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Saturday, April 9, 2011

Density of Metals *


http://www.bibliobase.com/Chemistry/Brenstein/pdf_doc/05.pdf

You are certainly aware that different materials weigh differently, even if they have the same volume.  In this experiment, you will determine the density of several unknown solid metal samples from their masses and volumes.  You will then identify the unknown samples based on your calculated densities.

  Densities of selected metals
Metal
Density (g/cm3)
Aluminum
2.699
Chromium
7.13
Copper
8.89
Gold
19.33
Iron
7.86
Lead
11.347
Magnesium
1.738
Mercury
13.596
Nickel
8.85
Osmium
22.5
Potassium
0.87
Silver
10.6
Tin
7.184
zinc
7.19
Note that the density of metals actually varies somewhat depending on how the metal is produced.  Your measured value may thus be different from the value listed above, but it may still be correct.

Friday, April 1, 2011

Helical Gears

Introduction
Helical gears are similar to spur gears except that the gears teeth are at an angle with the axis of the gears.  A helical gear is termed right handed or left handed as determined by the direction the teeth slope away from the viewer looking at the top gear surface along the axis of the gear.   ( Alternatively if a gear rests on its face the hand is in the direction of the slope of the teeth) .   Meshing helical gears must be of opposite hand. Meshed helical gears can be at an angle to each other (up to 90o ).  The helical gear provides a smoother mesh and can be operated at greater speeds than a straight spur gear.   In operatation helical gears generate axial shaft forces in addition to the radial shaft force generated by normal spur gears.


In operation the initial tooth contact of a helical gear is a point which develops into a full line contact as the gear rotates.   This is a smoother cycle than a spur which has an initial line contact.  Spur gears are generally not run at peripheral speed of more than 10m/s. Helical gears can be run at speed exceeding 50m/s when accurately machined and balanced.

Standards ... The same standards apply to helical gears as for spur gears
  • AGMA 2001-C95 or AGMA-2101-C95 Fundamental Rating factors and Calculation Methods for involute Spur Gear and Helical Gear Teeth
  • BS 436-4:1996, ISO 1328-1:1995..Spur and helical gears. Definitions and allowable values of deviations relevant to corresponding flanks of gear teeth
  • BS 436-5:1997, ISO 1328-2:1997..Spur and helical gears. Definitions and allowable values of deviations relevant to radial composite deviations and runout information
  • BS ISO 6336-1:1996 ..Calculation of load capacity of spur and helical gears. Basic principles, introduction and general influence factors
  • BS ISO 6336-2:1996..Calculation of load capacity of spur and helical gears. Calculation of surface durability (pitting)
  • BS ISO 6336-3:1996..Calculation of load capacity of spur and helical gears. Calculation of tooth bending strength
  • BS ISO 6336-5:2003..Calculation of load capacity of spur and helical gears. Strength and quality of materials


Helical gear parameters

A helical gear train with parallel axes is very similar to a spur gear with the same tooth profile and proportions.  The primary difference is that the teeth are machined at an angle to the gear axis.
Helix Angle ..
The helix angle of helical gears β is generally selected from the range 6,8,10,12,15,20 degrees. The larger the angle the smoother the motion and the higher speed possible however the thrust loadings on the supporting bearings also increases.   In case of a double or herringbone gear β values 25,30,35,40 degrees can also be used.  These large angles can be used because the side thrusts on the two sets of teeth cancel each other allowing larger angles with no penalty

Pitch /module ..
For helical gears the circular pitch is measured in two ways
The traverse circular pitch (p) is the same as for spur gears and is measured along the pitch circle
The normal circular pitch p n is measured normal to the helix of the gear.
The diametric pitch is the same as for spur gears     ... P = z g /dg = z p /d p ....d= pitch circle dia (inches).
The module is the same as for spur gears     ... m = dg/z g = d p/z p.... d = pitch circle dia (mm).

