TABLE OF CONTENTSNo.TitlePage
1STRUCTURAL SCHEME, TORQUES AND ROTATIONS FOR EACH SHAFT2
1.1Structural scheme2
1.2The torques and rotations for each shaft2
2GEAR CALCULUS2
2.1Predimensioning gear calculus2
3DIMENSIONING CALCULUS6
3.1Inputs6
3.2Results7
4GEAR FORCES CALCULUS9
4.1Calculus of forces9
4.2Scheme and direction of the forces10
5SHAFT CALCULUS10
5.1Predimensioning calculus10
5.2Choosing the bearing mountings for both input and output
shafts10
5.3Checking the input shaft for compound loads11
5.3.1Determining the horizontal and vertical reactions in the
bearings11
5.3.2Calculating the reactions in the bearings12
5.3.3Identifying the loads12
6CHOOSING AND CHECKING OF THE KEY ASSEMBLY BETWEEN THE DRIVEN
WHEEL AND THE OUTPUT SHAFT13
7CHOOSING AND CHECKING OF THE BALL-BEARING MOUNTING FOR THE
INPUT SHAFT14
8CHOOSING AND JUSTIFICATION OF THE OILING SYSTEM15
9CHOOSING THE SEALING DEVICES16
10CHAIN TRANSMISSION CALCULUS16
11JUSTIFICATION FOR THE MATERIALS USED19
EXECUTION DRAWING
ENSEMBLE DRAWING
1. STRUCTURAL SCHEME, TORQUES AND ROTATIONS FOR EACH SHAFT
1.1 Structural schemeFig 1.1
1.2 The torques and rotations for each shaft
The motor's shaft:
The input shaft:
The output shaft:
2. GEAR CALCULUS
2.1 Predimensioning gear calculusTable
2.1No.ParameterCalculus
1INPUT DATA
1.1The rotational speed of the pinion
1.2The torque at the pinion
1.3The gear ratio
1.4The imposed lifetime
1.5Functioning conditionsActuator: Asynchronous electric
motor
Load type: Medium shocks
1.6Loading cyclesCycle type: Pulsatory
1.7No of loadings/cycle for the pinion and for the wheel
1.8Refference rack's parameters
2CHOOSING THE STEELS, TREATMENTS AND LIMIT TENSIONS
2.1Choosing the steel, treatments and durabilities for the two
wheels
2.2The limit stresses for contact and bending
3PREDIMENSIONING CALCULUS
3.1Number of teeth - pinion and wheel
3.2The real gear ratio
3.3The contact calculus factors
3.3.1The elasticity factor
3.3.2The contact zone factor
3.3.3The gearing factor
3.3.4The inclination factor
3.4The bending factor calculus
3.4.1Equivalent wheel's teeth number
3.4.2Normal plan profile displacement's coefficients
3.4.3YFa1,2 shape factor for the teeth
3.4.4Tension correction factors for teeth base
3.4.5Covering degree factor
3.4.6teeth slope factor
3.5Load correction factors
3.5.1Working condition factor
3.5.2Dynamic factor
3.5.3Uniform repartition factors for the teeth with load
KH - contact load
KF - bending load
3.5.4Uniform repartition factors for frontal load: contact,
bending
3.6Allowable strengths for contact loading
3.6.1Predimensioning calculus factors
3.6.2Lubrication factor
3.6.3Speed factor
3.6.4Roughness factor
3.6.5Material torque factor
3.6.6Size factor
3.6.7Durability factors for contact loads on the pinion and
driven wheel
3.6.8Minimum safety coefficient for contact load
3.6.9Allowable strengths for bending loading
3.6.10Predimensioning calculus factorsCorrection factor for
bending
Sensibility relative factor at tension concentrator for tooth
base of pinion and wheel
Relative roughness factor for the connection zone at tooth base
of pinion and wheel
Size factor
Resistance factors for bending load of pinion and driven
wheel
Minimum safety coefficient at bending stress
3.7PREDIMENSIONING - AXES DISTANCE
3.7.1Width coefficients
