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Department of Agricultural and Food Department of Agricultural and Food EngineeringEngineering
Indian Institute of Technology, KharagpurIndian Institute of Technology, Kharagpur
Design and Development of Hydrostatic Design and Development of Hydrostatic Transmission for Agricultural TractorTransmission for Agricultural Tractor
ByTanaji Shinde
09AG6103
Under the Supervision of
Prof. V. K. TewariProf. V. K. Tewari
Professor, Farm Machinery and Power, Professor, Farm Machinery and Power, Agricultural & Food Engineering Agricultural & Food Engineering
Department, IIT KharagpurDepartment, IIT Kharagpur
Mr. R. Murli KrishnanMr. R. Murli Krishnan
Associate Vice PresidentAssociate Vice President
R & D Division, TAFE Ltd. R & D Division, TAFE Ltd. ChennaiChennai
Hydrostatic transmission is one of the type of continuously variable transmission
A continuously variable transmission (CVT) is a transmission which can change steep lessly through an infinite number of effective gear ratio between maximum and minimum values
This can provide better fuel economy than other transmission by inability the engine to run at its most efficient rpm for a range of vehicle speeds
Introduction
Fig. Torque-speed curves of (a) an engine and (b) vehicle rear axle.
Justification of work
The greatest advantage of a hydrostatic transmission
is the ability to infinitely vary the ground speed .
Transmit high power in compact size.
Operates efficiently over a wide range of torque to
speed ratio.
Maintains controlled speed (even in reverse
direction) regardless of load within the design limit
Maintains a preset speed accurately against driving
or braking loads.
Advantages of hydrostatic transmission system
1. Design and development of hydrostatic transmission system for agricultural tractor.
2. To evaluate the performance of the developed hydrostatic transmission system in laboratory condition
OBJECTIVES
How does Hydrostatic transmission system work?
A hydrostatic CVT is formed by the combination of at least one hydrostatic pump and at least one motor.
A hydrostatic transmission consists of a pump hydraulically coupled to a hydraulic motor
Power input is in to pump, while power output is from motor
We can operate pump from engine which converts mechanical power in to hydraulic power and motor converts hydraulic power in to mechanical power.
Speed control lever regulates the position of the
swash plate . As the control handle is moved
forward to increase speed due to increase
hydraulic flow to the hydraulic motor. This
increases the speed of the motor
Design of hydrostatic transmission system for agricultural tractor.
1 . Sizing of hydrostatic transmission. a) Determination of maximum tractive effort . b) Selection of Hydraulic pump and Hydraulic
Motor. c) Determination of Gear reduction d) Hydraulic circuit of Hydrostatic Transmission. e) Sizing of charge pump , Reservoir f) Selection of Hydraulic hoses, Hydraulic Fluid.
Objective -I
Wheel
Engine Hydrostatic transmission
Planetary gear box with 4:1, 1:1 reduction
Differential with
3.4545 reduction
Final drive 3.143 reduction
Final drive 3.143 reduction
Wheel
Fig. Block diagram of proposed drive train of hydrostatic transmission for tractor
1) Type of vehicle : 2 wheel drive agricultural tractor equipped with pneumatic tires
2) Size of prime mover : 50 hp engine at 2250 rpm of governed speed and 700 at low idle
3) Loaded radius
: 640 mm
4) Gross vehicle weight : 2090 kg. 5) Pump drive : Direct from engine. 6) Gear reduction from motor to wheel:
Gear reduction with planetary gear box
Differential unit
Final drive
:43.429, 10.8574
: 4:1, 1:1
: 3.4545
: 3.143
Following are the parameters required for sizing of hydrostatic transmission system
Sizing of hydrostatic transmission
YES
NO
Calculate Machine Corner power
Calculate Transmission Ratio ( TR)
TR> 4
Used Fixed Motor
Calculate Required Motor corner power
