Erwin V. Zaretsky

A NASA/Pratt & Whitney Life Prediction Cooperative Program

MMC Life System Development (phase I)

Nasa Technical Reports Server (Ntrs)

Emanuel T. Branzai

For nearly four decades it has been a practice in commercial and military aircraft application that rolling-element bearings removed at maintenance or overhaul be reworked and returned to service. The work presented extends previously reported bearing life analysis to consider the depth (Z(45)) to maximum shear stress (45) on stressed volume removal and the effect of replacing the rolling elements with a new set. A simple algebraic relationship was established to determine the L(10) life of bearing races subject to bearing rework. Depending on the extent of rework and based upon theoretical analysis, representative life factors (LF) for bearings subject to rework ranged from 0.87 to 0.99 the lives of new bearings. Based on bearing endurance data, 92 percent of the bearing sets that would be subject to rework would result in L(10) lives equaling and/or exceeding that predicted for new bearings with the remaining 8 percent having the potential to achieve the analytically predicted life of new bearings when one of the rings is replaced at rework.. The potential savings from bearing rework varies from 53 to 82 percent that of new bearings depending on the cost, size and complexity of the bearing.

Effect of Rolling Bearing Refurbishment and Restoration on Bearing Life and Reliability

Richard J. Parker

Rolling-element fatigue studies were conducted with VIM-VAR AISI M-50, EFR 18-4-1, and VAR 18-4-1. Groups of 12.7-mm (1/2-in.) diameter balls of each material were tested in the five-ball fatigue tester. Test conditions included a load of 1540 N (347 lbf) giving a maximum Hertz stress of 5520 MPa (800 000 psi), a shaft speed of 10 700 rpm, and a contact angle of 30 deg. Tests were run at a race temperature of 339 K (150 deg F) with a type II ester (MIL-L-23699) lubricant. The rolling-element fatigue life of VIM-VAR AISI M-50 was not significantly different from that of EFR 18-4-1 or VAR 18-4-1 based on a statistical comparison of the test results.

Rolling-element Fatigue Life of AISI M-50 and 18-4-1 Balls

Richard Jay Parker

Salvatore J. Grisaffe

The five-ball fatigue tester was used to study the behavior of nickel- bonded titanium carbide cermet and self-bonded silicon carbide balls under repeated stresses applied in rolling contact. Test conditions were 800 and 700 F, a contact angle of 200, and a shaft speed of 950 rpm with a mineral oil lubricant. The nickel-bonded titanium carbide - cermet balls were tested at maximum Hertz stresses from 400,000 to 550,000 psi, while the silicon carbide balls were tested at maximum Hertz stresses from 300,000 to 400,000 Psi. Failure appearance for both materials was unlike the fatigue pits found in bearing steels and a crystallized-glass ceramic. However, the pits were similar to those found in alumina balls (both hot pressed and cold pressed and sintered). A typical failure was a shallow eroded area less than 1 mil deep for the titanium carbide cermet and 3 to 5 mils for the silicon carbide. These failures evolved by progressing slowly from a very small pit to one spanning the running track width. Tests at 800 F with mineral oil lubrication over a range of stresses show that life varies inversely with stress to a power that ranges from 9.7 to 10.5 for the titanium carbide cermet and from 6.9 to 8.6 for the silicon carbide. The load capacity of the titanium carbide and silicon carbide balls was 3 and 1 percent, respectively, that of typical bearing steel balls, and 41 and 18 percent, respectively, that of hot-pressed alumina balls.

Surface Failure of Titanium Carbide Cermet and Silicon Carbide Balls in Rolling Contact at Temperatures to 2000 ̊F

David W. Reichard

Reduced-pressure Environment Effects on Rolling-element Fatigue Life with Super-refined Mineral Oil Lubricant

William J. Anderson

Thomas L. Carter

Effect of Hardness and Other Mechanical Properties on Rolling-contact Fatigue Life of Four High-temperature Bearing Steels

Residual stress measurements were made in the zone of resolved shearing stress on five SAE 52100 upper test ball specimens with an average Rockwell C hardness of 63.2. These upper test ball specimens were run against lower test balls of nominal Rockwell C hardnesses of 60, 62, 63, 65, and 66 until either ft conditions included an or both components failed because of fatigue. Test conditions included an average race temperature of 150 F, a maximum (Hertz) compressive stress of 800,000 psi, and a highly purified naphthenic mineral oil lubricant. Fatigue lives of the upper test balls were compared with measured residual stresses in the subsurface zone of resolved shearing stress. The compressive residual stress induced in the upper test ball during running is a function of delta H, the hardness of the lower test balls minus the upper test ball hardness An interrelation is indicated among differences in com- ponent hardness, induced compressive residual stress, and fatigue life. Measured values of compressive residual stress within the zone of resolved maximum shearing stress ranged from 178,000 and 294,000 psi. The apparent maximum residual stress occurs above where delta H = 0.

Effect of Component Differential Hardnesses on Residual Stress and Rolling-contact Fatigue

Eric N. Bamberger

Hans R. Signer

Operating Characteristics of 120-millimeter-bore Ball Bearings at 3x106 DN

Steven T. Miller

An experimental apparatus was designed and constructed at the NASA Lewis Research Center to study one aspect of spinning friction based on interfacial slip over the entire ball-race contact area. The apparatus is capable of measuring spinning moments of less than 0. Ol inch-pound at maximum Hertz contact stresses to over 400 000 psi, speeds to 3500 rpm, and under varying contact configurations and conditions. From the torque measurement, a coefficient of spinning friction can be calculated. The apparatus comprises a drive assembly, a dead-weight load assembly, a spherical upper test specimen, a cylindrically grooved lower test specimen of varying conformity, a lower test specimen housing assembly incorporating a hydrostatic air-bearing assembly, and a torque-measuring system. In operation, the upper test specimen is loaded against the lower test specimen through the drive shaft assembly by the dead- weight load assembly. As the drive assembly is rotated, the upper test specimen rotates against the stationary lower test specimen actuating the torque-measuring system. Preliminary tests were conducted under varying Hertz stress to 136 000-psi maximum Hertz stress at a spinning speed of 950 rpm, with a 51 percent ball-groove conformity using a polyphenyl ether (5P4E), a highly purified naphthenic mineral oil, and a di-2-ethylhexyl sebacate (MIL-7808). It was found that the coefficient of spinning friction decreased with increasing maximum Hertz stress to an intermediate stress level and had a minimum value of 0. 122, 0. 089, and 0. 050, respectively, for the lubricants listed previously.

Apparatus for Studying Ball Spinning Friction

Erwin V. Zaretsky

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