1.1 INTRODUCTION
1.1.1 Status of agricultural mechanization in India
Most of the developing countries of Asia have the problem of high population and low level of land productivity as compared to the developed nations. One of the main reasons for low productivity is insufficient power availability on the farms and low level of farm mechanization. This is especially true for India.
It is now realized the world over that in order to meet the food requirements of the growing population and rapid industrialization, modernization of agriculture is inescapable. It is said that on many farms, production suffers because of improper seedbed preparation and delayed sowing, harvesting and threshing. Mechanization enables the conservation of inputs through precision in metering ensuring better distribution, reducing quantity needed for better response and prevention of losses or wastage of inputs applied. Mechanization reduces unit cost of production through higher productivity and input conservation.
Agricultural implement and machinery program of the Government has been one of selective mechanization with a view to optimize the use of human, animal and other sources of power. In order to meet the requirements, steps were taken to increase availability of implements ,irrigation pumps ,tractors ,power tillers, combine harvesters and other power operated machines and also to increase the production and availability of improved animal drawn implements. Special emphasis was laid on the later as more than 70% of the farmers fall in small and, marginal category. Liberal credit has helped in acquiring new machines. For example, Faridkot district in Punjab recorded 137 tractors per thousand hectares in 1986-87 where as many of the districts in the country may not have a single tractor even today. The availability of farm power through mechanical means was estimated as 2.71 hp per hectare in Punjab in 1986-87 where as many states may not have one tenth of it.
It is generally said that mechanization of small farms is difficult .But Japan having average land holding even smaller than ours, with proper mechanization has led agriculture to great heights. In order to minimize the drudgery of small fanners ,to increase efficiency and save fanner's time for taking up additional /supplementary generating activities , the use of modern time saving machines/implements of appropriate size needed to be suitably promoted.
1.1.2 Level of Mechanization in India
Industrialized countries of the west and in the Asian sub-continent have achieved almost 100 % mechanization in agriculture. Among the developing countries even in China, South Korea and Pakistan are much ahead of India. Facts recorded from FAO Year Book 1990, indicating the density of tractors and combine harvesters per thousand hectare may be seen. The world average is 19.15 tractors per thousand hectare and the tractor density in India is only 5.59 per thousand hectare which is far below the international average.
In India, the introduction of improved implements was initiated in 1880, with the advent of the Department of agriculture. With the organization and expansion of the departments in the year 1905, steps were taken to accelerate the pace of introduction of improved farm implements. A modest beginning of mechanization was made by utilizing World War -II surplus machines by the Central tractor Organization in late 40s and early 50s
Manufactures of tractors and power tillers in India commenced in 60swhere as manufacture of engines and pumps started much earlier. Till mid 70s, part of demand of fann machines was met from indigenous sources. Though substantial units of agriculture machines have been introduced in the recent past, yet adoption had mostly been in the northern states and in the scattered pockets/areas where better irrigation facilities were available.
1.1.3 Research & Development System
The Indian Council of Agricultural Research (ICAR) is the main organization looking after all agricultural research, including agricultural implements and machinery. It coordinates a number of research projects with centers at different places in the country. Some of the State Governments have also facilitated in setting up of research organizations at state level. Each of the state has at least one agricultural university.
A research programme usually concentrates on the development of equipment suitable to a given farming conditions. The objective is to improve upon the performance of indigenous implements or develop a new implement that can either enhance labour productivity or appropriately mechanize the operation where a labour or power shortage hinders completing the task in time.
Major tractor manufactures have set-up their own R&D facilities with well equipped laboratory with well equipped test track. Eg- Escorts, Eicher, HMT, TAFE, PLT, Mahindra&Mahindra. How ever, the small scale industries hardly have any facility for research and development. Most of the items being manufactured by them have been adopted from the designs available within and outside the country.
1.1.4 Future Prospects
Technology in the developed countries has undergone sea change in recent years. Products being manufactured in India require a similar approach to provide more reliable machines in terms of economy in operation, comfort, safety, easy maintenance and higher efficiency. Turbo charging and supercharging of the engines have become quite common now a days in the developed countries. Similarly, synchromesh transmission system on agricultural machines has become a common feature.
Fluid couplings or turbo clutches are being incorporated to cushion both engine and transmission against shock load, jerking, vibration and reducing clutch wear. Monitoring and control systems are needed on machines to assist the operator by way of automation in control and information's on wheel slip, area covered, maintenance requirement etc. These developments are required for tractors, power tillers, combine harvesters, engines and other similar machines.
Indian Farm Machinery Industry has not made significant achievements in exports expect a small quantity of tractors .There fore tractor and farm machinery manufactures will have to strive for marketing in the world wide competition market to get reasonable market share in the exports.
CHAPTER 2 2.1 DESIGN OF GARDEN TILLER
2.1.1 Selection of engine
As per standards the torque required for the output shaft was taken as 75Nm at a speed of about 300 rpm. We studied the details of the engine available. We selected the diesel engine that could provide a torque of around lOONm.The engine selected was the GARUDA engine which had a rated power of 6.4hp @ 3600 rpm. We found that as the speed of the engine was increased above 2500 rpm large amount of emissions was obtained .Hence we decided to restrict the speed of the engine to be less than 2000 rpm and set the maximum acceleration as 2000rpm. This was achieved by controlling the fuel fo be supplied by acceleration. We know that as the speed of the engine is decreased the brake power output also decreases. We assume brake power output is proportional to the speed of the engine. Since we have set the maximum engine speed as 2000 rpm, the brake power output will be 2000*6.4/3600 = 3.56hp.But in actual practice the decrease in power will be a parabolic variation. Io order to take account of this factor we assume a service factor 1.2. Hence the output power of the engine will be = 3.56*1.2= 4.3hp.Thus we have selected a four stroke single cylinder diesel engine of GARUDA as the power source.