Select Motor Size DM
Calculate Motor Corner Power (CP)
Use variable Motor
A
(1)
(2)
(3)
(4)
Determine maximum Tractive Effort
Flow Chart for Selection of Hydraulic Pump and Motor
YES
Motor Cp>=required motor cp
Determine Required Final Drive
A
(5)
Select final drive >= required final drive
Calculate motor speed required at maximum displacement of Pump (Nmr)
Increase Motor size
Calculate required pump displacement for max. Motor speed (Dpr)
Choose Pump displacement Dp >= Required Dpr
Nmr Nml
Determine Required Final Drive
(6)
(7)
NO
YES
Increase Motor size Dm
NO
YES
Flush control valve
Check valve
Motor
DCV
pump
High press.relief valve(500 bar )
Purgerelief valve(10bar )
Fig . Hydraulic circuit diagram of Hydrostatic transmission ( Forward position of DCV )
filter
Relief valve( 25bar )
Flush control valve
Check valve
Motor
DCV
pump
High press.relief valve(500 bar )
Purgerelief valve(10bar )
Fig . Hydraulic circuit diagram of Hydrostatic transmission ( Reverse position of DCV )
filter
Relief valve( 25bar )
Flush control valve
Check valve
Motor
DCV
pump
High press.relief valve(500 bar )
Purgerelief valve(10bar )
Fig . Hydraulic circuit diagram of Hydrostatic transmission ( Neutral position of DCV )
filter
Relief valve( 25bar )
The charge pump is critical component of the hydrostatic transmission. It is the heart of the hydrostatic transmission, without the charge flow and charge pressure, the transmission will ceases to function.
The charge pump provides several functions to hydraulic circuit
The primary function of charge pump is to replenish fluid lost through leakages.
Provide flow, under pressure for maintaining back pressure on pump, motor piston.
Provide fluid for the servo piston on the system having servo – controlled transmission.
Provides cooled, cleaned fluid for temperature control and flushing
To properly size a charge pump, several consideration must be taken in to account including the system pressure, input speed, Control requirement, Loop flushing, cooling flow etc.
Charge pump sizing
1)Pump leakage
Assume vol. efficiency of pump is 95 %
Pump leakages =
= 8.538 lpm2) Motor leakage Assume vol. efficiency of pump is 95 %
Motor leakages =
= 8.538 lpm
3) Servo control requirement
For most of the application with stroke time 1-3 second and following table
Shows different size of pump and there required servo volume
Servo control flow required
=47.3586/3
= 0.9471 lpm
Series Servo volume (cm3)
Series 40
M 46 41.787
Series 42
28 cc 27.858
Series 90
42 cc
55 cc
75.9 cc
100cc
27.858
36.2154
47.3586
69.645
4) Loop flushing The amount of loop flushing will normally vary from 7 to 15
lpm depending on the charge pump displacement, input speed, and relative setting between the pump and motor charge relief valves. So loop flushing is 13 lpm
Total charge Flow required = Pump leakages + Motor leakages + Servo
control requirement + Loop flushing
Total charge Flow required = 8.538+8.538+0.9471+ 13 = 31.0231 lpmSo, 11 cc gear pump is selected which is operated by electric motor at
3000 rpm.
Charge flow = Charge pump displacement x Input speed x vol. efficiency of charge pump
= 31.35 lpm.
P = Q x Δp = = 1.375 kW.
Electric motor of 3.730 kW selected which is having maximum speed of 3000 rpm.
=
Fig. Charge pump performance curve
Hose size is designated by nominal inside diameter of the hose, expressed in fraction of inches. Hydraulic hoses are also classified By SAE as R1, R2 etc on the basis of pressure rating.
The connection of hydraulic component in hydraulic system and the relative transportation of hydraulic energy is done by means of pipes, hoses, control block ports etc.
Friction loss is reduced to its minimum by keeping flow speed within certain limit. The Recommended flow velocities for different hoses are
Oil velocities above the recommended range cause excessive turbulence in the oil and creates power losses and heating of the oil.