Helical Gear geometrical proportions
  • p = Circular pitch = d g. p / z g = d p. p / z p
  • p n = Normal circular pitch = p .cosβ
  • P n =Normal diametrical pitch = P /cosβ
  • p x = Axial pitch = p c /tanβ
  • m n =Normal module = m / cosβ
  • α n = Normal pressure angle = tan -1 ( tanα.cos β )
  • β =Helix angle
  • d g = Pitch diameter gear = z g. m
  • d p = Pitch diameter pinion = z p. m
  • a =Center distance = ( z p + z g )* m n /2 cos β
  • a a = Addendum = m
  • a f =Dedendum = 1.25*m
  • b = Face width of narrowest gear


Herringbone / double crossed helical gears
Crossed Helical Gears
When two helical gears are used to transmit power between non parallel, non-intersecting shafts, they are generally called crossed helical gears.  These are simply normal helical gears with non-parallel shafts.  For crossed helical gears to operate successfully they must have the same pressure angle and the same normal pitch.   They need not have the same helix angle and they do not need to be opposite hand.   The contact is not a good line contact as for parallel helical gears and is often little more than a point contact.  Running in crossed helical gears tend to marginally improve the area of contact.

The relationship between the shaft angles E and the helix angles β 1 & β2 is as follows
E = (Same Helix Angle) β 1 + β 2 ......(Opposite Helix Angle) β 1 - β 2
For gears with a 90o crossed axis it is obvious that the gears must be the same hand.

The centres distance (a) between crossed helical gears is calculated as follows
a = m * [(z 1 / cos β 1) + ( z 1 / cos β 1 )] / 2
The sliding velocity Vsof crossed helical gears is given by
Vs = (V1 / cos β 1 ) = (V 2 / cos β 2 )


Strength and Durability calculations for Helical Gear Teeth
Designing helical gears is normally done in accordance with standards the two most popular series are listed under standards above: The notes below relate to approximate methods for estimating gear strengths. The methods are really only useful for first approximations and/or selection of stock gears (ref links below). � Detailed design of spur and helical gears should best be completed using :
a) Standards.
b) Books are available providing the necessary guidance.
c) Software is also available making the process very easy.    A very reasonably priced and easy to use package is included in the links below (Mitcalc.com)


The determination of the capacity of gears to transfer the required torque for the desired operating life is completed by determining the strength of the gear teeth in bending and also the durability i.e of the teeth ( resistance to wearing/bearing/scuffing loads ) .. The equations below are based on methods used by Buckingham..

Bending
The Lewis formula for spur gears can be applied to helical gears with minor adjustments to provide an initial conservative estimate of gear strength in bending.    This equation should only be used for first estimates. 
σ = Fb / ( ba. m. Y )
  • Fb = Normal force on tooth = Tangential Force Ft / cos β
  • σ = Tooth Bending stress (MPa)
  • ba = Face width (mm)
  • Y = Lewis Form Factor
  • m = Module (mm)
When a gear wheel is rotating the gear teeth come into contact with some degree of impact.  To allow for this a velocity factor is introduced into the equation.   This is given by the Barth equation for milled profile gears.
K v = (6,1 + V ) / 6,1

V = the pitch line velocity = PCD.w/2

The Lewis formula is thus modified as follows
σ = K v.Fb / ba. m. Y
The Lewis form factor Y must be determined for the virtual number of teeth z' = z /cos3β The bending stress resulting should be less than the allowable bending stress Sb for the gear material under consideration.   Some sample values are provide on this page ef Gear Strength Values

Surface Strength
The allowable gear force from surface durability considerations is determined approximately using the simple equation as follows
Fw = K v d p b a Q K / cos2β
Q = 2. dg /( dp + dp ) = 2.zg /( zp +zp )
Fw = The allowable gear load. (MPa)