2.1.2 4-Stroke compression ignition engines
In 4-stroke cycle engine the cycle of operation of engine, the cycle of operation of engine is completed in 4-strokes of piston or two revolutions of the crankshaft. In CI engine high compression ratio is used. During suction stroke air is alone inducted. Due to high compression ratio, the temperature at the end of the compression stroke is sufficient to ignite the fuel which is injected in to the combustion chamber. In the CI engine a high pressure fuel pump and an injector is provided to inject fuel in to combustion chamber.
The ideal sequence of operation for the 4-stroke CI engine is as follows;
Suction stroke: Suction stroke 0-1 starts when the piston is at top dead centre and about to move downwards. The inlet valve is open at the time and the exhaust valve is closed. Due to the suction created by the motion of the piston towards bottom centre, the air is drawn in to the cylinder. At the end of suction stroke the inlet valve closes.
Compression stroke; The fresh air taken in to the cylinder during suction stroke is compressed by the return stroke of the piston 1-2.During this strokes both inlet and exhaust valves remain closed.
Expansion stroke OR Power stroke; Fuel is injected in the beginning of expansion stroke 2-3 The rate of injection is such that the combustion maintains the pressure constant. After the injection of fuel is over (i.e. after the fuel is cut-off) the products of combustion expand. Both valves remain closed during expansion stroke.
Exhaust stroke; at the end of the expansion stroke the exhaust valve opens, the inlet valve remaining closed, and the piston is moving from bottom dead centre sweeps out the burnt gases from the cylinder, stroke 4-0.
2.2 BELT DRIVES
The belts are used to transmit power from one shaft to another by means of pulleys which rotate at the same speed or different speeds. The amount of power transmitted depends on the following factors:
1. The velocity of the belt.
2. The tension under which the belt is placed on the pulleys.
3. The arc of contact between the belt is used.
It may be noted that
a. The shafts should be properly in line to ensure uniform tension across the belt
section.
b. The pulleys should not be so far apart as to cause the belt to weigh heavily on the
shafts, thus increasing the friction load on the bearings.
c. The pulleys should not be too close together, in order that the arc of contact on the
smaller pulley may be as large as possible.
d. A long belt tends to swing from side to side, causing the belt to run out of the
pulleys, which in turn develops crooked spots in the belt.
e. The tight side of the belt should be at the bottom, so that whatever sag is present on
the loose side will increase the arc of contact at the pulleys.
f. In order to obtain good results with flat belts, the maximum distance between the
shafts should not exceed 10 meters and the minimum should not be less than 3.5
times the diameter of the larger pulley.
2.2.1 Selection of a Belt Drive
Following are the various important factors upon which the selection of a belt drive depends:
1. Speed of the driving and driven shaft.
2. Speed reduction ratio.
3. Power to be transmitted.
4. Centre distance between the shafts.
5. Positive drive requirements.
6. Shafts layout.
7. Space available.
8. Service conditions.
2.2.2 Types of belt drives:
The belt drives are usually classified into following three groups:
1. Light drives. These are used to transmit small powers at belt speeds up to about lOm/s as in agricultural machines and small machine tools.
2. Medium drives. These are used to transmit medium powers at belt speeds over
3. 1 Om/s but up to 22m/s, as in machine tools
4. Heavy drives. These are used to transmit large powers at belt speeds above 22m/s as in compressors and generators.
2.2.3 Types of belts:
Though there are many types of belts used these days, yet the following are important:
1. Flat belts. It is mostly used in factories and workshops, where a moderate amount of power is to be transmitted, from one pulley to another when the two pulleys are not more than 8 meters apart.
2. V-belts. It is mostly used in factories and workshops, where a great amount of power is transmitted, from one pulley to another, when the two pulleys are very near to each other.
3. Circular belt or rope. It is mostly used in factories and workshops, where a great amount of power is transmitted, from one pulley to another, when the two pulleys are more than 8 meters apart.
If a huge amount of power is transmitted, then a single belt may not be sufficient, in such case, wide pulleys (for V-belts and circular belts) with a number of grooves are used. Then a belt in each groove is provided to transmit the required amount of power form one pulley to another.
2.2.4 Materials used for belt
1. Leather belts. The most important material for flat belt is leather. The best leather belts are made from 1.2 meters to 1.5 meters long strips cut from either side of the backbone of the top grade steer hides. The hair side of the leather is smoother and harder than the flesh side, but the flesh side is stronger. The fibre on the hair side are perpendicular to the surface and give more intimate contact between belt and pulley and places the greatest tensile strength of the belt section on the outside, where the tension is maximum as the belt passes over the pulley.