Suction line : 0.5- 1.5 m/secPressurized line : For pressure < 50 bar, 4-5 m/sec 50 – 100 bar, 5-6 m/sec pressure > 100 bar, 6-7 m/secReturn line : 1.5- 3 m/sec
Selection of hoses
Fig . Hose Size Selection Nomograph
Function of hydraulic fluid Transfer of hydraulic power from the pump to hydraulic motor or
cylinder. Lubrication of moving parts, such as the sliding surface of pistons and
spools, bearing and switches elements etc. Protection of metal surfaces actually wetted by the hydraulic fluid The removal of contamination and dirt, abrasion, water, air etc. The removal of heat loses which have been caused by leakages and
friction loses.Classification1. H :- Hydraulic fluid without additive. These hydraulic fluid are hardly
used at all in hydraulic engineering 2. HL :- Hydraulic fluid with additive for increasing the degree of
protection against rust and for increasing resistant to aging. Generally, these hydraulic fluids are employed in hydraulic engineering at pressure up to approx 200 bar.
3. HLP :- Hydraulic fluid with special high pressure additives which result in increased protection against wear. These hydraulic fluids are used in system above operating pressure 200 bar.
4. HV:- Hydraulic Fluid with extremely low viscosity-temperature interdependence. There other properties are same as HLP fluid
Hydraulic Fluid
Fig 6 . Viscosity-Temperature curve for hydraulic oil Hydraulic oil of ISO 68 grade is chosen so as to ensure proper
functioning of the system. It has a viscosity of 67.5 centistokes at 40°C and 8.66 centistokes at 100°C.
ISO
Grade
Equivalent
SAE Grade
Viscosity Density
Centistokes
Kg/m340 0 C 100 0 C
32 10W 32 5.4 857
46 20 46 6.8 861
68 20W 68 8.7 865
100 30 100 11.4 869
150 40 150 15 872
220 50 220 19.4 875
Table . Comparison between SAE and ISO viscosity classes and there oil properties
The tank used in hydraulic circuit for the storage of fluid is generally called
Reservoir.
Functions of Reservoir
To provide an expansion chamber to accommodate changes in fluid volume
in the working part of the circuit. These will naturally arise from fluid
expansion and contraction with the change in temperature.
Fluid cooling.
Removal of entrained air from the fluid.
Removal of fluid contamination by settlement in the bottom of the tank.
Reservoir Sizing
Reservoir capacity should be three times the rated delivery of the pump in
liters per minute. So pump delivery per minute is 33 lpm , In order to avoid
the Heat exchanger in the hydraulic circuit , Reservoir selected having
capacity of 200 liters which is near about 6 times the rated delivery of
pump.
Reservoir
Flange coupling consist of one keyed to the driving shaft and other to the driven shaft as shown in fig .
The two flanges are connected together by means of bolts arranged on circle concentric with the axis of the shaft.
Power is transferred from the driving shaft to the left side flange through key. It is then transmitted from the left side flange to right side flange through bolts.
Rigid flange coupling designed for connecting two shafts. The input shaft transmit 7.46 kW power at 3000 rpm to output shaft through Flange Gear coupling.
Design of flange coupling
Step 1: Selection of material
1) Flange has a complex shape and easiest method to make flange is
casting. Grey cast iron FG200 ( Sut =200 N/mm2) is selected as
the material for the flange from manufacturing consideration. It
is assumed that ultimate shear strength is one half of the ultimate
tensile strength. The factor of safety for the flange is assumed as
6.
2) The keys and bolts are subjected to shear and compressive
stresses. On the basis of strength criteria plain carbon of grade
30C8 (Syt= 400 N/mm2) is selected .It is assumed that
compressive yield strength is 150 % of the tensile strength. The
factor of safety for keys and bolts is taken as 2.5. Step 2: Permissible stresses
I. Flange N/mm2
II. Keys and bolts N/mm2
16.67
6
2000.5
f
ssτ
s
u
80
2.5
4000.5
f
ssτ
s
u
Step 3: Dimension of flange
The dimension of flanges are as follows,
dh =outer diameter of hub = 1.5 d = 1.5( 38) =
57 mm
lh =length of hub = 1.5 d = 1.5 (38) =
57 mm
D=pitch circle diameter of bolts = 2 d = 2(38) = 72
mm
t = thickness of flanges = 0.5 d = 0.5 (38)=19
mm
Do=outside diameter of flange = (1.5d 2t) = (1.5 =
95mm
d= Shaft diameter = 38 mm
The hub is treated as a hollow shaft subjected to torsional
moment
The torsional shear stress in hub is given by,
Where, Mt = design torque
678131625.22
32
3857π
32
ddπJ
4444h
4mm
J
rMτ t
N/mm2
N/mm2
Stress in the flange at the junction of the hub is determined by ,
N/mm2
N/mm2
where , P Power ,kW N rpm Mt Design torque , N-mm. 1.5 Service factor
N-mm
The stress in the flange within the limit.