K = Gear Wear Load Factor (MPa) obtained by look up ref Gear Strength Values





Lewis Form factor for Teeth profile α = 20o , addendum = m, dedendum = 1.25m
Number of teethY Number of teethY Number of teethY Number of teethY Number of teethY
12 0.245170.303220.33134 0.37175 0.435
130.261180.309240.33738 0.3841000.447
14 0.277190.314260.346450.4011500.460
150.290200.322280.353500.4093000.472
160.296210.328300.35960 0.422Rack0.485


http://www.roymech.co.uk/Useful_Tables/Drive/Hellical_Gears.html

Worm Gears

Introduction

A worm gear is used when a large speed reduction ratio is required between crossed axis shafts which do not intersect.   A basic helical gear can be used but the power which can be transmitted is low.  A worm drive consists of a large diameter worm wheel with a worm screw meshing with teeth on the periphery of the worm wheel.   The worm is similar to a screw and the worm wheel is similar to a section of a nut.   As the worm is rotated the wormwheel is caused to rotate due to the screw like action of the worm.  The size of the worm gearset is generally based on the centre distance between the worm and the wormwheel.

If the worm gears are machined basically as crossed helical gears the result is a highly stress point contact gear.  However normally the wormwheel is cut with a concave as opposed to a straight width.   This is called a single envelope worm gearset.   If the worm is machined with a concave profile to effectively wrap around the wormwheel the gearset is called a double enveloping worm gearset and has the highest power capacity for the size.  Single enveloping gearsets require accurate alignment of the worm-wheel to ensure full line tooth contact. Double enveloping gearsets require accurate alignment of both the worm and the wormwheel to obtain maximum face contact.

The worm is shown with the worm above the wormwheel.  The gearset can also be arranged with the worm below the wormwheel.   Other alignments are used less frequently.




Nomenclature

As can be seen in the above view a section through the axis of the worm and the centre of the gear shows that , at this plane, the meshing teeth and thread section is similar to a spur gear and has the same features

αn = Normal pressure angle = 20o as standard
γ = Worm lead angle = (180 /π ) tan-1 (z 1 / q)(deg)   ..Note: for α n= 20o  γ should be less than 25o
b a = Effective face width of worm wheel. About 2.m √ (q +1) (mm)
b l = Length of worm wheel. About 14.m. (mm)
c = clearance   c min = 0,2.m cos γ ,  c max = 0,25.m cos γ (mm)
d 1 = Ref dia of worm (Pitch dia of worm (m)) = q.m (mm)
d a.1 = Tip diameter of worm = d 1 + 2.h a.1 (mm)
d 2 = Ref dia of worm wheel (Pitch dia of wormwheel) =( p x.z/π ) = 2.a - d 1 (mm)
d a.2 = Tip dia worm wheel (mm)
h a.1 = Worm Thread addendum = m (mm)
h f.1 = Worm Thread dedendum , min = m.(2,2 cos γ - 1 ) , max = m.(2,25 cos γ - 1 )(mm)
m = Axial module = p x /π (mm)
m n = Normal module = m cos γ(mm)
M 1 = Worm torque (Nm)
M 2 = Worm wheel torque (Nm)
n 1 = Rotational speed of worm (revs /min)
n 2 = Rotational speed of wormwheel (revs /min)
p x = Axial pitch of of worm threads and circular pitch of wheel teeth ..the pitch between adjacent threads = π. m. (mm)
p n = Normal pitch of of worm threads and gear teeth (m)
q = diameter factor selected from (6   6,5   7   7,5   8   8,5   9  10   11   12   13   14   17   20 )
p z = Lead of worm = p x. z 1 (mm).. Distance the thread advances in one rev'n of the worm.   For a 2-start worm the lead = 2 . p x
R g = Reduction Ratio
q = Worm diameter factor = d 1 / m - (Allows module to be applied to worm )
μ = coefficient of friction
η= Efficiency
Vs = Worm-gear sliding velocity ( m/s)
z 1 = Number of threads (starts) on worm
z 2 = Number of teeth on wormwheel






Worm gear design parameters
Worm gears provide a normal single reduction range of 5:1 to 75-1.  The pitch line velocity is ideally up to 30 m/s.  The efficiency of a worm gear ranges from 98% for the lowest ratios to 20% for the highest ratios.  As the frictional heat generation is generally high the worm box is designed disperse heat to the surroundings and lubrication is and essential requirement.  Worm gears are quiet in operation.  Worm gears at the higher ratios are inherently self locking - the worm can drive the gear but the gear cannot drive the worm.   A worm gear can provide a 50:1 speed reduction but not a 1:50 speed increase....(In practice a worm should not be used a braking device for safety linked systems e.g hoists.  . Some material and operating conditions can result in a wormgear backsliding )

The worm gear action is a sliding action which results in significant frictional losses.   The ideal combination of gear materials is for a case hardened alloy steel worm (ground finished) with a phosphor bronze gear.  Other combinations are used for gears with comparatively light loads.





