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
Typical wear life for power screws and nuts

Wear Mechanism

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
Microscopic schematic of mating wear surfaces before break-in wear
Figure #39
Microscopic schematic of mating wear surfaces before break-in wear

Screw and Nut Material Selection

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

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

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

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
Static load capacity of weakest mating material

Life Testing Methods

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

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
nut threads

Wear Life - Power Screws

The wear life of power screws is a function of load, speed, lubrication, contamination, heat and other factors. The operating loads listed in the Screw/Nut Engineering section for each screw series provide acceptable wear life for most applications.

Wear in a power screw is generally in proportion to usage. Each movement of the screw surface against the mating nut surface removes a microscopic amount of material, usually from the softer nut material. As these wear increments add up over time, and backlash increases, the nut threads become thinner. When the shear strength of the remaining threads is exceeded by the load, failure occurs.

Although their wear life is not as predictable as Ballscrews, well lubricated power screws, without side loads or moment loads, can provide excellent service lives for many applications. Heavy loads and duty cycles which generate significant amounts of heat will cause material and lubricant breakdown and should be avoided.

Every power screw application is unique in terms of loads, environment, duty cycle, etc. Operational and life testing of prototypes is highly recommended especially for OEMs anticipating large volume production. Customers are encouraged to contact Roton’s application engineers who are available for consultation and to discuss wear life objectives for specific applications. Often, a short evaluation early in the application development can save many hours of design revision and testing.

Wear Life - Ballscrews

The wear life of Ballscrews is much more predictable than power screws due to the large body of research and testing that has been conducted on ball bearings and bearing balls. Assuming that a Ballscrew is a ball bearing arranged with helical inner and outer races, the listed operating loads have been determined.

The operating load ratings are based upon a theoretical 90% survival rate of Ballscrews at 1,000,000 in. of travel. Ratings also assume pure axial loading of the screw and nut with no side loads or moment loads, and a clean, well lubricated, room temperature environment. The presence of unfavorable loading, dirt, dust, lack of lubricant and external heat will dramatically reduce the service life.

Ballscrew life is proportional to the inverse cube of the load. If the load is cut in half, the life increases by 2 cubed or a factor of 8. For example, a 1 x .250 Ballscrew is to be operated at 1,000 lbs. The expected travel life of the Ballscrew with a 90% survival rate would be 4,100,000 inches of travel. Dividing the operating load rating of 1,600 lbs. for this size Ballscrew (from Table 15) by the actual load of 1,000 lbs., cubing the result and multiplying by 1,000,000 inches yields the expected life: (1,600/1,000)3 x 1,000,000 = 4,100,000 inches. The formula for Ballscrew wear life can be found in Useful Formulas.

Every application is unique in terms of loads, environment, duty cycle, etc. Operational and life testing of prototype Ballscrews is recommended especially for OEMs anticipating large volume production. Customers are encouraged to contact Roton’s application engineers who are available for consultation and to discuss wear life objectives for specific applications.

Cost Considerations

The products in this catalog are arranged in increasing cost order from front to back. Acmes are the least expensive and are the most widely used. Hileads®, Torqsplines® and Ballscrews offer increased performance at increased costs.

The final choice depends upon the user’s economics, the market for the end product, reliability objectives, and many other factors. Bear in mind that initial cost is only one element in the cost equation. Installed cost, maintenance, consequences of failure and many other items need to be weighed before finalizing any design.