2. Cotton or fabric belts. Most of the fabric belts are made by folding convass or cotton duct to three or more layers (depending upon the thickness desired) and stitching together. These belts are woven also into a strip of the desired width and
thickness. They are impregnated with some filler like linseed oil in order to make the belt water proof and to prevent injury to the fibres. The cotton belts are cheaper and suitable in warm climates, in damp atmospheres and in exposed positions. Since the cotton belts require little attention, therefore these belts are mostly used in farm machinery, belt conveyors etc...
3. Rubber belt. Rubber belts are made of layers of fabric impregnated with rubber composition and have a thin layer of rubber on the faces. These belts are very flexible but are quickly destroyed if allowed to come into contact with heat, oil or grease. One of the principal advantages of these belts is that they may be easily made endless. These belts are found suitable for saw mills, paper mills where they are exposed to moisture.
4. Balata belts. These belts are similar to rubber belts except that balata gum is used in place of rubber. These belts are acid proof and water proof and it is not affected by animal oils or alkalis. The balata belts should not be at temperatures above 40° Celsius because at this temperature the balata begins to soften and becomes sticky. The strength of balata belts is 25% higher than rubber belts.
Advantages and disadvantages of V-belt drive over flat belt drive Advantages
1. The V-belt drive gives compactness due to small distance between centres of pulleys.
2. The drive is positive, because the slip between the belt and pulley groove is negligible.
3. Since the V-belts are made endless and there is no joint trouble, therefore the drive is smooth.
4. It provides longer life, 3-5 years.
5. It can be easily installed or removed.
6. The operation of belt and pulley is quiet.
7. The belts have the ability to cushion the shock when machines are started.
8. The high velocity ratio (maximum 10) may be obtained.
9. The wedging action of the belt in the groove gives high value of limiting ratio of tensions. Therefore the power transmitte4d by V- belts is more than flat belts for the same coefficient of friction, arc of contact and allowable tension in the belt.
10. The V-belt may be operated in either direction, with tight side of belt at the top or bottom. The centre line may be horizontal, vertical or inclined.
Disadvantages
1. The V-belt drive cannot be used with large centre distances, because of larger weight per unit length.
2. The V-belts are not as durable as flat belts.
3. The construction of pulleys for V -belts is more complicated than pulleys of flat belts.
4. Since the V-belts are subjected to certain amount of creep, therefore these are not suitable for constant speed applications such as synchronous machines and timing devices.
5. The belt life is greatly influenced with temperature changes, improper tension and mismatching of belt length.
6. The centrifugal tension prevents the use of V-belts at speeds below 5m/s and above 50m/s.
2.2.5 V-belts
The V-belts are made of fabric and cords moulded in rubber and covered with fabric and rubber a.) These belts are moulded to a trapezoidal shape and are made endless. These are particularly suitable for short drives. The included angle for the V-belts is usually from 30° to 40°. The power is transmitted by the wedging action between the belt and the V- groove in pulley and sheave. A clearance must be provided at the bottom of the groove, b.) In order to prevent touching of the bottom as it becomes narrower from wear. The V-belt drive maybe inclined at any angle with tight side either at top or bottom. In order to increase the power output, several V-belts maybe operated side by side. It may be noted that in multiple V-belt drive all the belts should stretch at the same rate so
that the load is equally divided between them. When one of the set of belts breaks, the entire set should be replaced at the same time. If only one belt is replaced the new unworn and unstretched belt will be more tightly stretched and will move with different velocity.
2.2.6 Types of V-belts and pulleys
According to Indian standards (IS: 2494 -1974), the V -belts are made in five types ie. A, B, C, D and E. The pulleys for V-belts maybe made of cast iron or pressed steel in order to reduce weight.
Table 2.1 Dimensions of standard V-belts according to IS 2494-1974
Type of belt Power ranges in kW Min. pitch dia. of pulley (D) mm Top width (b) mm Thickness (t) mm Weigh t/m length in N
A 0.7-3.5 75 13 8 1.06
B 2-15 125 17 11 1.89
C 7.5-75 200 22 14 3.43
D 20-150 355 32 19 5.96
E 30-350 500 38 23 -
2.3 DESIGN OF THE BELT DRIVE
Nl speed of engine shaft
N2 speed of intermediate shaft
D diameter of the pulley
9 lap angle
a area of the belt in mm2
m mass of the belt in kg/m
V speed of the belt in m/s
Tc centrifugal tension in N
T T,
T2
maximum tension in N
tension in the tight side of the belt in N
tension in the slack side of the belt in N
Velocity ratio for the drive = N1/N2 = D2/D1
= 2000 / 500 = 4
Assuming the diameter of the pulleys to be of 25cm and 6.25cm according to the standard dimensions.
The distance between the two pulleys x = 40cm
For an open belt drive,
Sina = (0.25-0.0625)0.4 = 0.04685
a = 27.953
Angle of lap on the smaller pulley
0 = 180-2a
= 124.093 degrees = 2.165 radians.
Area of the belt =130 mm2
Mass of the belt = 0.052kg/m
Angular velocity of the shaft
co = (2TI*2000)/60
= 209.439 rad/sec.
Speed of the belt
r*co
= 0.03125*209.439 = 6.544m/sec.
Centrifugal tension
Tc = m*v2
= 0.052*6.542
= 2.226N
Maximum tension in the belt
T =2.5*130
= 325N
Tension in the tight side of the belt
T, = T - Tc
= 325-2.226
= 322.774
We know that,
2.3 log(T!/T2) = uB cosec P
log(T,/T2) = (0.25*2.165*cosec 17)/2.3
T,/T2 = 6,380
Tension in the slack side of the belt
T2 = 322.774/6.380
= 50.591N
Power transmitted = (Tl -T2)* v
= (322.774- 50.591)* 6.544 = 1.781
No. of belts = 321/ 1.781 = 1.802 = 2
2.4 CHAIN DRIVES
A chain drive consists of an endless chain wrapped around two sprockets. The chain consists of a number of links connected by pin joints, while the sprockets are toothed wheels with a special profile for teeth. The chain drive is intermediate between the belt and gear drives. It has some features of gear drive and some features of belt drive. The advantages and disadvantages of the chain drive compared to the belt and gear drives are as follows:
1) Chain drives can be used for long as well as short distances. They are particularly suitable for medium centre distance, where gear drives will require additional idler gears.
2) A number of shafts can be driven in the same or opposite direction by means of the chain drive from a single driving sprocket.
3) Chain drives are used for shafts which are parallel, whereas some gears like bevel and worm gears can be used for non parallel shafts.
4) Chain drives are more compact than belt and gear drives.
5) A chain drive does not slip and to that extend, it is a positive drive compared to belt drive. However it is unsuitable where precise motion is required due to polygonal effect and wear in the joints. They also require adjustment for slack, such as a tensioning device.
6) Compared to belt drives chain drives require precise alignment of shaft. However, the center distance is not as critical as in case of gear drives.
7) The efficiency of chain drives is high at times as high as 98%.
8) Compared to belt drives chain drives requires proper maintenance,
particularly lubrication and slack adjustment. However, chains can be
easily replaced.
9) They are unaffected by fire hazards and atmospheric conditions.
2.4.1 Roller chains
Figure 2.2 construction of roller chain
There are five parts for a roller chain namely pin, bushing, roller and inner and outer page link plates. The pin is press fitted to two outer page link plates, while bushing is press fitted to the two inner page link plates. The bush and the pin form a swivel joint and the outer page link is free to swivel with respect to inner link. The rollers are freely fitted on bushes and, during engagement, turn with the teeth of the sprocket wheels. This results in rolling friction instead of sliding friction between the roller and the sprocket teeth and reduces wear. The pins bushes and rollers are made of alloy steels.
The pitch of the chain is defined as the linear distance between axes of adjacent rollers. Roller chains are standardized and are manufactured on the basis of pitch. These chains are available in single -row or multi- row constructions such as simple duplex or triplex strands.
2.4.2 Sprocket wheels
A wheel that drives or is driven by a chain is usually
Figure 2.3 Sprocket
referred to as a sprocket. Small sprockets up to 100mm in diameter are usually made of a disc or a solid disc with a hub on one side. They are machined from low carbon steel bars. Large sprockets with diameter more than 100mm diameter are either welded to steel hubs or bolted to cast iron hubs. In general for most applications the sprockets are made of low carbon or medium carbon steels and, on rare occasions, stainless steel is also used for making sprockets. When the chain velocity is less than 180m/min, the teeth of the sprocket wheels are heat-treated to obtain a hardness of 180 B.H.N. For high speed applications the hardness recommended is 300 to 500 B.H.N. The teeth are hardened either by carburizing in case of low carbon steels or by quenching and tempering in case of high carbon steels. The sprocket that we used is shown below.
2.5 DESIGN OF CHAIN DRIIVE FOR THE BLADE SHAFT
P power transmitted in watts
Z| no. of teeth on the smaller sprocket or pinion
Z2 no. of teeth in the larger sprocket or gear
p pitch of the chain in mm
WB breaking load of the chain in N
PCD pitch circle diameter in mm
v pitch line velocity in m/s
Power transmitted = 3210watts Velocity Ratio N1/N2 = 500/250 = 2
For roller chain, the number of teeth on the smaller sprocket or pinion (Zi) for a velocity ratio of 2 is 27.
Therefore number of teeth on the larger sprocket or gear, Z2 =Zi *(N1/N2) = 17*(500/250) = 54
Design Power = rated power * service factor
=3.2 *1.5 = 4.8kW
For the range 3-15 kW
Chain no. 10 B should be used.
For chain 10B
Pitch = 15.875 mm
Breaking load = 22.2kN ¦
PCD for pinion di = p*cosec (180/Zi)
= 15.875*cosec (180/27)
= 136.74mm
PCD for sprocket = p*cosec( 18O/Z2)
= 15.875* cosec (180/54) = 273.02mm
Pitch line velocity on smaller sprocket,vl = (7r*d!.Ni)/60 = (7t*0.136*500)/60 = 3.56m/sec
Load on the chain, W
= 3.210/3.56 = 901.68N
2.6 SHAFT
The term transmission shaft is usually referred to a rotating machine element, circular in cross section, which supports transmission elements like gears, pulleys and sprockets and transmits power. Such shafts are subjected to bending tensile, bending or torsional shear stresses, or to a combination of these. The design of transmission consists of determining the correct shaft diameter from strength and rigidity considerations. The materials usually used are mild steel and alloy steels, such as nickel, nickel-chromium and molybdenum steels.
The shaft used here is subjected to a combination of both bending and torsional moments. The shaft material used is standard steel which is
tensile in nature and hence the principle shear stress theory of failure is used to determine the shaft diameter.