Specifications
BS721 Pt2 1983 Specification for worm gearing � Metric units.
This standard is current (2004) and provides information on tooth form, dimensions of gearing, tolerances for four classes of gears according to function and accuracy, calculation of load capacity and information to be given on drawings.






Worm Gear Designation
Very simply a pair of worm gears can be defined by designation of the number of threads in the worm ,the number of teeth on the wormwheel, the diameter factor and the axial module i.e z1,z2, q, m .

This information together with the centre distance ( a ) is enough to enable calculation of and any dimension of a worm gear using the formulea available.






Worm teeth Profile

The sketch below shows the normal (not axial) worm tooth profile as indicated in BS 721-2 for unit axial module (m = 1mm) other module teeth are in proportion e.g. 2mm module teeth are 2 times larger
Typical axial modules values (m) used for worm gears are
0,5    0,6     0,8    1,0     1,25    1,6    2,0    2,5    3,15     4,0     5,0     6,3    8,0    10,0     12,5    16,0    20,0    25,0    32,0    40,0    50,0




Materials used for gears




MaterialNotesapplications
Worm
Acetal / Nylon Low Cost, low duty Toys, domestic appliances, instruments
Cast Iron Excellent machinability, medium friction. Used infrequently in modern machinery
Carbon Steel Low cost, reasonable strength Power gears with medium rating.
Hardened Steel High strength, good durability Power gears with high rating for extended life
Wormwheel
Acetal /Nylon Low Cost, low duty Toys, domestic appliances, instruments
Phos Bronze Reasonable strength, low friction and good compatibility with steel Normal material for worm gears with reasonable efficiency
Cast Iron Excellent machinability, medium friction. Used infrequently in modern machinery






Design of a Worm Gear
The following notes relate to the principles in BS 721-2
Method associated with AGMA are shown below..
Initial sizing of worm gear.. (Mechanical)
1) Initial information generally Torque required (Nm), Input speed(rpm), Output speed (rpm).
2) Select Materials for worm and wormwheel.
3) Calculate Ratio (R g)
4) Estimate a = Center distance (mm)
5) Set z 1 = Nearest number to (7 + 2,4 SQRT (a) ) /R g
6) Set z 2 = Next number < R g . z 1
7) Using the value of estimated centre distance (a) and No of gear teeth ( z 2 )obtain a value for q from the table below
8) d 1 = q.m (select) ..
9) d 2 = 2.a - d 1
10) Select a wormwheel face width b a (minimum =2*m*SQRT(q+1))
11) Calculate the permissible output torques for strength (M b_1 and wear M c_1 )
12) Apply the relevent duty factors to the allowable torque and the actual torque
13) Compare the actual values to the permissible values and repeat process if necessary
14) Determine the friction coefficient and calculate the efficiency.
15) Calculate the Power out and the power in and the input torque


6) Complete design of gearbox including design of shafts, lubrication, and casing ensuring sufficient heat transfer area to remove waste heat.






Initial sizing of worm gear.. (Thermal)
Worm gears are often limited not by the strength of the teeth but by the heat generated by the low efficiency. It is necessary therefore to determine the heat generated by the gears = (Input power - Output power). The worm gearbox must have lubricant to remove the heat from the teeth in contact and sufficient area on the external surfaces to distibute the generated heat to the local environment. This requires completing an approximate heat transfer calculation. If the heat lost to the environment is insufficient then the gears should be adjusted (more starts, larger gears) or the box geometry should be adjusted, or the worm shaft could include a fan to induced forced air flow heat loss.