3.7.2Axes distance from contact load resistance condition
3.7.3Axes distance from bending load resistance condition
3.7.4Axes distance selection for predimensioning
3. DIMENSIONING CALCULATION
3.1 InputsTable 3.1ParameterSymbolValue
Inputs
PowerP, kW18.5
LifetimeLh, hrs10000
Functioning factorKA1,6
Type of engineAsynchronous electric motor
Load variationMedium shocks
Rotation at pinionnI, rpm1523.08
Gear ratioudat4
Helical angle, 10
Center distanceaW, mm125
Width factora0,35
Materials
SteelCarburised
Type18MnCr11
Superficial hardnessHRC>58
Inside hardnessHB270...360
Recommended limit contact stressMPa1500
Limit contact stressHlim1,2, MPa1500
Recommended limit bending stressMPa500
Limit bending stressFlim1,2, MPa500
Teeth numbers
z118
z272
Machining information
Profile roughnessRa, m>0,4
Root roughnessRa, m3,2
Functioning conditionsPulsatory
Minimal safety coefficients
For contact stressSHmin1,2
For bending stressSFmin1,5
3.2 ResultsTable 3.2ParameterGear
PinionWheeel
Gear Parameters
Center distanceaW= 125mm
Refference center distancea= 124,26283 mm
Normal modulemn= 2,75 mm
Frontal modulemt= mm
Helix angle= 10
Helix angle on the base circleb= 9,39129
Frontal pressure anglet= 20,28356
Gearing angle:frontalwt= 21,17892
normalwn= 20,88168
Total addendum correction coefficient, in normal planexsn=
0,27377
Contact degree:frontal= 1,52
helical= 0,88
total= 2,40
Speed on the pitch circlev= 4,01 m/s
Precission8
Lubricant viscosity50= 180 cSt
Roughness:active profileRa= 0,80 m
fillet (root)Ra= 1,60 m
Pinion and wheel parameters
Addendum diametersda1= 57,71223 mmda2= 203,25638 mm
Deddendum diametersdf1= 45,36862 mmdf2= 190,91277 mm
Pitch diametersd1= 50,26362 mmd2= 198,26205 mm
Rolling circle diametersdw1= 50,56180 mmdw2= 199,43820 mm
Base circle diametersdb1= 47,14669 mmdb2= 185,96751 mm
Numbers of teethz1= 18z2= 71
Widthb1= 42 mmb2= 40 mm
Normal addendum correction coefficientxn1= 0,36xn2= -0,08623
Frontal addendum correction coefficientxt1= xt2=
Minimum normal addendum correction coefficientxnmin1=
-0,105xnmin2= -3,357
Tooth width on the addendum circle, on the normal planeSan1=
1,48 mmSan2= 2,23 mm
Minimum tooth width on the addendum circle, on the normal
planeSanmin1= 1,1 mmSanmin2= 1,1 mm
Equivalent gear parameters
Numbers of teethzn1= 18,78zn2= 74,07
Pitch diametersdn1= 51,63857 mmdn2= 203,68545 mm
Base circle diametersdbn1= 48,52438 mmdbn2= 191,40172 mm
Addendum diametersdan1= 59,08718 mmdan2= 208,67979 mm
Center distanceawn= 128,339631 mm
Contact degreen= 1,56
ParameterGear
ContactBending
Calculus factors
Functioning factorKA= 1,6
Dynamic factorK= 1,00
Axial load factorKH= 1,57KF= 1,48
Transverse load factorKH= 1,47KF= 1,47
Elastic factorZE= 189,8-
Contact factorZH= 2,41-
Helix angle factorZ= 0,99Y= 0,92
Shape factor-YFa1= 2,32
YFa2= 2,27
Correction factor for bending stress-YSa1= 1,73
YSa2= 1,73
Lubricating factorZL= 1,04-
Speed factorZV= 0,98-
Roughness factorZR= 0,97YR= 1
Sensitive factor-Y1= 0,993
Y2= 0,987
Number of cycles:pinionNL1= 3,3*108
wheeelNL2= 4,6*107
Lifetime factorsZN1= 1
ZN2= 1YN1= 1
YN2= 1
Minimal safety factorsSHmin= 1,2SFmin= 1,5
Stresses
Limit stress (chosen)Hlim1= 1500 MPaFlim1= 500 MPa
Hlim2= 1500 MPaFlim2= 500 MPa
Permissible stressHP1= 1234,5 MPaFP1= 664,4 MPa
HP2= 1234,5 MPaFP2= 660,5 MPa
Real stressH1= 1217,2 MPaF1= 373,6 MPa
