Power Screw & Nut Wear
Introduction back to top
The wear life of power screw and nut drive systems is difficult to predict theoretically. The number
of variables involved in such a prediction is large; load, speed, screw material, nut material,
surface finishes, lubrication, duty cycle, operating temperature, and environmental factors such as
the presence of abrasive contaminants, corrosives, vibration, etc... (See equation 1 in Figure 49)
and our understanding of how these factors interact is limited. Because of this, the only proper
approach to evaluating the service life of power screws is to thoroughly life test each application
prior to final specification and production. However, even under laboratory conditions, results may
vary quite markedly as rubbing friction and wear are notoriously capricious. Test life cycle variations
of two or three to one are not uncommon.
A general understanding of the wear mechanism, some simple design and operating guidelines, and
recommendations for life testing will help you get the best performance from your screw and nut drive
system.
FIGURE 37
Wear Mechanism back to top
The study of wear is a field called tribology. There is much research on the subject, but little
definitive work that can help determine the wear rate of two surfaces in any specific application.
The wear mechanism itself is simple to understand. With reference to Figure 38, two rubbing
surfaces contact only at their highest microscopic aspersions. When the contact stress is high
enough and under relative motion, these aspersions shear off and become debris. Lower aspersions
then come into contact and the contact area increases until the unit pressure and the underlying
materials shear strengths are in balance. At this point break-in wear has occurred and the
surfaces appear as in Figure 39, and can be represented by the curved line between A and B in
Figure 37.
After break-in a steady state, continuous wear pattern begins, as represented by the straight
line between points B and D in Figure 37. Unless the surfaces are completely separated with a
lubricant film some wear will occur continually as the mating surfaces rub each other in normal
service life.
FIGURE 38
FIGURE 39
Screw and Nut Material Selection back to top
To reduce the costs of wear in power screw systems, we recommend designs where nuts are made of
softer material and screws of harder material. This ensures that the nuts will wear and the screw
will remain relatively wear free, which is desirable because replacement nuts are usually much less
expensive than replacement screws. Typically bronze or plastic nuts are mated with carbon or
stainless steel screws.
In general, plastic nuts offer the best possibility for long life at low loads. They can be used
with minimal lubrication and at light pressures experience little wear. They have a lower coefficient
of friction and therefore run cooler, and require less drive torque for the same load compared with
metallic nuts. They outwear bronze for low load applications, usually reaching a plateau after a
wear-in period. In contrast, bronze and copper alloys tend to wear a certain amount under light
loads simply by the mechanism of two surfaces rubbing together like the wear on the steps of an
old building or water dripping on a rock for many years. However, bronze and copper alloys are the
preferred choice where loads are high, or heat build up is a concern.
Power screws and nuts made from the same materials make poor candidates for good wear life. Under
the pressure of loading as the screw and nut are rubbed together the molecules of the screw and nut
will bond with each other. The result is galling where material is rapidly transferred across the
screw / nut interface. This phenomenon is especially evident in steel on steel and stainless steel
on stainless steel. Run without lubricant or with poor lubrication in the presence of high loading
steel on steel can actually weld together in just a few cycles. Like materials are generally only
used in applications that position and support loads, such as scaffolding, jackstands, or mechanical
stops. For moving loads, bronze or plastic nuts should be used.
Design and Operational Considerations back to top
Here are the most important keys to maximizing service life:
- Maintain low surface contact pressure.
Increasing the screw size and nut size will reduce thread contact pressure for the same working
load. The higher the unit pressure and the higher the surface speed, the more rapid the wear will
be.
- Maintain low surface speed.
Increasing the screw lead will reduce the suface speed for the same linear speed.
- Keep the mating surfaces well lubricated.
The better the lubrication, the longer the service life. Power screws and nut should be treated
as any other wear surfaces. If grease fittings or other lubrication means are provided for other
wear elements in the application, the designer will be well served by providing a like means to
lubricate the power screw and nut.
- Keep the mating surfaces clean.
Dirt, especially hard particle type dirt, can easily embed itself in the soft nut material. Once
established the dirt will act as a file and readily abrade the mating screw surface. The soft nut
material backs away during contact leaving the hard dirt particles to scrap away the mating screw
material. Approximately 2/3 of the drive energy in an ACME screw and nut system goes into heat.
When the mating surfaces heat up, they become much softer and are more easily worn away. Means to
remove the heat such as limited duty cycles or heat sinks must be provided so that rapid wear of
over heated materials can be avoided.
Some applications and tests indicate that wear is proportional to load and speed, however, others
show proportionality to load and speed to the 2nd - 4th power. The general relationship of more wear
with higher loads and speeds is well accepted and has been demonstrated in laboratory and field tests.