For a shaft subjected to a bending moment Mb, the bending stress at any fibre is given by
ob = (Mb*y) II
where,
ab ” Bending stress at a distance y from the
neutral axis in N/mm2.
Mb ” Applied bending moment in Nmm.
I ” moment of inertia of the cross section about the
neutral axis in mm4
For circular cross section
the moment of inertia, I = fld4/64.
Therefore, Ob
Maximum bending stress occurs at y = d/2.
= (32 Mb)/ (nd4). ” 1
torsional shear stress, x
Where, i
Mt r
in mm.
J
the axis of rotation in mm4.
For the shaft subjected to torsional moment, the = Mt*r/J.
torsional shear stress in N/mm . applied torque in Nmm.
radial distance of the fibre from the axis of rotation
polar moment of inertia of the cross section about
inertia is given by,
For a circular cross section the polar moment of
Therefore, x
J = nd4/32.
= (16Mt)/(rid4).
The maximum shear stress on the shaft can be determined by constructing the Mohr's circle
Ob
Figure 2.4 Stresses acting on a body
Tmax =V {(0b/2)2+X2}.
Substituting the value of maximum shear and bending stress from above equations number 1 and 2, we get,
Tmax = 16/(IId3)* V {(Mb)2+(M,)2}
From the above equation the diameter of the solid shaft subjected to both bending and torsion,
d = [16/ (n*ximx) {(Mb)2 +(Mt)2}(a5)](1/3) -3
Applying the Numerical combined shock and fatigue factor to be applied to both computed twisting moment (Kt) and bending moment (Kb) and taking the maximum shear stress as the design shear stress value, xecj we get the diameter of the shaft as,
d = [16/ (FI*fs) {(KbMb)z + (KM)'}
2.6.1 Design of intermediate shaft
60N
r \902N
C A Ra D B
11 Wm Wm
¦* In.

Figure 2.6 Free body diagram of the intermediate shaft
Ti tension on the tight side of the pulley in N
T2 tension on the slack side of the pulley in N
|i coefficient of friction
9 lap angle
N rpm
P power transmitted in watts
T torque to be transmitted in Nm
FT chain load
FTH horizontal component of the chain load
FTV vertical component of the chain load
M design bending moment in Nm
Te Torque effective
fs shear stress in N/mm2
We know that,
T,/T2
2.193 T2
Power transmitted
P
T
= (2TI*N*T) / 60
= (3210 * 60)/ (2r* 2000)
15.326Nm
(Ti-T2) * 0.125 = 15.326 From 1
(2.193 T2-T2)*0.125 = 15.326 1.193 T2*0.125 =15.326 T2 = 102.772 N Ti = 122.606N
Chain load
FT =3210/3.56
901.68N
FT cos 60
901.68*cos 60
450.93
FT sin 60
781.03N
Bending moments
Horizontal
MA = 22.537Nm
MC = 33.824Nm
Vertical
MA = 6Nm
MC = 75.10Nm
Resultant bending moments
Res MA = (22.532+62)l/2
= 23.315Nm
Res MC = (33.8242+75.1032)1/2
= 82.368Nm
Torque effective, Te = (M2+T2)1/2
= (82.3 6 82+15.3622)1/2 = 83.788Nm
Kt*Te =(7t/16)*fs*d3
1.5*83.788 = (7i/16)*44*106*d3 d = 24.41mm
Nearest standard size available = 25mm
Figure 2.7 Free body diagram of the blade shaft
Ra reaction at the left side bearing, in N
Rb reaction at the right side bearing, in N
P power transmitted in Watts
N speed of the shaft in revolutions per second
T torque to be transmitted in Nm
d diameter of the shaft, in mm
FT chain load in N
FTV vertical component of chain load
FJH horizontal component of chain load
Te torque effective in Nm
fs shear stress in N/mm2
M bending moment in Nm
d diameter of shaft in mm
Chain load acting on the shaft FT = 902 N
FTV = FT*sin30 =902*sin30 =45 IN
FTH =FT *COS30 =902*COS30 =781.15
Power transmitted 3200
Torque
= (2TT*N*T)/60 = (2 7t*250*T)/60 = 3200*60/500 TC =122.23Nm
Bending moments Vertical
MA -106.2
MB =106.2
Horizontal
MA =156.3
MR =156.3
Resultant bending moment,M=( MA2+ MB2)1/2
=(106.22+156.32)1/2 =188.96Nm
Torque effective, Te
=(M2+T2)1/2
=(188.962+1232)1/2
=225.465Nm
Kt*Te
= (Tt/16)*fs*d3
1.875*225.465 =( n/16)*42*106*d3
d =37.14mm
Nearest standard size available = 40mm
2.7 BEARINGS
A bearing is a machine part whose function is to support a moving element and to guide or confine its motion, while preventing the motion in the direction of applied load. They take up the radial and axial loads imposed on the shaft or axle they carry, and transmit these to the casing or machine frame.
2.7.1 Classification
Bearing can be classified in the following many ways:
1) Depending up on the direction of load to be supported:
a) Radial bearing
b) Thrust bearing.
2) Depending up on the nature of contact between the working surfaces:
a) Sliding contact bearing
b) Rolling contact bearing.
3) Depending up on the type of loading:
a) Bearing with steady load
b) Bearing with variable or fluctuating load.