Formulae
The reduction ratio of a worm gear ( R g )
R g = z 2 / z 1
eg a 30 tooth wheel meshing with a 2 start worm has a reduction of 15

Tangential force on worm ( F wt )= axial force on wormwheel
F wt = F ga = 2.M 1 / d 1

Axial force on worm ( F wa ) = Tangential force on gear
F wa = F gt = F wt.[ (cos α n - μ tan γ ) / (cos α n . tan γ + μ ) ]

Output torque ( M 2 ) = Tangential force on wormwheel * Wormwheel reference diameter /2
M 2 = F gt* d 2 / 2
Relationship between the Worm Tangential Force F wt and the Gear Tangential force F gt
F wt = F gt.[ (cos α n . tan γ + μ ) / (cos α n - μ tan γ ) ]
Relationship between the output torque M 2and the input torque M 1
M 2 = ( M 1. d 2 / d 1 ).[ (cos α n - μ tan γ ) / (cos α n . tan γ + μ ) ]

Separating Force on worm-gearwheel ( F s )
F s = F wt.[ (sin α n ) / (cos α n . sin γ + μ .cos γ ) ]


Sliding velocity ( V s )...(m/s)
V s (m/s ) = 0,00005236. d 1. n 1 sec γ
= 0,00005235.m.n (z 12 + q 2 ) 1/2
Peripheral velocity of wormwheel ( V p) (m/s)
V p = 0,00005236,d 2. n 2





Friction Coefficient

Cast Iron and Phosphor Bronze .. Table x 1,15
Cast Iron and Cast Iron.. Table x 1,33
Quenched Steel and Aluminum Alloy..Table x 1,33
Steel and Steel..Table x 2
Friction coefficients - For Case Hardened Steel Worm / Phos Bros Wheel
Sliding Speed Friction Coefficient Sliding Speed Friction Coefficient
m/s μ m/s μ
0 0,145 1,5 0,038
0,001 0,12 2 0,033
0,01 0,11 5 0,023
0,05 0,09 8 0,02
0,1 0,08 10 0,018
0,2 0,07 15 0,017
0,5 0,055 20 0,016
1 0,044 30 0,016






Efficiency of Worm Gear

The efficiency of the worm gear is determined by dividing the output Torque M2 with friction = μ by the output torque with zero losses i.e μ = 0

First cancelling [( M 1. d 2 / d 1 ) / M 1. d 2 / d 1 ) ] = 1
Denominator = [(cos α n / (cos α n . tan γ ] = cot γ
η = [(cos α n - μ tan γ ) / (cos α n . tan γ + μ ) ] / cot γ
= [(cos α n - μ .tan γ ) / (cos α n + μ .cot γ )]

Graph showing worm gear efficiency related to gear lead angle ( γ )








Worm Design /Gear Wear / Strength Equations to BS721
Note: For designing worm gears to AGMA codes AGMA method of Designing Worm Gears

The information below relates to BS721 Pt2 1983 Specification for worm gearing � Metric units.  BS721 provides average design values reflecting the experience of specialist gear manufacturers.   The methods have been refined by addition of various application and duty factors as used.  Generally wear is the critical factor..
Permissible Load for Strength
The permissible torque (M in Nm) on the gear teeth is obtained by use of the equation
M b = 0,0018 X b.2σ bm.2. m. l f.2. d 2.
( example 87,1 Nm = 0,0018 x 0,48 x 63 x 20 x 80 )

X b.2 = speed factor for bending (Worm wheel ).. See Below
σ bm.2 = Bending stress factor for Worm wheel.. See Table below
l f.2 = length of root of Worm Wheel tooth
d 2 = Reference diameter of worm wheel
m = axial module
γ = Lead angle


Permissible Torque for Wear
The permissible torque (M in Nm) on the gear teeth is obtained by use of the equation
M c = 0,00191 X c.2σ cm.2.Z. d 21,8. m
( example 33,42 Nm = 0,00191 x 0,3234 x 6,7 x 1,5157 x 801,8 x 2 )

X c.2 = Speed factor for wear ( Worm wheel )
σ cm.2 = Surface stress factor for Worm wheel
Z = Zone factor.
Length of root of worm wheel tooth
Radius of the root = R r= d 1 /2 + h ha,1 (= m) + c(= 0,25.m.cos γ )
R r= d 1 /2 + m(1 +0,25 cosγ)

l f.2 = 2.R r.sin-1 (2.R r / b a)
Note: angle from sin-1(function) is in radians...