35618..873000π2
1.57.461060
Nπ2
1.5P1060M
66
t
1.62831625.226
3835618.8762τ
16.67τ
0.36735719π
35618.87622
dhtπ
M2τ
22t
16.67τ
Step 4: Diameter of bolts
The diameter of electric motor shaft is 38 mm , d 40 mm , The number of bolts (N) are 4 .
mm2
Diameter of 4 mm bolt was selected.
The compressive stress in the bolt is determined by ,
N/mm2
N/mm2
3.729580476π
35618.87628
τNDπ
M8d t2
Dtd1N
Mt2σ c
893.2721963
8762.356182
C
240σC
Step 5: Dimension of key
From Table 3 , the standard cross-section of the flat key for 38 mm
diameter shaft is 12 x 8 mm.
The length of the key is equal to lh = 57 mm
The dimension of flat key is 12 x 8 x 57
The torsional shear stress in key is given by ,
,
The shear and compressive stresses induced in the key are
within permissible limit
The compressive stress in the key is determined by
N/mm2
2.74571238
35618.87622
lbd
Mt2τ
h
16.67τ
hc lhd
Mt4σ
8.22257838
35618.8764σ c
N/mm2
240σ c N/mm2
Fig. Drawing of Flange coupling
Fig. Drawing of Flange Gear coupling
Pressure Gauge
Pressure Relief Valve
Hand Pump
Pressure Relief Valve
Fig . Calibration of Pressure Relief valve
Valve with hand knob for pr. adjustment
Development of hydrostatic transmission system.
Phase 2
1.Electric motor 2. Flange gear coupling 3. Flexible drive gear coupling 4. Variable displacement pump 5.Fixed displacement motor 6. Charge gear pump 7. Reservoir 8. Control panel 9. Mechanical control for variable displacement pump 10. Differential 11. Final drive.
Fig. Layout of hydrostatic transmission system
1
245
3
6
7
8
9
10
11
Charge pressure port
Motor leakage port
Vent portPump leakage port
Fig. Variable displacement pump and fixed displacement motor with port location
Flexible Drive gear coupling
Pressure relief valve
Suction filter (40 𝜇)
Fig . Gear pump set up
1. Charge pump 2.Pressure relief valve 3.Flexible drive gear coupling 4.Electric motor
5. Filter 6.Reservoir 7. Temperature indicator
5
2
3
4
1
6
7
Mechanical lever for speed adjustment
Fig. Hydrostatic Transmission system setup
1.Elecric motor 2. Flange gear coupling 3. Mechanical control for variable displacement pump 4. Variable displacement pump 5.Fixed displacement motor 6.Control panel 7. Data acquisition
1
23
4 5
6
7
1.Potentiometer 2.Potentiometer mounting fixture 3. Protractor
4. Mechanical lever for control of variable displacement pump
Fig. Setup for measurement of angle of mechanical lever for control of
variable displacement pump.
1
23
4
The setup for axle speed measured to determine velocity is
shown in Fig.4.17.
In this case better results were obtained by using a proximity
sensor.
This instrument actually detects the proximity of a magnetic
material projecting out of the circumference bolts of a rear
axle and senses it, sending peak signals at the closest
interaction point
Axle speed measurement
1. Proximity switch 2. Bolt 3. Rear axle Fig. Setup for measurement of axle
speed
1
2
3
t
P
8
t
P
8
60
60
2 N
The rear axle has eight bolts around it, so eight consecutive
peaks will mean that it has turned one revolution.