2.7.2 Selection of bearing type
There is no hard or fast rule or formula existing, for deciding between sliding and rolling bearings. For this, the performance of each bearing must be compared in relation to load capacity, friction, space requirements, accuracy, noise etc
2.7.3 Bearing selection based on mechanical requirements
2.7.4 Sliding bearings or plain bearings
In sliding bearings the primary motion between the bearing and the moving element is sliding one. The sliding bearings can be classified in to two groups depending up on the nature of motion of the moving element.
They are 1) linear bearings: - In these bearings sliding action is guided in a straight line.
2) Sliding bearing with motion of rotation: - In these bearing, the sliding motion between the bearing and the moving element, is he motion of rotation.
Sliding bearings are also classified according to the type of friction present between the bearing and the moving element:
1) Dry friction bearing: No lubricant is supplied in between the rubbing surface and there is metal contact. The coefficient of friction may range fromO.l to 0.25.
2) Boundary friction bearings: - In these bearings, the lubricant is supplied to the rubbing surfaces in scarce quantities. The lubricant is neither constant nor abundant .Such bearings are suitable for low load and low speed conditions since at increased loads , the thin film of lubricant will break and the bearing surfaces approach each other, resulting in metal to metal contact and wear the surfaces.
3) Semi fluid friction: - This is the transition between the boundary and fluid type of friction. Lubricant is supplied constantly but not abundantly and so the lubricant film is very thin and when it breaks, there is metal to metal contact. The coefficient of friction range from 0.005 to 0.10.
4) Bearing with fluid friction: - In these bearings there is always a thick film of lubricant between the working surfaces and there will never be a metal to metal contact.
2.7.5 Rolling contact bearings
In these bearings the contact between the bearing elements is rolling instead of sliding as in plain bearings. Since rolling friction is very less as compared to the sliding friction, such bearings are also known as "antifriction bearings".
2.7.5.1 Types of rolling contact bearings
Following are the two types of rolling contact bearings
1) Ball bearings
2) Roller bearings
The ball and roller bearings consists of an inner race which is mounted on the shaft or journal and an outer race which is carried by the housing or casing. In between the inner and outer race, there are balls or rollers. A number of balls or rollers are used and these are held at proper distances by retainers so that they do not touch each other. The retainers are thin strips and are usually in two parts which are assembled after the balls have been properly spaced. The ball bearings are used for light loads and the roller bearings are used for heavier loads.
The roller bearings depending up on the load to be carried are classified as
1) Radial bearings
2) Thrust bearings.
2.7.6 Types of radial ball bearings
1) Single row deep groove bearings: A single row deep groove bearings are used due to their high load carrying capacity and suitability for high running speeds. The load carrying capacity of a ball bearing is related to the size and number of the balls. This bearing is usually made with deep groove. This is most widely used type of ball bearing.
2) Filling notch bearing: These bearings have notches in the inner and outer races which permits more balls to be inserted than deep groove ball bearing.
3) Angular contact bearing: These bearings have one side of the outer race cut away to permit the insertion of more balls than in deep groove bearing but without having a notch cut in to both races.
4) Double row bearing: These bearings may be made with radial or angular contact between the balls and races. The double row bearing is narrower than two single row bearings.
5) Self aligning bearings: These bearings permit shaft deflection with in 2-3 degrees. Following are the two types of self aligning bearings
1) Externally self aligning bearings
2) Internally self aligning bearings.
Where,
D-Outside diameter d-Bore diameter W-Width
These dimensions are functions of the bearing bore and the series of bearing. The standard dimensions are given in millimeters. There is no standard for the size and number of steel balls.
The bearings are designated by a number. In general the number consists of at least three digits .Additional letters or digits are used to indicate special features. The last three digits give the series and bore of bearing. The last two digits from 04 onwards, when multiplied by 5, give the bore diameter in millimeters. The third from the last digit designates the series of the bearing. The most common ball bearings are available in 4 series as follows;
1) Extra light (100)
2) Light (200)
3) Medium (300)
4) Heavy (400)
2.8 DESIGN OF BEARINGS 2.8.1 Bearing B
We have selected deep groove ball bearings since deep groove ball bearings have high radial load capacity. The shaft has the diameter of 25 mm.
60N \902N

C A Ra D B Rb
r4
”<
Figure 2.9 Free body diagram of the intermediate shaft
L bearing life in millions of revolutions
Co static load capacity in N
C dynamic load capacity in N
N rotation of the shaft in rpm
LH bearing life in hours
X radial thrust factor
V race rotation factor
Y axial thrust factor
Fr radial load
Fa axial load
E clearance
P equivalent dynamic load
D outer race diameter of bearing
B width of bearing
The required life of bearing in millions of revolution is given by L= ((60*n*Lž) /10a6)
Where,
n= 500 rpm.
From table 24.40, typical values of bearing life for various applications. LH=600hrs, for light vehicle Therefore L= ((60*500*600)/10a6)
=18 millions of revolutions
The radial load acting on the bearing,(Rb),Fr =865N.
Since the shaft is horizontal the axial load acting on the bearing is assumed to be zero i.e., Fa=0 N.
From table 24.58,
Index FL of dynamic stressing varies from 1.4-1.9.
Here we take FL =1.4
Therefore radial load, Fr=865*1.4
=1211N
From table 24.60, deep groove ball bearing -diameter series-2, (SKF)
Assume the bearing number is 6205.