Speed Factor for Bending
This is a metric conversion from an imperial formula..
X b.2 = speed factor for bending = 0,521(V) -0,2

V= Pitch circle velocity =0,00005236*d 2.n 2 (m/s)

The table below is derived from a graph in BS 721. I cannot see how this works as a small worm has a smaller diameter compared to a large worm and a lower speed which is not reflected in using the RPM.

Table of speed factors for bending

RPM (n2) X b.2 RPM (n2) X b.2
1 0,62 600 0,3
10 0,56 1000 0,27
20 0,52 2000 0,23
60 0,44 4000 0,18
100 0,42 6000 0,16
200 0,37 8000 0,14
400 0,33 10000 0,13

Additional factors
The formula for the acceptable torque for wear should be modified to allow additional factors which affect the Allowable torque M c
M c2 = M c. Z L. Z M.Z R / K C


The torque on the wormwheel as calculated using the duty requirements (M e) must be less than the acceptable torque M c2 for a duty of 27000 hours with uniform loading.   For loading other than this then M e should be modified as follows
M e2 = M e. K S* K H
Thus
uniform load < 27000 hours (10 years) M e ≤ M c2
Other conditions M e2 ≤ M c2

Factors used in equations
Lubrication (Z L)..
Z L = 1 if correct oil with anti-scoring additive else a lower value should be selected

Lubricant (Z M)..
Z L = 1 for Oil bath lubrication at V s < 10 m /s
Z L = 0,815 Oil bath lubrication at 10 m/s < V s < 14 m /s
Z L = 1 Forced circulation lubrication

Surface roughness (Z R ) ..
Z R = 1 if Worm Surface Texture < 3μ m and Wormwheel < 12 μ m
else use less than 1

Tooth contact factor (K C
This relates to the quality and rigidity of gears . Use 1 for first estimate
K C = 1 For grade A gears with > 40% height and > 50% width contact
= 1,3 - 1,4 For grade A gears with > 30% height and > 35% width contact
= 1,5-1,7 For grade A gears with > 20% height and > 20% width contact

Starting factor (K S) ..
K S =1 for < 2 Starts per hour
=1,07 for 2- 5 Starts per hour
=1,13 for 5-10 Starts per hour
=1,18 more than 10 Starts per hour


Time / Duty factor (K H) ..
K H for 27000 hours life (10 years) with uniform driver and driven loads
For other conditions see table below
Tables for use with BS 721 equations

Speed Factors
X c.2 = K V .K R
Note: This table is not based on the graph in BS 721-2 (figure 7) it is based on another more easy to follow graph.   At low values of sliding velocity and RPM it agrees closely with BS 721.   At higher speed velocities it gives a lower value (e.g at 20m/s -600 RPM the value from this table for X c.2 is about 80% of the value in BS 721-2

Table of Worm Gear Speed Factors


Note -sliding speed = Vs and Rotating speed = n2 (Wormwheel)
Sliding speed K V Rotating Speed K R
m/s rpm
0 1 0,5 0,98
0,1 0,75 1 0,96
0,2 0,68 2 0,92
0,5 0,6 10 0,8
1 0,55 20 0,73
2 0,5 50 0,63
5 0,42 100 0,55
10 0,34 200 0,46
20 0,24 500 0,35
30 0,16 600 0,33





Stress Factors

Table of Worm Gear Stress Factors
Other metal
(Worm)
P.B.     C.I.     0,4%
C.Steel
0,55%
C.Steel
C.Steel 
Case. H'd
Metal
(Wormwheel)
Bending
bm )
Wear ( σ cm )
MPa
MPa
Phosphor Bronze
Centrifugal cast
69 8,3 8,3 9,0 15,2
Phosphor Bronze
Sand Cast Chilled
63 6,2 6,2 6,9 12,4
Phosphor Bronze
Sand Cast
49 4,6 4,6 5,3 10,3
Grey Cast Iron 40 6,2 4,1 4,1 4,1 5,2
0,4% Carbon steel 138 10,7 6,9
0,55% Carbon steel 173 15,2 8,3
Carbon Steel
(Case hardened)
276 48,3 30,3 15,2





