# The Surface Strength of Gears, Elements of Metric Gear Technology (Cont.)

**17.2 Surface Strength Of Spur And Helical Gears**

The following equations can be applied to both spur and helical gears, including double helical and internal gears, used in power transmission. The general range of application is:

Module:

*m*, 1.5 to 25 mm

Pitch Circle:

*d*, 25 to 3200 mm

Linear Speed:

*v*, less than 25 m/sec

Rotating Speed:

*n*, less than 3600 rpm

**17.2.1 Conversion Formulas**

To rate gears, the required transmitted power and torques must be converted to tooth forces. The same conversion formulas,

*Equations (17-1)*,

*(17-2)*and

*(17-3)*, of

**SECTION 17**are applicable to surface strength calculations.

**17.2.2 Surface Strength Equations**

As stated in

**SECTION 17.1**, the tangential force,

*F*, is not to exceed the allowable tangential force,

_{t}*F*. The same is true for the allowable Hertz surface stress, σ

_{t lim}*. The Hertz stress σ*

_{H lim}*is calculated from the tangential force,*

_{H}*F*. For an acceptable design, it must be less than the allowable Hertz stress σ

_{t}*. That is:*

_{H lim}The tangential force, σ

*, in kgf, at the standard pitch circle, can be calculated from*

_{H lim}*Equation (17-13)*.

The Hertz stress σ

*(kgf/mm²) is calculated from*

_{H}*Equation (17-14)*, where

*u*is the ratio of numbers of teeth in the gear pair.

The "+" symbol in Equations

*(17-13)*and

*(17-14)*applies to two external gears in mesh, whereas the "–" symbol is used for an internal gear and an external gear mesh. For the case of a rack and gear, the quantity u/(u ± 1) becomes 1.

#### 17.2.3 Determination Of Factors In The Surface Strength Equations

**17.2.3.A Effective Tooth Width,**

*b*(mm)_{H}The narrower face width of the meshed gear pair is assumed to be the effective width for surface strength. However, if there are tooth modifications, such as chamfer, tip relief or crowning, an appropriate amount should be subtracted to obtain the effective tooth width.

**17.2.3.B Zone Factor,**

*Z*_{H}The zone factor is defined as:

where: β

*= tan*

_{b}^{-1}(tan β cos α

*)*

_{t}The zone factors are presented in

**Figure 17-2**for tooth profiles per JIS B 1701, specified in terms of profile shift coefficients

*x*and

_{1}*x*, numbers of teeth

_{2}*z*and

_{1}*z*and helix angle β.

_{2}The "+" symbol in

**Figure 17-2**applies to external gear meshes, whereas the "–" is used for internal gear and external gear meshes.

**17.2.3.C Material Factor,**

*Z*_{M}where:

ν = Poisson's Ratio, and

*E*= Young's Modulus

**Table 17-9**contains several combinations of material and their material factor.

**17.2.4 Contact Ratio Factor,**

*Z*_{ε}

This factor is fixed at 1.0 for spur gears.

For helical gear meshes,

*Z*

_{ε}is calculated as follows:

**17.2.5 Helix Angle Factor,**

*Z*_{β}

This is a difficult parameter to evaluate. Therefore, it is assumed to be 1.0 unless better information is available.

**17.2.6 Life Factor, K**

_{HL}This factor reflects the number of repetitious stress cycles. Generally, it is taken as 1.0. Also, when the number of cycles is unknown, it is assumed to be 1.0. When the number of stress cycles is below 10 million, the values of

**Table 17-10**can be applied.

**17.2.7 Lubricant Factor,**

*Z*_{L}The lubricant factor is based upon the lubricant's kinematic viscosity at 50°C. See

**Figure 17-3**.

**17.2.8 Surface Roughness Factor,**

*Z*_{R}This factor is obtained from

**Figure 17-4**on the basis of the average roughness

*R*(µm). The average roughness is calculated by

_{maxm}*Equation (17-19)*using the surface roughness values of the pinion and gear,

*R*and

_{max1}*R*, and the center distance,

_{max2}*a*, in mm.