Let, the number of peaks over a time range t sec be = P
Number of revolution over the time range =
Therefore revolutions per minute, Nt =
Now angular velocity, ωt =
Hence, theoretical velocity,
where r = rolling radius of the tire in m.
Calculation of axle rpm from peak signals
rwV tt
Objective 2
To evaluate the performance of the developed
hydrostatic transmission system in
laboratory condition
These are presented under following section:
Calibration of Potentiometer
Performance of Axle Speed Measuring Device
Laboratory testing of the Prototype
Comparison of Designed Hydrostatic
transmission with the existing transmission
system
Fig. Calibration curve for Potentiometer
Calibration of
Potentiometer
-50
0
50
100
150
0 100 200 300 400 500 600Time in second
Pea
k si
gnal
in m
v
Fig. Plot of peak signals versus time for axle rpm
Performance of Axle Speed
Measuring Device
Laboratory testing of the
Prototype
The prototype of Hydrostatic transmission
was tested for its performance in laboratory
condition. Testing of prototype was conducted in
two phase namely preliminary trial tests and final
test at No load condition
I.Preliminary trial tests included the prototype test in controlled condition i.e. short run and low rpm for figuring out any manufacturing errors, tolerance check, and leakage points
II.The final test at no load condition was conducted after preliminary checks and the prototype was operated with electric motor for final testing at no load.
Fig. Angle Vs axle rpm, velocity at 2250 rpm with 43.42 reduction after hydraulic motor (Forward).
Fig. Angle vs axle rpm, velocity at 2250 rpm with 43.42 reduction after hydraulic motor (Reverse).
Fig. Angle vs axle rpm, velocity at 2250 rpm input speed
with 10.85 reduction after hydraulic motor (Forward).
Fig. Angle vs axle rpm, velocity at 2250 rpm input speed with 10.85 reduction after hydraulic motor (Reverse)
SUMMARY AND CONCLUSIONS
The designed hydrostatic transmission for agricultural
tractor offers gear
reduction in two stages i.e. hydrostatic low and
hydrostatic high.
In hydrostatic low range provides gear reduction from
10.85 to and in high
range it provides gear reduction from 43.42 to .
The designed hydrostatic transmission for agricultural
tractor offers velocity in two stages i.e. hydrostatic low
and hydrostatic high. In hydrostatic low range provides
speed from 0 to 11.998 km/hr in the forward direction and
0 to 4.915 km/hr in the reverse direction. Hydrostatic high
range provides speed from 0 to 49.87 km/hr in the
forward in the forward and 0 to 19.663 km/hr in the
reverse direction.
The developed hydrostatic transmission is expected to
have lesser cost as several existing components such as
clutch, large no of gears in transmission system would
be removedThe designed hydrostatic transmission system gives
infinite no. of gear and therefore hydrostatic
transmission has the ability to adjust engine speed and
transmission gear ratio together to operate at the
point of maximum fuel efficiency for given travel speed
and power requirementThe developed hydrostatic transmission was found
to be lesser in weight as well as compact in size as
compared to the existing gear box.
ReferencesAdarsh, K.,1986.Design & development of hydrostatic steering and
hydrostatic transmission system for agricultural tractor .Unpublished M.
Tech. Thesis, Agricultural and Food engineering Department, IIT,
Kharagpur, India.
Bainer, R.,Kepner, R.A. and Barger,E.L.,1960. Principle of farm
machinery,John wiley and sons, Inc. New York.
Bhandari , V.B, 2010. Design of machine element. 3rd ed. Tata Mc-Graw-
Hill, New delhi. , pp 362-168.
Coffman, B.,and Kocher,M., 2010. Testing Fuel Efficiency of a Tractor with
a Continuously Variable Transmission.Applied Eng. In Agric.26 (1):31-36.