Zone Factor (Z)
If b a < 2,3 (q +1)1/2 Then Z = (Basic Zone factor ) . b a /2 (q +1)1/2
If b a > 2,3 (q +1)1/2 Then Z = (Basic Zone factor ) .1,15


Table of Basic Zone Factors
q
z1 6 6,5 7 7,5 8 8,5 9 9,5 10 11 12 13 14 17 20
1 1,045 1,048 1,052 1,065 1,084 1,107 1,128 1,137 1,143 1,16 1,202 1,26 1,318 1,402 1,508
2 0,991 1,028 1,055 1,099 1,144 1,183 1,214 1,223 1,231 1,25 1,28 1,32 1,36 1,447 1,575
3 0,822 0,89 0,989 1,109 1,209 1,26 1,305 1,333 1,35 1,365 1,393 1,422 1,442 1,532 1,674
4 0,826 0,83 0,981 1,098 1,204 1,701 1,38 1,428 1,46 1,49 1,515 1,545 1,57 1,666 1,798
5 0,947 0,991 1,05 1,122 1,216 1,315 1,417 1,49 1,55 1,61 1,632* 1,652 1,675 1,765 1,886
6 1,131 1,145 1,172 1,22 1,287 1,35 1,438 1,521 1,588 1,625 1,694 1,714 1,733 1,818 1,928
7 1,316 1,34 1,37 1,405 1,452 1,54 1,614 1,704 1,725 1,74 1,76 1,846 1,98
8 1,437 1,462 1,5 1,557 1,623 1,715 1,738 1,753 1,778 1,868 1,96
9 1573 1,604 1,648 1,72 1,743 1,767 1,79 1,88 1,97
10 1,68 1,728 1,748 1,773 1,798 1,888 1,98
11 1,732 1,753 1,777 1,802 1,892 1,987
12 1,76 1,78 1,806 1,895 1,992
13 1,784 1,806 1,898 1,998
14 1,811 1,9 2





Duty Factor

Duty - time Factor K H
Impact from Prime mover Expected life
hours
K H
Impact From Load
Uniform Load Medium Impact Strong impact
Uniform Load
Motor Turbine Hydraulic motor
1500 0,8 0,9 1
5000 0,9 1 1,25
27000 1 1,25 1,5
60000 1,25 1,5 1,75
Light impact
multi-cylinder engine
1500 0,9 1 1,25
5000 1 1,25 1,5
27000 1,25 1,5 1,75
60000 1,5 1,75 2
Medium Impact
Single cylinder engine
1500 1 1,25 1,5
5000 1,25 1,5 1,75
27000 1,5 1,75 2
60000 1,75 2 2,25





Worm q value selection
The table below allows selection of q value which provides a reasonably efficient worm design.   The recommended centre distance value "a" (mm)is listed for each q value against a range of z 2 (teeth number values).   The table has been produced by reference to the relevant plot in BS 721
Example
If the number of teeth on the gear is selected as 45 and the centre distance is 300 mm then a q value for the worm would be about 7.5

Important note: This table provides reasonable values for all worm speeds. However at worm speeds below 300 rpm a separate plot is provided in BS721 which produces more accurate q values.    At these lower speeds the resulting q values are approximately 1.5 higher than the values from this table. The above example at less than 300rpm should be increased to about 9

Table for optimum q value selection
Number of Teeth On Worm Gear (z 2)
q 20 25 30 35 40 45 50 55 60 65 70 75 80
6 150 250 380 520 700
6.5 100 150 250 350 480 660
7 70 110 170 250 350 470 620 700
7.5 50 80 120 180 240 330 420 550 670
8 25 50 80 120 180 230 300 380 470 570 700
8.5 28 90 130 130 180 220 280 350 420 500 600 700
9 40 70 100 130 170 220 280 330 400 450 520
9.5 25 50 70 100 120 150 200 230 300 350 400
10 26 55 80 100 130 160 200 230 270 320
11 25 28 55 75 100 130 150 180 220 250
12 28 45 52 80 100 130 150 100
13 27 45 52 75 90 105



AGMA method of Designing Worm Gears
The AGMA method is provided here because it is relatively easy to use and convenient- AGMA is all imperial and so I have used conversion values so all calculations can be completed in metric units..


