Maximizing Bearing Life with EHD Lubricants

Sibtain Hamid, Santotrac Fluids
Tags: bearing lubrication

Early theories assumed that metal-to-metal contact was the mechanism that transmitted force in rolling contact, with the lubricant acting primarily as a coolant. But experimental measurements of the interfacial film thickness have shown that a distinct lubrication regime exists between hydrodynamic and boundary lubrication. This lubrication regime is termed the elastohydrodynamic (EHD) regime because of the presence of a lubricant film and because elastic deformations occur at the metal surface.

Today the EHD property that many lubricants exhibit under high stress is well known. Basically, for a few microseconds the fluid lubricant becomes a semisolid and conforms to the changing geometry of the metal surface which is itself slightly deformed. This brief transformation enhances protection between metal surfaces, limits wear on the surface of a ball or roller bearing, and prevents fatigue failure (Figure 1).


Figure 1. EHD Property Limits
Wear on the Surface of Ball
or Roller Bearing

The thickness of the EHD film depends on the viscosity of the lubricant - thicker films generally give more protection. But if the viscosity becomes too high, the bearing elements begin to skid, drag and overheat. The optimum lubricant is one that provides a viscosity high enough to put an adequate film thickness between the asperities on the opposing surfaces while at the same time keeping viscosity low enough to prevent skidding, dragging and overheating.

Bearing life is conventionally measured by exposing a set of bearings to identical test conditions and determining the point at which a given percentage of the bearings may be expected to fail. L10, for example, indicates the point at which 10 percent of a specified bearing type will fail under described conditions. While L10 is the most frequently used parameter, other guides such as L50 and even L1 are sometimes used.

In L10 tests, the basic dynamic capacity of a bearing is usually measured at one million revolutions. Recently, however, improvements in steel manufacturing and surface treatment have allowed ball and roller bearing manufacturers to use standards as high as three million revolutions.

The basic dynamic capacity (the fatigue life) of a given ball bearing can be predicted by the equation:

Capacity = K x N 2/3 x d 1.8 (or: K x N2/3 x d1.4)

Where K is a constant based on the design and material of the bearing, N is the number of balls, and d is the diameter of the ball. This equation is used with bearings having a diameter less than 25.4 mm (1 inch); for larger bearings, the d exponent is changed to 1.4. These equations are often further modified to accommodate specific situations.

In simpler terms, the basic equation for dynamic capacity would be:

Capacity = (bearing life x load3 x RPM x 60/3,000,000)1/3

The above equations assume that bearing failure will be caused by fatigue alone. In actual use, bearings can fail for many other reasons, including overheating, insufficient lubrication, mechanical misalignment, dirt, vibration, fretting, cage wear and excessive preload or shock loads during operation. The equations given here are best used as approximations of basic dynamic capacity.

Fatigue failure is related to the degree of stress and the number of stress cycles, but is initiated by minor imperfections in the bearing material. In a specific bearing application, the number of revolutions before failure is related to the location and size of the load stress points, the thickness of the film at these points and mechanisms that cause film loss. When failure finally occurs, it typically begins as a fatigue crack at a point of stress concentration.

Lubrication conditions at the areas constantly in contact therefore have great impact on the number of revolutions before failure. Bearing finish also plays a role, because the finish interacts with the fluid properties of the lubricant and the thickness of the film.

Bearing finish is significant during idle mode because minute localized vibrations can cause fretting of the surface. When rotation begins, the film develops and the properties of the bearing finish become less important. Under high-speed rotation, part of the load is borne by the fluid ahead of the contact area. At the same time, asperities function to some degree to drag the lubricant into the contact area to create the lubricant pad that carries the rest of the load.

In real-world use, of course, a lubricant faces changing conditions of temperature, stress and exposure to corrosion. The development of a synthetic lubricant therefore focuses on satisfying six key conditions:

  1. The lubricant must be thermally stable at the maximum operating temperature.
  2. It must form an adequate film between the sliding and rolling contact areas.
  3. It must serve as a coolant.
  4. It must prevent corrosion both at the maximum operating temperature and in idle mode.
  5. It must be free from dirt and contaminants.
  6. It must protect bearings from normal wear.