Heatwole, B.and Perumpral,J., (1990). Knowledge based assistant for
component selection & preliminary design of hydrostatic drive. Transaction
of ASAE 6(3):359-366
Heyan, LI. Baorui, C. Biao, MA. ,Man, C., 2009.Study on braking capacity
of hydrostatic transmission vehicle.Intelligent Vehicle symposium,2009
IEEE, 3-5 June 2009, 848 – 851
Michael,B.,1991.Hydrostatic transmission application
consideration for commercial turf maintenance vehicles. ASAE
Distinguished Lecture Series No.16,1-15
Pinches,M .and Ashby,J.1989. Power hydraulics, pp 163-179
Renius, K. T and R. Resch.,2005. Continuously Variable Tractor
Transmission. ASAE Distinguished Lecture Series No.29, 1-35.
Rodger,W. and Borghoff,W.1968. Hydrostatic transmission In
farm and light industrial tractor . SAE Paper No 68-680570, 1-
12.
Rydberg, K.1998.Hydrostatic drive in heavy mobile machinery-
new concept and development trends.SAE Paper No.98-
981989,35-41.
Sauer-sundstrand.(1987). BLN-9885.Selection of Driveline
Components.Rev.A
Sauer-sundstrand.(1987). BLN-9886.Transmission Circuit
Recommendation.
Worn, C. and Walker, A.1965. A gear box Replacement
Hydrostatic drive.SAE Paper No 65- 650689, 286-297
THANK YOU
REVIEW OF LITERATURE
Author Work Concluded
Worn C et .al ( 1965)
A gear box Replacement Hydrostatic drive
The performance obtained from hydrostatic transmission compared with current conventional tractor transmission having shown that the present day hydrostatic unit meets all vehicle requirement with obvious advantages of true steeples transmission the reason for selection of gear box replacement unit are considered and further the use of variable displacement motor with variable displacement pump
Author Work Finding
Rodger W et al .(1968)
Hydrostatic Transmission In Farm and Light Industrial Tractor
Infinitely variable hydrostatic drive provides versatility to meet the diverse demand placed on these vehicle .Adopting the hydrostatic transmission to farm or industrial tractor required development of controls for each application e.g Hand operated speed ratio control
Adarsh Kumar ( 1987)
Design and Development of Hydrostatic Steering and Hydrostatic Transmission for Agricultural Tractor
Hydrostatic transmission was designed for 24 hp tractor the arrangement considered to have one fixed displacement motor & one variable pump according to torque & speed requirement motor selected & on the basis of motor requirement pump, valve block , cooler, filter & hose were selected for above arrangement
Author Work Concluded
Heatwole B et.al (1990)
Knowledge Based assistant for component selection & preliminary design of Hydrostatic Drive
•The rule based expert system, called HSTX, was developed to aid in the selection & sizing of the major component pump ,motor and final drive of hydrostatic transmission. HSTX can serve as an educational tool for teaching HST design.
Michael B(1991)
Hydrostatic Transmission Application Consideration for Commercial Turf Maintenance Vehicles.
•Selecting the proper hydrostatic transmission component for application in commercial turf maintenance vehicles requires close cooperation between vehicle designer and transmission supplier . The information provided is intended to serve as an applicationGuideline for selecting a hydrostatic transmission.
Author Work Concluded
Rydberg et al.(1998)
Hydrostatic Drive in Heavy Mobile Machinery New Concept and Development
Increased overall efficiency of the drive train is also of great importance since transmission is responsible for 60-80 per. Of the total fuel consumption is most further improvement to fully utilize the advantages of hydrostatic drives
Renius K et.al (2005)
Continuously Variable Tractor Transmission
•The efficiency demanded for tractor CVT above 50 kw can not be met by “Direct ” Hydrostatic . A special transmission system is needed called power split system in order to increase efficiency it often needs a planetary for splitting or merging the power .
Author Work Concluded
Xu L et al.( 2009)
Multi-range Hydro-Mechanical Continuously Variable Transmission of Tractor
HMCVT is new type steeples transmission device with wide range, high power, top efficiency
Heyan L et al.(2009)
Study on Breaking Capacity Hydrostatic Transmission Vehicle
The study revels the influencing factor of hydraulic breaking capacity .