**17.2.9 Sliding Speed Factor,**

*Z*_{V}This factor relates to the linear speed of the pitch line. See

**Figure 17-5**.

**17.2.10 Hardness Ratio Factor,**

*Z*_{W}The hardness ratio factor applies only to the gear that is in mesh with a pinion which is quenched and ground. The ratio is calculated by

*Equation (17-20)*.

where: HB

_{2}= Brinell hardness of gear range: 130 ≤ HB

_{2}≤ 470

If a gear is out of this range, the

*Z*is assumed to be 1.0.

_{W}**17.2.11 Dimension Factor,**

*K*_{HX}Because the conditions affecting this parameter are often unknown, the factor is usually set at 1.0.

**17.2.12 Tooth Flank Load Distribution Factor,**

*KH*_{β}

**(a) When tooth contact under load is not predictable:**This case relates the ratio of the gear face width to the pitch diameter, the shaft bearing mounting positions, and the shaft sturdiness. See

**Table 17-11**. This attempts to take into account the case where the tooth contact under load is not good or known.

**(b) When tooth contact under load is good:**In this case, the shafts are rugged and the bearings are in good close proximity to the gears, resulting in good contact over the full width and working depth of the tooth flanks. Then the factor is in a narrow range, as specified below:

**17.2.13 Dynamic Load Factor,**

*K*_{V}Dynamic load factor is obtainable from

**Table 17-3**according to the gear's precision grade and pitch line linear speed.

**17.2.14 Overload Factor,**

*K*_{o}The overload factor is obtained from either

*Equation (17-11)*or from

**Table 17-4**.

**17.2.15 Safety Factor For Pitting,**

*S*_{H}The causes of pitting involves many environmental factors and usually is difficult to precisely define. Therefore, it is advised that a factor of at least 1.15 be used.

**17.2.16 Allowable Hertz Stress,**σ

_{H lim}

The values of allowable Hertz stress for various gear materials are listed in Tables

**17-12**,

**17-2 part 2**,

**17-13**,

**17-14**,

**17-14A**,

**17-15**,

**17-16**,

**17-16A**and

**17-16B**. Values for hardness not listed can be estimated by interpolation. Surface hardness is defined as hardness in the pitch circle region. Also see

**Figure 17-6**.

**17.3 Bending Strength Of Bevel Gears**

This information is valid for bevel gears which are used in power transmission in general industrial machines. The applicable ranges are:

Module:

*m*, 1.5 to 25 mm

Pitch Diameter:

*d*, less than 1600 mm for straight bevel gears, less than 1000 mm for spiral bevel gears

Linear Speed:

*v*, less than 25 m/sec

Rotating Speed:

*n*, less than 3600 rpm

**17.3.1 Conversion Formulas**

In calculating strength, tangential force at the pitch circle,

*F*, in kgf; power,

_{tm}*P*, in kW, and torque,

*T*, in kgf • m, are the design criteria. Their basic relationships are expressed in Equations

*(17-23)*through

*(17-25)*.

**17.3.2 Bending Strength Equations**

The tangential force,

*F*, acting at the central pitch circle should be equal to or less than the allowable tangential force,

_{tm}*F*, which is based upon the allowable bending stress σ

_{tm lim}_{F lim}. That is:

The bending stress at the root, σ

*, which is derived from*

_{F}*F*should be equal to or less than the allowable bending stress σ

_{tm}_{F lim}.

The tangential force at the central pitch circle,

*F*(kgf), is obtained from

_{tm lim}*Equation (17-28)*.

where: β

*: Central spiral angle (degrees)*

_{m}*m*: Radial module (mm)

*R*: Cone distance (mm)

_{a}And the bending strength σF (kgf/mm2) at the root of tooth is calculated from

*Equation (17-29)*.