H2= 1217,2 MPaF2= 385,5 MPa
Width coefficient:preliminarya= 0,35
finalarec= 0,31
Control elements
Dimension over teeth
Number of teeth for dimension over teethN1= 3N2= 9
Normal dimension over teethWNn1= 21,69735 mmWNn2= 71,70059
mm
Tooth width
Normal tooth widthScn1= 4,45074 mmScn2= 3,66195 mm
Frontal tooth widthSct1= 4,44267 mmSct2= 3,65531 mm
Height of the normal tooth widthhcn1= 2,91434 mmhcn2= 1,83075
mm
Height of the frontal tooth widthhct1= 2,90334 mmhct2= 1,82169
mm
4.GEAR FORCES CALCULUS
4.1 Calculus of forces
4.2 Scheme and direction of the forces
Fig 4.1
5. SHAFT CALCULUS
5.1 Predimensioning calculus
5.2 Choosing the bearing mountings for both input and output
shafts
Table 5.1ShaftSymbold (mm)D (mm)B (mm)CrC0
input6204204714127006550
output6205255215140007800
5.3 Checking the input shaft for compound loadsFig 5.1
Input shaft's length: l=73 mm (measured on the drawing)
5.3.1 Determining the horizontal and vertical reactions in the
bearings
Fig 5.2
Fig 5.3
5.3.2 Calculating the reactions in the bearings
5.3.3 Identifying the loads
Compression:
Torsion:
Bending:
6. CHOOSING AND CHECKING OF THE KEY ASSEMBLY BETWEEN
THE DRIVEN WHEEL AND THE OUTPUT SHAFT
Fig 6.1
7. CHOOSING AND CHECKING OF THE BALL-BEARING
MOUNTING FOR THE INPUT SHAFT
Fig 7.1
We select a radial bearing mounting with balls, in XFig 7.2
Checking the ball-bearing mounting by the dynamic load
capacity
Durability of the bearing
Necesary dynamic capacity
for Fa / Fr e , X = 1 and Y = 0;
for Fa / Fr > e, X = 0,4 and Y is chosen from the tableTable
7.1Fa / C00,0250,040,0650.120,170,5
e0,40,420,440,480,50,56
Y1,421,361,271,161,111
8. CHOOSING AND JUSTIFICATION OF THE OILING SYSTEM
Oiling the gearings:The gears from speed reducers are grease
through splashing in the oil bath. For this aim in which a gear
from the gearing mechanism is introduce in the oil bath until a
tooth is covered with oil, not more than 10 mm, and without passing
six time the modulus.
In case of speed reducers with more steps (when the wheels dont
reach the bath), the grease is made with a parasite gear, or with
the help of some discs or splashing spoons which are creating am
oil fog.
The grease through splashing is applied on gearing mechanisms
that are working periodically, with speeds up to 15m/s. For greater
peripheral speeds, the grease is done with oil injectors. The oil
pressure is about 0.1-0.8at. For greasing, mineral oils are use
with the viscosity of 3-60 degrees E50^C.With how much the
peripheral speed is smaller, the contact pressure and the roughness
are higher, and more viscous oils are used.On speed reducers with
more gears, the oil is choused with a viscosity corresponding to
the steps that transmit the biggest torque. For the oil bath volume
are considered 0.25-0.5l of oil over a horsepower. The period of
oil change is about 1000-5000 hours of functioning (for the case
when the gearing mechanism is sealed and the oil is filtrate every
2500 hours). For filtering can be used magnetic filters. When the
speed reducer is new, the oil must be changed after
200-300hours.Oiling the ball-bearings:The choose of lubricants for
ball bearings and establishing the grease intervals, is done
considering the dimension, number of revolutions, load and work
temperature of the bearing.