Then, d=25mm. D=52mm. B=15mm. C0=6965 N. C =10690N e= Fa/Co
=0/6965
=0
From table 24.47 b,
Factor Fn, for deep groove ball bearing is 13.8
Therefore (Fž.Fa/C0)= 13.8*0/10690 =0
Since there is no axial load is acting on the bearings, we take X=l and Y=0
The equivalent dynamic load is given by
P=X*V*Fr +Y*Fa
Where
X=l
V=l, for all types of bearing if the inner race is rotating Fa=0
Substituting the values in the above equation, we get =1*1211 P=1211 N
*Check
The bearing life in millions of revolutions,
L=(C/P) aP
C=10690 N
p=3, for ball bearings
P=1211N
L=(10690/1211)a3 =687 millions of revolutions
Since life of selected bearing is more than required life, the bearing suitable for the purpose.
So we select the bearing 6205
2.8.2 Bearing A
We have selected deep groove ball bearing, since the radial load acing on the bearing is high. The shaft has the diameter of 35 mm.
MO
C A Ra E B Rb 1
Hal- 4

Figure 2.10 Free body diagram of the blade shaft
L bearing life in millions of revolutions
Co static load capacity in N
C dynamic load capacity in N
N rotation of the shaft in rpm
LH bearing life in hours
X radial thrust factor
V race rotation factor
Y axial thrust factor
F, radial load
Fa axial load
E clearance
P equivalent dynamic load
D outer race diameter of bearing
B width of bearing
The required life of bearing in millions of revolution is given by L= ((60*n*Lž) /10a6)
Where,
n=250 rpm.
From table 24.40, typical values of bearing life for various applications.
Ln=600hrs, for light vehicle
Therefore, life of bearings in millions of revolutions, L= ((60*250*600)/10A6) =9 millions of revolutions
The radial load acting on the bearing, (Rb),Fr =3500N.
Since the shaft is horizontal the axial load acting on the bearing is assumed to be i.e., Fa=0 N.
From table 24.58,
Index FL of dynamic stressing varies from 1.4-1.9. Here we take FL =1.4
Therefore radial load, FR= 3500*1.4
=4900 N
From table 24.60, deep groove ball bearing -diameter series-2, (SKF)
Assume the bearing number is 6208, since we start from light series.
Then, d=40mm. D=80mm. B=18mm. C0=15495 N. C =22165N
e= Fa/Co =0/15495 =0
From table 24.47 b,
Factor F0 ,for deep groove ball bearing is 13.8
Therefore (F0.Fa/C0)= 13.8*0/15495 =0
Since , there is no axial load is acting on the bearings ,we take X=l and Y=0
The equivalent dynamic load is given by
P=X*V*Fr +Y*Fa
Where
X=l
V=l, for all types of bearing if the inner race is rotating.
Fa=0
Substituting the values in the above equation, we get =1*4900
Therefore equivalent dynamic load, P =4900 N
*Check
The bearing life in millions of revolutions, L=(C/P) aP
Dynamic load carrying capacity C=22165 N
p=3, for ball bearings
P=4900N
L=(22165/4900)a3 = 93 millions of revolutions
Since life of selected bearing is more than required life, the bearing suitable for the purpose.
So we select the bearing 6208
2.9 SPECIFICATION OF BLADES FOR ROTAVATOR OF POWER TILLER
Blade is an important soil engaging component of the rotavator. It wears out earlier than other components causing its replacement very often. For the purpose of deciding whether a particular requirement of this standard is compiled with , the final value ,observed or calculated, expressing the result of a test or analysis ,shall be rounded off in accordance with IS : 2-1960. The number of significant places retained in the rounded off value should be the same as that of the specified value in this standard.
2.9.1 Scope
This standard specifies material, hardness, dimensions and other requirements for blades used in rotavators operated by power tillers.
2.9.2 Types
Blades shall be of the following types :
a) Type A- Straight blade.
b) Type B- Hatchet blade.
The blade that we are using is the hatchet blade.
Figure 2.11 Hatchet blade
2.9.3 Materials
The chemical composition of the steels to be used for the manufacture of blades shall be as follows.
a) Carbon steel:
Carbon 0.70 to 0.85 %
Silicon 0.10 to 0.40%
Manganese 0.50 to 1.0 %
Sulphur 0.05 % ,Max
Phosphorous 0.05 % ,Max
b) Silico Manganese Steel:
Carbon 0.50 to 0.60 %
Silicon 1.50 to 2.00%
Manganese 0.50 to 1.00%
Sulphur 0.05 %, Max
Phosphorous 0.05 %, Max
Some of the typical steels that may be used are T 70 Mn 65, T 75, T 80 Mn65 and 55 Si 2 Mn 90.
2.9.4. Hardness
The blades shall be heat treated, quenched and tempered. The hardness in edge portion shall be 56 ± 3 HRC and shank portion shall be 37 to 45 HRC.
2.9.5. Dimensions and tolerances
The essential dimensions of Type B blades are given below.
Table 2.3 Essential dimension of blades
SI no Description Dimension Tolerance
1 A 25.0,26.0 -0.3 -0.8
/ 2 I B 1 10.0 1 ±0.5 j
3 C 25.0 j ±0.5 j
4 D 10.5 +0.3 0.0
5 E 210OR225 ±0.5
6 E 240,245 ±0.5
All dimensions are in millimeters.