#### 17.3.3 Determination of Factors in Bending Strength Equations

**17.3.3.A Tooth Width,**

*b*(mm)The term

*b*is defined as the tooth width on the pitch cone, analogous to face width of spur or helical gears. For the meshed pair, the narrower one is used for strength calculations.

**17.3.3.B Tooth Profile Factor,**

*Y*_{F}The tooth profile factor is a function of profile shift, in both the radial and axial directions. Using the equivalent (virtual) spur gear tooth number, the first step is to determine the radial tooth profile factor,

*Y*, from

_{FO}**Figure 17-8**for straight bevel gears and

**Figure 17-9**for spiral bevel gears. Next, determine the axial shift factor,

*K*, with

*Equation (17-33)*from which the axial shift correction factor,

*C*, can be obtained using

**Figure 17-7**. Finally, calculate

*Y*by

_{F}*Equation (17-30)*.

Should the bevel gear pair not have any axial shift, then the coefficient

*C*is 1, as per

**Figure 17-7**. The tooth profile factor,

*Y*, per

_{F}*Equation (17-31)*is simply the

*Y*. This value is from

_{FO}**Figure 17-8**or

**17-9**, depending upon whether it is a straight or spiral bevel gear pair. The graph entry parameter values are per

*Equation (17-32)*.

where:

*h*= Addendum at outer end (mm)

_{a}*h*= Addendum of standard form (mm)

_{a0}*m*= Radial module (mm)

The axial shift factor,

*K*, is computed from the formula:

**17.3.3.C Load Distribution Factor,**

*Y*_{ε}

Load distribution factor is the reciprocal of radial contact ratio.

The radial contact ratio for a straight bevel gear mesh is:

See Tables

**17-17**,

**17-18**and

**17-19**for some calculating examples of radial contact ratio for various bevel gear pairs.

**17.3.3.D Spiral Angle Factor,**

*Y*_{β}

The spiral angle factor is a function of the spiral angle. The value is arbitrarily set by the following conditions:

**17.3.3.E Cutter Diameter Effect Factor,**

*Y*_{C}This factor of cutter diameter,

*Y*, can be obtained from

_{C}**Table 17-20**by the value of tooth flank length, b / cos β

*(mm), over cutter diameter. If cutter diameter is not known, assume*

_{m}*Y*= 1.00.

_{C}**17.3.3.F Life Factor,**

*K*_{L}We can choose a proper life factor,

*K*, from

_{L}**Table 17-2**similarly to calculating the bending strength of spur and helical gears.

**17.3.3.G Dimension Factor Of Root Bending Stress,**

*K*_{FX}This is a size factor that is a function of the radial module,

*m*. Refer to

**Table 17-21**for values.

**17.3.3.H Tooth Flank Load Distribution Factor,**

*K*_{M}Tooth flank load distribution factor,

*K*, is obtained from

_{M}**Table 17-22**or

**Table 17-23**.

**17.3.3.I Dynamic Load Factor,**

*K*_{V}Dynamic load factor,

*K*, is a function of the precision grade of the gear and the tangential speed at the outer pitch circle, as shown in

_{V}**Table 17-24**. Also see

**Table 17-24A**and

**Table 17-24B**.

**17.3.3.K Reliability Factor,**

*K*_{R}The reliability factor should be assumed to be as follows:

- General case:
*K*= 1.2_{R} - When all other factors can be determined accurately:
*K*= 1.0_{R} - When all or some of the factors cannot be known with certainty:
*K*= 1.4_{R}

**17.3.3.L Allowable Bending Stress at Root,**σ

_{F lim}

The allowable stress at root σ

_{F lim}can be obtained from

**Tables 17-5**through

**17-8**, similar to the case of spur and helical gears.

**17.4 Surface Strength Of Bevel Gears**

This information is valid for bevel gears which are used in power transmission in general industrial machines. The applicable ranges are:

Radial Module:

*m*, 1.5 to 25 mm

Pitch Diameter:

*d*, Straight bevel gear under 1600 mm

Spiral bevel gear under 1000 mm

Linear Speed:

*v*, less than 25 m/sec

Rotating Speed:

*n*, less than 3600 rpm

**17.4.1 Basic Conversion Formulas**

The same formulas of

**SECTION 17.3**apply.

**17.4.2 Surface Strength Equations**

In order to obtain a proper surface strength, the tangential force at the central pitch circle,

*F*, must remain below the allowable tangential force at the central pitch circle,

_{tm}*F*

_{tm lim}, based on the allowable Hertz stress σ

_{F lim}.