Coffman B. et al .( 2010)
Testing Fuel Efficiency of a Tractor With a Continuously Variable Transmission
The result indicate that the c v t automatic transmission was more fuel efficient than the conventional gear transmission .
Machine CP F max , Maximum vehicle
tractive effort, N S ,Maximum vehicle design
speed , km/hr Machine CP , Machine
Corner power, kW
Vehicle velocity with additional gear reduction of 4: 1 before
differential at 2250 rpm i.e. rated engine speed is 12.50 km/hr
Machine CP Machine CP = 115.2027 kW
3600
SFmax
3600
100012.533.1784
T R
TR = 3.8606
TR < 4 use fixed motor
746500.8
1000115.20227TR
HP0.8
Machine
CP
HP , Engine power , kWTR , Effective transmission ratio
Required Motor cp =
(kW) E
E , final drive efficiency assume 95 %
Required Motor CP
Required Motor CP = 121.266 KW.
Machine CP
0.95
115.2027
Motor CP
Dm , cc/rev -maximum motor displacement , 75.9 cc/rev
Nm , rpm -design maximum speed , 2500 rpm
Pm , bar -design maximum pressure, 500 bar
Taking motor torque efficiency as 90 percent
Motor CP
Motor CP = 142.3125 kW
142.3125 > 121.266
Motor cp is greater than required cp value the selected motor satisfies our requirement
60000
PmNmDm0.90
600000
500250075.90.90
Required FD
LR, Wheel loaded radius ,m
Dm ,Max motor displacement ,cc/rev
E , Final drive efficiency
Pm, Maximum pressure, bar
Required FD
Required FD = 41.1185
Design Check : FD ≥ Required final drive
Total final drive = Differential reduction x Final drive reduction x Additional gear
reduction
Total final drive = 3.4545 x 3.143 x 4 = 43.4299. Final drive ≥ Required final drive
EPmDm0.90
100002πLRFmax
0.9550075.90.90
100002π0.6433.1784
Nmr = 37.4899 rps
Nmr = 2249.3970 rpm
Design Check : Nmr Nml
Nml ,Motor speed Limit
From appendix B value of Nml is 3800 rpm so our required motor speed at max. Vehicle
speed is less than motor speed limit
Where ,
Nmr - Required motor speed at max. displacement of pump, rps
Sm - Maximum vehicle speed , m/s
π20.64
0.277712.543.4299Nmr
2πLR
SmFDNmr
Where,
Dpr - Required max pump displacement , cc/rev
Nmr - required motor speed at max vehicle speed, rpm
Dm - Max motor displacement ,cc /rev
NP - maximum pump speed ,rpm
Assuming volumetric efficiency of pump and motor as 95 %
Dpr = 75.66 cc/rev.
So we select 75.9 cc/rev size of pump
75.9 ≥ 75.66
So selected size of pump satisfies our requirement.
Np0.950.95
DmNmrDpr
25000.950.95
75.92249.397Dpr
Tractor parameters Implement parameters Operating Parameters
Front axle static weight = 820 kg
Rear axle static weight =1270 kg
Wheel base (WB) = 193 cm
Hitch point from rear axle = 83.5cm
Tyre size: front = 6 - 16
rear = 14.9 – 28Location of c.g from rearAxle center = 0.766 m
Name = 2-bottom 30 cm
MB plough
Type = mounted
Weight = 250 kg
Width = 2 x 30 cm
CG from hitch point = 55 cm
Depth of operation = 20 cm
Cone index =1200 Kpa
Tractive effort
Max. Gross tractive effort= weight coming on the drive wheels x max. gross traction coefficient
Tractive effort refers to the amount of force available at the wheel or wheel of
the vehicle could exert if it had no resistance to movement.