Wear Equations back to top
As discussed earlier, estimating the service life of power screw systems is a very complex task and
inexact at best. The only reliable predictor is actual testing. Variations in service life are widely
scattered and multiples of two to one or three to one in life test performance are not uncommon. The
field of tribology is not as yet mature enough for accurate life estimating. The variables affecting
life are too great in number and can vary too widely preventing reliable performance predictions. To
illustrate the complexity of predicting wear theoretically, consider the potential equation for
predicting wear in Figure 40.
The exponents x thru j may be unity or greater than or less than unity and much work remains to be
done by tribologists to determine a workable formula which includes all variables known to influence
wear. In the meantime, the assumption commonly used for power screw and nut wear is that within the
mating materials' PV limits, wear is proportional to the operating PV. This is known as the "linear
assumption of wear" and while not completely true it is a useful estimator.
Using the linear assumption and with reference to Figure 41, two mating materials are assumed to have
a "PV LIMIT" which is constant except for the extremes of unit pressure and surface speed.
This relationship is represented by the graph in Figure 41. Below the PV limit curve, the linear wear
life assumption states that wear is proportional to the PV product. Thus, simplifying Equation 1 in
Figure 40, x and y both become unity and all the variables beyond V are eliminated. This yields the
following linear equation for wear life:
Wear = KPV Equation 2
In English units of measure for the above equation, Wear is the rate of wear in inch/hr, K is the
"wear factor" in in.3-min./ ft.-lb.-hr., P is the pressure in lbs./in2 (PSI),V is the rubbing
surface speed in ft./min. (FPM). Note that Wear from this equation is the slope of the linear section of
the graph in Figure 37.
FIGURE 40
Example Using Equation 2 back to top
Equation 2 may be more useful in relative application than in the absolute application because
"K" factors for common lubricated bronzes and lubricated plastics are not plentiful. Let€(tm)s
assume however that testing of a 3/4 - 6 ACME screw with lubricated bronze sleeve nut yields a service
life of 20,000 cycles for a 10-inch travel at 1,000 lbs. and 300 RPM. The target service life is 30,000
cycles. We could calculate the "K" factor from this test result, but that is not necessary to
evaluate the expected increase in service life by changing to a larger 1 - 5 Acme size with lubricated
bronze sleeve nut. For the 3/4 - 6 size, the speed V at 300 RPM is 59 FPM and using the thread contact
area of 1.414 sq. inches, the pressure P is 707 PSI at 1,000 lbs. load. This yields an operating PV of
41,713 PSIFPM.
For the larger ACME size 1 - 5 note that the speed can be reduced to 250 RPM to get the same linear
speed for the end use because the lead is .200 in./rev. versus .16667 in./rev. for the 3/4 - 6 size.
Using the same load of 1,000 lbs. but the larger contact area of 2.55 sq. inches, the pressure P becomes
392 PSI and the speed V at 250 RPM is now 65.6 FPM reducing the operating PV to 25,715 PSIFPM. The
theoretical expected life would then increase by 62% (41,713 / 25,715 = 1.62) to 32,400 and the target
service life would be achieved.
FIGURE 41
Life Testing Methods back to top
As mentioned above the best method of assessing performance and life of power screws and nuts is actual
field testing. This may present difficulties because of time constraints so accelerated lab testing is
often conducted instead. All lab testing should monitor drive torque and nut temperature along with some
method of regulating speed and load. Dead weight testers are popular and reliable. Hydraulic or
pneumatic load testers can also be used but load cells are recommended to detect unexpected variations
in the load pressure. The use of cooling fans is often necessary when artificially high duty cycles are
used to shorten the testing duration.
Monitoring Wear back to top
Most users measure the backlash in the screw and nut set both initially and on going during the life
test (see Figure 42). Doing so will most likely produce a classic "S curve" graph similar to that shown
schematically in Fig. 35. The section of the graph from point "A" to point "B" is the break-in section
and is characterized by rapid initial run-in of the screw and nut surfaces. Both screw and nut surfaces
are virgin at this point and some polishing of the surfaces occurs over a short period of time until the
system reaches point "B" and begins steady state or normal service wear. Normal service life continues
until some point "D" after which the wear becomes quite rapid until the screw and nut seize or one
member, usually the nut, fractures at point "E". It is good practice to use a factor of safety and use
point "C" as the practical service limit. Many labs continue testing thru point "D" to point "E" just to
discover what can happen if service life is extended beyond the practical limit.
Two points are noteworthy. The first is that the slope of the linear portion of the curve between points
"B" and "D" is wear over time (the "K" factor in equation 2). The second is
that point "C" is an arbitrary point of wear set by the judgment of the engineer or designer. It
may be an increase in the backlash, say a doubling of same, or it may be some portion, say 1/2 of the life
between points "B" and "C". The key is that at some point in time a decision must be
made as to what constitutes a practical service life for each application considering the factor of safety,
the failure mode and the consequences of failure.
FIGURE 42