Alternately, the Hertz stress σ

_{H}, which is derived from the tangential force at the central pitch circle must be smaller than the allowable Hertz stress σ

_{H lim}.

The allowable tangential force at the central pitch circle,

*F*

_{tm lim}, in kgf can be calculated from

*Equation (17-39)*.

The Hertz stress, σ

_{H}(kgf/mm

_{2}) is calculated from

*Equation (17-40)*.

#### 17.4.3 Determination of Factors In Surface Strength Equations

**17.4.3.A Tooth Width,**

*b*(mm)This term is defined as the tooth width on the pitch cone. For a meshed pair, the narrower gear's "

*b*" is used for strength calculations.

**17.4.3.B Zone Factor,**

*Z*_{H}The zone factor is defined as:

If the normal pressure angle α

*is 20°, 22.5° or 25°, the zone factor can be obtained from*

_{n}**Figure 17-10**.

**17.4.3.C Material Factor,**

*Z*_{M}The material factor,

*Z*, is obtainable from

_{M}*Table 17-9*.

**17.4.3.D Contact Ratio Factor, Z**

_{ε}

The contact ratio factor is calculated from the equations below.

where: ε

_{α}= Radial Contact Ratio

ε

_{β}= Overlap Ratio

**17.4.3.E Spiral Angle Factor,**

*Z*_{β}Little is known about these factors, so usually it is assumed to be unity.

**17.4.3.F Life Factor,**

*K*_{HL}The life factor for surface strength is obtainable from

*Table 17-10*.

**17.4.3.G Lubricant Factor,**

*Z*_{L}The lubricant factor,

*Z*, is found in

_{L}*Figure 17-3*.

**17.4.3.H Surface Roughness Factor,**

*Z*_{R}The surface roughness factor is obtainable from

**Figure 17-11**on the basis of average roughness,

*R*, in µ m. The average surface roughness is calculated by

_{maxm}*Equation (17-44)*from the surface roughnesses of the pinion and gear (

*R*and

_{maxm1}*R*), and the center distance,

_{maxm2}*a*, in mm.

**17.4.3.I Sliding Speed Factor,**

*Z*_{V}The sliding speed factor is obtained from

*Figure 17-5*based on the pitch circle linear speed.

**17.4.3.J Hardness Ratio Factor,**

*Z*_{W}The hardness ratio factor applies only to the gear that is in mesh with a pinion which is quenched and ground. The ratio is calculated by

*Equation (17-45)*.

where Brinell hardness of the gear is: 130 ≤ HB2 ≤ 470

If the gear's hardness is outside of this range,

*Z*is assumed to be unity.

_{W}**17.4.3.K Dimension Factor,**

*K*_{HX}Since, often, little is known about this factor, it is assumed to be unity.

**17.4.3.L Tooth Flank Load Distribution Factor,**

*K*_{H}_{β}

Factors are listed in

**Tables 17-25**and

**17-26**. If the gear and pinion are unhardened, the factors are to be reduced to 90% of the values in the table. Also see

**Tables 17-26A**and

**17-26B**.

**17.4.3.M Dynamic Load Factor,**

*K*_{V}The dynamic load factor can be obtained from

*Table 17-24*.

**17.4.3.N Overload Factor,**

*K*_{0}The overload factor can be computed by

*Equation 17-11*or found in

*Table 17-4*.

**17.4.3.O Reliability Factor,**

*C*_{R}The general practice is to assume

*C*to be at least 1.15.