( ) ( ) ( )
( )CG r m CM r y CM r x
r f
TSW X e W X e P X e PYFWD
WB e e
( ) ( )( )
( )CG f m Y CM f x
r f
TSW WB X e W P X WB e PYRWD
WB e e
Fig. Forces acting on tractor-mould board plough combination
….(1)
….(2)
Where
er,ef offset distance of rear wheel and front wheel reaction reaction respectively, m
Px, Py horizontal and vertical component of soil reaction forces respectively, N
Wm weight of the implement, kg
y centre of resistance of the implement below the ground, m
xcm Location of C G of the implement from rear axle centre, m
RWD, FWD rear axle dynamic weight , front axle dynamic weight, N
X cg location of CG of the tractor from rear axle centre. m
TSW total static weight of the tractor, N
MRRr ,MRRf motion resistance of front and rear tire respectively
For firm soil MRRr = 0.1256, er=MRRr x rr = 0.1256 x 0.64 = 0.080384 m.
MRRf= 0.05818 , ef= MRRf x rf , ef= 0.05818 x 0.3296= 0.019176 m.
Y = 2/3 x depth of operation = 2/3 x 0.20= 0.1333 m.
Py/Px = 0.2
Maximum pull Px based at which front wheel starts lifting
FWD = 0 in equation … (1)
Px = 65.48764 KN
By putting this value of Px in equation … (2) and find out RWD
Wm ( Xcm + er) + 0.2 Px ( Xcm + er) – TSW( Xcg- er)
YPx =
Px = 2452.5(1.385+0.08038) + 0.2 x Px (1.385+0.08038) – 20502.9 (0.766 –0.08038)
0.1333
20502.9 ( 1.93- 0.766+0.019176) + ( 13097.528+2452.5) x (1.385+1.93+0.019176)- 65487.64 x 0.1333
1.93- 0.080384 +0.019176RWD=
RWD = 36063.5176 N
Fmax available = 0.92 x RWD …… From Brixius Equation
Fmax = 0.92 x 36.0635176 = 33.1784 kN
So maximum Tractive effort available or tire can generate is 33.18447 kN.
Motor characteristics Units
Displacement Maximum V max. Cm3/rev 75
SpeedMaximum continuous speed(at max. displacement)
rpm 2400
Max. Speed ( interment)(at maximum displacement , higher speedOn request)
rpm 3800
Pressure Max. operating pressure( intermittent) bar 500
Continuous pressure bar 250
TorqueContinuous output torque(at continuous pressure)
Nm 302
Max. output torque(at maximum operating pressure)
Nm 508
PowerContinuous power(at max. continuous speed, max Displacementand max. continuous pressure)
kW 120
Continuous power(at max. continuous speed, max Displacementand max. operating pressure)
kW 202
Permissible shaft loadsAxial force N 2000
Axial input force N 2000
Weight Fixed displacement motor kg 26
Table No. 1 Technical data of
Hydraulic Motor
Pump characteristics Units
Displacement Maximum V max. Cm3/rev 75.9
SpeedMaximum continuous speed(at max. displacement)
rpm 3100
Max. Speed ( interment)(at maximum displacement , higher speedOn request)
rpm 3500
Minimum continuous speed rpm 500
PressureMax. operating pressure( intermittent) bar 500
Continuous pressure bar 250
TorqueContinuous input torque(at continuous pressure)
Nm 305
Max. input torque(at maximum operating pressure)
Nm 485
PowerContinuous power(at max. continuous speed, max Displacementand max. continuous pressure)
kW 98
Maximum power(at max. continuous speed, max Displacementand max. operating pressure)
kW 157
Permissible shaft loadsAxial force N 2000
Axial input force N 2000Weight Fixed displacement motor kg 49
Table No. 2 Technical data of Hydraulic Pump
Shaft diameter, mm Key size, mm Key depth, mm Key way depth
Above up to and including
b x h mm
68
10121722303844505865758595
110130
810121722303844505865758595
110130150
2 x 23 x 34 x 45 x 56 x 68 x 7
10 x 812 x 814 x 9
16 x 1018 x 1120 x 1222 x 1425 x 1428 x 1632 x 1836 x 20
1.21.82.53.03.54.05.05.05.56.07.07.59.09.0
10.011.012.0
Table 3 Dimension of square and rectangular sunk keys (in mm)