Generally, the liquid lubricants have more advantages then the
consistent ones: higher physical-chemical stability, can be used at
high speeds and temperatures, and also at very low temperatures,
easier evacuation of heat produced in the bearing, smaller
resistance sported by the rolling bodies.
Disadvantages: difficult bearing sealing, loses through leakages
in time, etc.
Grease lubrication is more advantageous because leads to:
simpler bearings construction, easy to seal, with a lower cost,
better protection of the balls to external impurities, lower
lubricant looses.Gaskets: due to high levels of revolutions trees
cuffs sealing is accomplished by rotation.9. CHOOSING THE SEALING
DEVICESFig 9.1
dm1=drul1 - (2...3)=20-2=18 mm
dm2=drul2 - (2...3)=30-3=22 mm
we select a felt cuff A 18x30 STAS 7950/2-72 with h=7 mm for the
input shaft
we select a felt cuff A 22x40 STAS 7950/2-72 with h=10 mm for
the output shaft10. CHAIN TRANSMISSION CALCULUS
Table 10.1No.ParameterCalculus
1INPUT DATA
1.1PowerP=4 kW
1.2Transmission ratioiL=1,95
1.3Torsion moment for the chain driving wheel
1.4Driving wheel speedn=2970 rpm
1.5Working conditionsStatic load, horizontal transmission, no
adjustement, periodic drip oiling, one shift working condition
2KINEMATIC GEOMETRICAL ELEMENTS
2.1Driving wheel's number of teethz1=28
2.2Driven wheel's number of teethz2=z1*iL=28*1,95=54,6~55 <
z2max=120
2.3Pitch
we select p1=19,05 mm (12B); p2=25,40 mm (16A); p3=25,40 mm
(16B) from STAS
2.4Diameter of the driving wheel
2.5Diameter of the driven wheel
2.6Medium speed
2.7Crushing area of bolt and sleeve
3ESTABLISHING THE OPTIMUM CHAIN TYPE
3.1Global correction coefficient
3.1.1Dynamic coefficient of load
3.1.2Axes distance coefficient
3.1.3Centers of wheels inclination towards the horizontal
coefficient
3.1.4Stretch adjustement method coefficient
3.1.5Oiling method coefficient
3.1.6Functioning conditions coefficient
3.2Admissible useful force
3.3Admissible useful power
3.4Number of rows
3.5Establishing the optimum chain variant
4FINALISING THE GEOMETRICAL ELEMENTS
4.1Preliminary axes distance
4.2Number of links
4.3Length of the chain
4.4Recalculated axes distance
5CHAIN TRANSMISSION FORCES
5.1The stretching force due to chain's own weight
5.1.1Centers of wheels position towards the horizontal
coefficient
5.1.2Weight pe liniar meter of chain
5.2The stretching force due to centrifugal forces
5.3The force in the passive branch of the chain
5.4Useful force
5.5The force in the active branch of the chain
5.6Checking the chain at brakage
5.7The force acting on the shafts
5.8Wheel's width
11. JUSTIFICATION FOR THE MATERIALS USEDIt is choosen 18MnCr11
steel hardening and tempering to achieve the shaft and gear wheel
transmission because this steel has good resistance to bending and
also has a high resistance to fatigue.Materials used for speed
reducer construction.Materials used for gears: Steel
It is used great steel: steel with carbon0.4-0.6 %C and steel
with 0.35-0.45%C low alloyed with Mn, Cr, Cr-Mo, Cr-Ni etc. Steel
non alloyed with Cr, Cr=Mo, Cr-Ni, with cyaniding
Cast irons
Cast irons are used at gearing which has a easy working, change
wheels which dosent functioning every time.When it is asking a
silent condition may be used normal iron ash.
Used material for axels execution:Generally the axel which don t
have a heat treatment are made by normally steel carbon: OL 50, OL
60, Stas 500-78
For axel which a big lifting power we can use carbon steel of
quality: OLC 35, OLC 45, OLC 60, according to STAS 880-66.
In case of axel which have a strong load and are required small
dimension are used steel alloyed with crom, Cr-Ni or
Cr-Mn.Marerials used for producing the body.The body because of the
stiffness are made by cast irons or by casting steel. Most of the
body are made by cast iron with average resistance Fc 200, Fc
250.19
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