_{R}**17.4.3.P Allowable Hertz Stress,**σ

_{H lim}

The values of allowable Hertz stress are given in

*Tables 17-12*through

*17-16*.

**17.5 Strength Of Worm Gearing**

This information is applicable for worm gear drives that are used to transmit power in general industrial machines with the following parameters:

Axial Module:

*m*, 1 to 25 mm

_{x}Pitch Diameter of Worm Gear:

*d*, less than 900 mm

_{2}Sliding Speed:

*v*, less than 30 m/sec

_{s}Rotating Speed, Worm Gear:

*n*, less than 600 rpm

_{2}**17.5.1 Basic Formulas:**

**17.5.2 Torque, Tangential Force and Efficiency**

(1) Worm as Driver Gear (Speed Reducing)

where:

*T*= Nominal torque of worm gear (kg • m)

_{2}*T*= Nominal torque of worm (kgf • m)

_{1}*F*= Nominal tangential force on worm gear's pitch circle (kgf)

_{t}*d*= Pitch diameter of worm gear (mm)

_{2}*u*= Teeth number ratio = z2 /zw

η

*= Transmission efficiency, worm driving (not including bearing loss, lubricant agitation loss, etc.)*

_{R}µ= Friction coefficient

(2) Worm Gear as Driver Gear (Speed Increasing)

where: η

*I*= Transmission efficiency, worm gear driving (not including bearing loss, lubricant agitation loss, etc.)

**17.5.3 Friction Coefficient,**µ

The friction factor varies as sliding speed changes. The combination of materials is important. For the case of a worm that is carburized and ground, and mated with a phosphorous bronze worm gear, see

**Figure 17-12**. For some other materials, see

**Table 17-27**.

For lack of data, friction coefficient of materials not listed in

**Table 17-27**are very difficult to obtain. H.E. Merritt has offered some further information on this topic. See

**Reference 9**.

**17.5.4 Surface Strength of Worm Gearing Mesh**

(1) Calculation of Basic Load

Provided dimensions and materials of the worm pair are known, the allowable load is as follows:

(2) Calculation of Equivalent Load

The basic load

*Equations (17-51)*and

*(17-52)*are applicable under the conditions of no impact and the pair can operate for 26000 hours minimum. The condition of "no impact" is defined as the starting torque which must be less than 200% of the rated torque; and the frequency of starting should be less than twice per hour.

An equivalent load is needed to compare with the basic load in order to determine an actual design load, when the conditions deviate from the above.

Equivalent load is then converted to an equivalent tangential force,

*F*, in kgf:

_{te}and equivalent worm gear torque,

*T*, in kgf • m:

_{2e}(3) Determination of Load

Under no impact condition, to have life expectancy of 26000 hours, the following relationships must be satisfied:

For all other conditions:

**NOTE:**If load is variable, the maximum load should be used as the criterion.

#### 17.5.5 Determination of Factors in Worm Gear Surface Strength Equations

**17.5.5.A Tooth Width of Worm Gear,**

*b*(mm)_{2}Tooth width of worm gear is defined as in

*Figure 17-13*.

**17.5.5.B Zone Factor,**

*Z*where: Basic Zone Factor is obtained from

**Table 17-28**.

**» Design of Plastic Gears - Continued on page 10**

Section 1: Introduction to Metric Gears Section 2: Introduction to Gear Technology Section 3: Details of Involute Gearing Section 4: Spur Gear Calculations Section 5: Internal Gears Section 6: Helical Gears Section 7:

Screw Gear or Crossed Helical Gear Meshes Section 8: Bevel Gearing Section 9: Worm Mesh Section 10: Tooth Thickness Section 11: Contact Radio Section 12: Gear Tooth Modications Section 13: Gear Trains Section 14: Backlash Section 15: Gear Accuracy Section 16: Gear Forces Section 17: Strength and Durability of Gears Section 18: Design of Plastic Gears Section 19: Features of Tooth Surface Contact Section 20: Lubrication of Gears Section 21: Gear Noise References and Literature of General Interest