Good proportions indicate that for a centre to centre distance = C the mean worm dia d 1 is within the range
Imperial (inches)
( C 0,875 / 3 )     d 1      ( C 0,875 / 1,6 )
Metric ( mm)
( C 0,875 / 2 )     d 1      ( C 0,875 / 1,07 )


The acceptable tangential load (W t) all
(W t) all = C s. d 20,8 .b a .C m .C v . (0,0132) (N)
The formula will result in a life of over 25000 hours with a case hardened alloy steel worm and a phosphor bronze wheel

C s = Materials factor
b a = Effective face width of gearwheel = actual face width. but not to exceed 0,67 . d 1
C m = Ratio factor
C v = Velocity factor

Modified Lewis equation for stress induced in worm gear teeth .
σ a = W t / ( p n. b a. y )(N)
W t = Worm gear tangential Force (N)
y = 0,125 for a normal pressure angle α n = 20o

The friction force = W f
W f = f.W t / (. cos φ n ) (N)
γ = worm lead angle at mean diameter
α n = normal pressure angle

The sliding velocity = V s
V s = π .n 1. d 1 / (60,000 )
d 1 = mean dia of worm (mm)
n 1 = rotational speed of worm (revs/min)

The torque generated γ at the worm gear = M b (Nm)
T G = W t .d 1 / 2000


The required friction heat loss from the worm gearbox
H loss = P in ( 1 - η )
η = gear efficiency as above.


C s values


C s = 270 + 0,0063(C )3... for C ≤ 76mm ....Else

C s (Sand cast gears ) = 1000 for d 1 ≤ 64 mm ...else... 1860 - 477 log (d 1 )

C s (Chilled cast gears ) = 1000 for d 1 ≤ 200 mm ...else ... 2052 -456 log (d 1 )

C s (Centrifugally cast gears ) = 1000 for d 1 ≤ 635 mm ...else ... 1503 - 180 log (d 1 )



C m values

NG = Number of teeth on worm gear.
NW = Number of stards on worm gear.
mG = gear ration = NG /NW

C v values
C v (V s > 3,56 m/s ) = 0,659 exp (-0,2167 V s )

C v (3,56 m/s ≤ V s < 15,24 m/s ) = 0,652 (V s) -0,571 )

C v (V s > 15,24 m/s ) = 1,098.( V s ) -0,774 )


f values
f (V s = 0) = 0,15

f (0 < V s ≤ 0,06 m/s ) = 0,124 exp (-2,234 ( V s ) 0,645

f (V s > 0,06 m/s ) = 0,103 exp (-1,1855 ( V s ) ) 0,450 ) +0,012

http://www.roymech.co.uk/Useful_Tables/Drive/Worm_Gears.html 

Perhitungan Power Motor

  1. Perhitungan power pada umumnya,
    P = W/t dalam watt (J/s).
    Dengan,
    P = Power, watt
    W = Usaha, J
Perhitungan dari usaha, W
W= F x s, dalam Nm = J
Dengan,
W = Usaha, Nm
F = Gaya, N
s = Jarak, m
t = Waktu, s
   karena W= F x s,
   sehingga P = ( F x s)/ t , dalam watt (J/s)
   Perhitungan kecepatan
   v = s/t, dalam m/s,
   Sehingga P = F x v , dalam watt
   2. Perhitungan efficiency
       
η = Pout/Pin, dalam %
         dengan,
        ŋ = Efficiency, %

        Pout = power keluar, kW
        Pin = power masuk, kW