2.9.6. Workmanship and finish
The blades shall be free from cracks, scrams and other visual defects which may be detrimental for their use. The blades shall be free from rust and shall have a protective coating which will prevent surface deterioration in transit and storage.
2.9.7. Marketing and packing
Marketing - Each blade shall be marked with the following particulars at suitable place avoiding the soil facing side:
a) Manufacturer's name or recognized trade mark;
b) 'L' or 'R' in case of hatchet type blades to denote left or right
c) Batch or code number.
Each blade may also be marked with the ISI Certification Mark.
Packing - The blades of the same type shall be packed for safe handling in transit as agreed to between the purchaser and the supplier.
2.9.8. Sampling for lot acceptance
Unless other wise agreed to between the purchaser and the supplier, the sampling of blades for the lot acceptance shall be done in accordance with 3 of IS: 7201¬1974.
CHAPTER 3
3.1 FABRICATION OF GARDEN TILLER
We used the following components for the fabrication of the Garden tiller.
> GARUDA Engine
> Starter pulley
> Bearings
> Solid shaft
> Hollow shaft
> Circular disc
> Tilling blade
> Chains & sprockets
> Trolley wheel
> Belt drive
For the fabrication of the GARDEN TILLER we used a 550cc GARUDA diesel engine. It produces 4.7 hp @ 2000rpm. As per design considerations we had to reduce the rpm to around 250 rpm at the main shaft (ie: blades). For this purpose we used a chain and sprocket drive and a belt drive as speed reduction in single step with chain drive will be out of the allowable speed ratio.
The engine is initially bolted to a standard steel angle iron frame of 40cm x 75cm by means of M10 bolts. The power of the engine is transmitted to the intermediate shaft through a belt drive. The speed ratio being 4, 2000 rpm of the engine is reduced to 500 rpm. The pulleys are of 2.5 inch and 10 inch diameter. Considering the power transmitted B type V43elts are used for the drive and two belts have been used to accommodate the power.
Figure 3.2 intermediate shaft The drive is now up to the intermediate shaft and for further transmission to the tilling shaft a chain drive is used. The shaft was designed to the specifications and the shaft was machined. The tilling shaft is 110cm long and of 40mm diameter. The shaft is supported at a distance of 35cm from both ends and the centre part of the shaft was made 45mm for creating the step for the bearings. This shaft carries a sprocket of 54 teeth and 8 pairs of tilling blades. The sprocket is keyed to the shaft. The sprockets on the two shafts are connected using a 10B chain. The bearings (6208) are housed in a MS plate of 0.5 inch thickness which acts as the support for the shaft to the frame. The tilling blades are attached using M8 bolts to MS circular discs of 6 inch diameter. The blades are attached in pairs of left and right orientations. A circular disc has 2 pairs of blades, the blades attached diametrically opposite on the disc. The circular discs are welded to the shafts equal intervals leaving the 40cm length at the centre. Thus the tilling shaft has 4 tilling sections with 4 blades each. The MS plate supports are fixed to a MS plate of 6mm thickness using M8 bolts. This MS plate is attached to the angle iron frame using M10 bolts and welded for extra strength.
A height adjusting wheel was fixed to the tilling shaft for adjusting the depth of blade on soil while working. Two wheels are attached to the front of the frame for the mobility of the tiller.
A mechanism for tightening the belts over the pulley is made using a 2B-type pulley. A handle is made out of a GI pipe and is fixed on to the angle iron frame. The accelerator switch is fixed on to the handle for easy manipulation of the engine power as required. The whole assembly was checked for any loose or faulty weldings. The assembly was painted and put to test.
3.2 COMPONENTS AND COST TABLE
CHAPTER 4
4.1 CONCLUSION AND FUTURE SCOPE 4.1.1Conclusion
The model of the garden tiller was constructed successfully and was found to work as per the requirements. The garden tiller that we developed can be used for many activities including the cultivation of tapioca, pulses ginger, turmeric etc. This can be achieved by using special attachments some of which are available with KAMCO. The cost of the garden tiller that we developed is around Rs.25000, whereas the cost of the power tiller is in terms of lakhs. Thus the equipment that we developed will be accessible to middle class farmers who are in deep crisis due to the unavailability of sufficient labor for working in farmland. The big scale farmers could only bear the costly equipments used in farmlands that have very specialized purpose. Thus this multipurpose equipment would be a boon to the small scale farmers.
4.1.2 Future scope
Providing of power to the wheels can greatly improve the mobility of the device. This will ease the effort of the worker.
Additional accessories can be incorporated which helps us to use the device for various applications. In order to improve safety we can use flap to prevent mud from striking the worker.
CHAPTER 5
5.1 REFERENCES
1. "Garden tillers'*, "Barbee, Jesse J., "1986"
2. American chain association, "chains for power transmission and material handling"(Marcel Dekker)
3. "Powered soil tillage device", by"Zach, Lawrence J.; Kosch, Alois J.;", "2005"
4. Harris TA., rolling bearing analysis, John Wiley, 1967
5. kamcoindia.com
6. chain-guide.com
7. lawn-aerator-attachment.com
8. cleanairgardening.com
9. agricoop.nic.in
10. Machine design data hand book data - K. Lingaiyah, Vol l,Suma publishers
11. Design of machine elements - V.B. Bhandari,Tata mac-graw hill publishers
12. Machine design - Robert L Norton, Pearson Education