Initial investigations involved determination of film thickness of several commercially available base fluids.

Film thickness of a bearing is primarily a function of:

ohC=KH(OD-ID)0.32(N(OD+ID))0.68

Z0 0.68-GOD, OD, ID and hC in inches,

KH=3.8x10-11

hc is film thickness

The coefficient of traction was determined by a Twin-Disc machine designed by Santotrac to understand the EHD film formation under high speed rolling contact discs under high pressure of at least 300,000 psi.

This machine has two large hardened steel discs that run parallel to each other at high speeds and thus allow an elastohydrodynamic film to form. The commercial MTM machine available from PCS Instruments can be used. A total of six base fluids were tested: polyalphaolephin (PAO), polyol ester, diester, polyglycol, paraffinic mineral oil and cycloaliphatic hydrocarbon.

The last-mentioned base fluid was of particular interest because previous work (see reference list) had shown that it has a high coefficient of traction. Traction is the resistance to shear from external forces that are acting on a film separating rolling elements. A high coefficient of traction correlates well with thicker film formation, even under extreme and variable conditions.


Figure 2. Measured Coefficient of
Traction Values for Six Base Fluids

The results of the base fluid testing are shown in Figure 2. All six base fluids have a measurable coefficient of traction, but the cycloaliphatic hydrocarbon base fluid has a coefficient nearly double that of the second-place paraffinic mineral oil. It appeared likely that a bearing lubricant made from the cycloaliphatic hydrocarbon base fluid might have desirable properties. For instance, when the pressure on bearings in production machinery becomes great enough, the bearings actually operate in the EHD regime. At these pressures, conventional lubricants will shear down, while the cycloaliphatic hydrocarbon will resist shearing and thus prevent damage to the bearings.

Next, bearing lubricants were prepared using four of the six base fluids tested (polyalphaolephin, diester, paraffinic mineral oil and cycloaliphatic hydrocarbon). All four lubricants included the same additive package, and were formulated to give a viscosity of ISO 46. Oxidation corrosion tests were carried out for these lubricants and wear properties were determined. The coefficients of traction were then measured. As one would expect, the additives had little impact on the coefficients of traction. Specifically, the coefficient of traction of the cycloaliphatic hydrocarbon, and presumably its film thickness, remained nearly twice that of the other base fluids.

To determine the performance of these newly formulated lubricants in bearings, fatigue life tests were carried out using a rolling contact testing machine (Figure 3). The machine has two disks made from GVM M-50 steel hardened to RC62. Each disk is 7.5 inches in diameter.


Figure 3. Coefficient of Traction Machine
(Twin-Disk Machine)

Pressure is applied to each disk by a hydraulic cylinder. For the tests using these four lubricants, pressure was maintained at 700,000 psi.

Results of testing the four lubricants are shown in Figure 4.


Figure 4. Results of B10 Fatigue Life Test

Spalling occurred at ranges from 1.3 million cycles (diester-based lubricant) to nearly five million cycles (cycloaliphatic hydrocarbon-based lubricant). The higher coefficient of traction of this lubricant creates a higher load-carrying capacity, but without causing skidding, dragging or overheating. This lubricant has since been added to the product line.

The fact that the cycloaliphatic hydrocarbon formulation tested here survived nearly five million revolutions before failure indicates that its higher traction coefficient creates a thicker film at load stress points. This in turn suggests that this formulation may permit higher bearing loads and will result in longer bearing life.

References

  1. Hamid, S. and Rose, N. “The Race for A Better CVT: Advanced Designs Use Solid-Liquid Transmission Fluid.” Lubes ‘N Greases, Vol. 9, Issue 5, May 2003.
  2. Herber, J. and Joaquim, M. “Traction Lubes Show Great Promise.” Lubrication Engineering, Volume 6, No. 12, December 1996.
  3. Kraus, C. “Bearing and Rolling Traction, Analysis and Design.” Excelermatic Inc., Austin, Texas, 1984.
  4. Santotrac Synthetic Traction Lubricants Bulletin. Findett Corporation, St. Charles, Mo.