Page 211 - Improving Machinery Reliability

182 Improving Machinery Reliability

ing velocity. Softer gear teeth would comply better with certain misalignment or tooth
inaccuracies. Of course, they would not have the same basic strength as harder gears.
Procurement of gears designed for acceptable long-term operation with medium-
hard teeth is advantageous because future uprates may be possible by simply pur-
chasing replacement gears with greater tooth hardness.

Design Appraisals Shortcuts

Vendor experience with the design and fabrication of special-purpose gearing is no
less important than vendor experience with any other critically important machinery
category. Questions to be asked relate to pitch-line velocities, gear-blank (web) con-
struction, horsepower levels, bearing design, gear-speed ratios, etc. When vendor expe-
rience has been established to the review engineer's satisfaction, he is ready to proceed
with a comparison of competing bids. This comparison is aimed at determining which
of the various offers may represent a stronger, potentially less failure-prone gear.
Design appraisals can be complex and time consuming if efforts are made to use
the full complement of AGMA (American Gear Manufacturers Association) rating
formulas. Moreover, cycles to failures calculated with some of these rating formulas
can be drastically influenced by minor changes in the assumed or anticipated surface
roughness, tooth spacing, etc. A sensible approach to gear design appraisals would
not, therefore, use calculated probable cycles to failure in an absolute way. The
review engineer would utilize the data only to make a comparison of competing
offers and to assign a ranking order.
In the late 1960s, Robert H. Pearson?l then chief engineer of the Sier-Bath Gear
Company, equated the mathematical expression for estimated gear-tooth compres-
N, = 3.8 x lo-'' [ dFI(~d3fO)z r'77
sive stress to that for allowable fatigue stress. Compressive stress is a measure of
surface durability and pitting. Pearson's work, summarized in an article published by
Machine Design magazine in 1968 determined

In this expression

N, = life in cycles to failure
d = pinion operating pitch diameter, in
F = face width, in
H =hardness, bhn
I = geometry factor
W, =tangential driving force, Ibs
The geometry factor I is obtained by dividing durability factor C3 (obtained from Figure
3-73) by a materials factor (s,,/C,)~ (obtained from Figure 3-74). W, is readily calculat-
ed by dividing the pinion output torque by the pinion pitch radius. Cd, however, is a
good deal more difficult to obtain. Five factors make up Cd and are defined as follows: