Rolling element bearings such as ball, roller and needle bearings are used in almost every kind of machine and device with rotating parts. They are currently the most widely used bearing.
Rolling element bearings typically include four parts: an inner ring, an outer ring, the balls or rollers and a cage or separator for separating the balls from one another. The balls in ball bearings are normally made of high carbon chromium steel. The balls are heat treated to high strength and hardness and the surfaces are ground and polished. Cylindrical roller bearings are usually made of case hardened steel.
Rolling element bearings are made in a wide variety of types and sizes. Regardless of their size or shape, rolling element bearings operate on the same basic principle of allowing low friction rotation of one member relative to the other while maintaining solid metal-to-metal contact between the two elements.
For a rotating shaft, relative rotation between shaft and bearing is usually prevented by mounting the inner ring with a press fit and securing it with a nut threaded on the shaft. Excessive interference of metal must be avoided in press-fits, or the stretching of the inner ring may decrease the small but necessary internal looseness of the bearing.
Conventionally, rolling element bearings are mounted to a fixed housing so that because there is little radial play. Although the outer ring, when the shaft rotates, is mounted more loosely than the inner ring, rotational creep between the ring and the housing should be prevented.
Compared to other bearings such as conventional journal bearings, rolling element bearings offer a number of advantages. These include low starting friction; the ability to support loads inclined at any angle in the transverse plane; the ability to support thrust components of loads; and low maintenance cost. In addition the bearings are easily replaced when worn out and require less axial space than for journal bearings.
There are, however, certain disadvantages associated with conventional rolling element bearing assemblies. The cost is typically higher, more radial space is generally required than with journal bearings and more noise is generated by ball bearings, especially after wear. In addition, rolling element bearings are more subject to fatigue failure and are more easily damaged by foreign matter. All rolling element bearings have a limited life, typically less than 20,000 hours depending on the application. Another disadvantage associated with rolling element bearings is that they have very little damping capability because of the metal-to-metal contact between elements. Thus, rolling element bearings are typically less well suited to overload and shock conditions. This is a significant drawback in high speed turbo machinery.
High speed equipment such as the compressor turbine in a jet engine and aeroderivative applications such as steam turbines, gas turbines and compressors must pass through several natural frequencies before reaching operating speed. When a system operates at its natural frequency or resonance, the system/rotor vibration amplitudes become large. These vibrations can be destructive or even catastrophic if not adequately dampened. Bearings with adequate damping characteristics limit or damp out the vibrations to allow the equipment to safely pass through the critical speeds. Likewise, smaller vibrations due to unbalance can be dampened by the bearings damping characteristics.
As previously noted, rolling element bearings where metal-to-metal contact exists have very little damping capability. Accordingly, when rolling element bearings are used in jet engines or other high speed turbo machinery they must be supported in complex expensive multipart assemblies which use a squirrel cage centering spring. Examples of such constructions are shown in the following U.S. Patents: U.S. Pat. No. 3,456,992 to Kulina disclosing fluid retained between sealing rings; U.S. Pat. No. 3,863,996 to Raimondi disclosing a fluid dampened journal bearing; U.S. Pat. No. 3,994,541 to Geary et al. disclosing a fluid dampened tilt pad bearing; U.S. Pat. No. 4,097,094 to Gardner disclosing a fluid dampened pad-type bearing and U.S. Pat. No. 4,213,661 to Marmol disclosing an O-ring type damper. Another form of damper was recently proposed by Messrs. Heshmat and Walton of Mechanical Technology Inc. These so-called multi-squeeze film dampers use a spiral foil to provide a spiral multi-film damper.
There are a number of disadvantages associated with known squeeze film damper bearings. Squeeze film dampers which use a squirrel cage centering spring typically occupy an axial space 2 to 3 times larger than the axial space available for the squeeze film land. Moreover, it is very difficult to install the centering spring and center the rotor within the squeeze film clearance. For this reason, the performance of the damper is often not consistent from one engine to another. The multi-piece design and precision required to assemble such an element is also undesirable.
In process type compressors, elastomer O-rings are used as a centering spring element in addition to providing sealing at the damper ends. The elastomer rings are not reliable as spring elements and have a very narrow range of stiffness. They degrade with time and temperature. Centering the damper with the O-rings is also difficult because they tend to creep due to the static loading. O-rings are also not capable of taking any thrust load which is required in certain applications.
Another problem experienced in most conventional damper bearings is cavitation and air ingestion caused by negative pressure in the squeeze film cavity. Such cavitation is a primary cause of poor performance of conventional damper bearings.
Fluid film bearings, on the other hand, have significant damping capability from the fluid film. Of the available fluid film bearings, the so-called tilt-pad radial bearing is by far the most universally-prescribed design for machines requiring maximum rotordynamic stability because of its exceptional stability characteristics. Consequently, it has become the standard by which many other radial bearings are measured when seeking a highly stable bearing design. The tilt-pad bearing's popularity is evidenced by the large number of applications found in industry, both as original equipment, and as aftermarket replacements. Applications range from small high-speed machines such as turbochargers and compressors, to very large equipment such as steam turbines and generators.
The high rotordynamic stability comes from the reduction of cross-coupled stiffness that occurs when pads are free to tilt about their individual pivot points. This attenuates the destabilizing tangential oil film forces that can induce catastrophic subsynchronous vibration in machines equipped with conventional fixed-geometry bearings. Since so many machines are susceptible to this type of bearing-induced instability, there is a large demand for quality tilt-pad bearings.
Because of its many moving parts and manufacturing tolerances, the tilt-pad design is also the most complex and difficult to manufacture of all journal bearing designs. The design complexity is evident in the number of highly-machined parts required to make up the bearing. Clearance tolerances are additive in the built-up assembly of shell, pivots, and pads, requiring a high degree of manufacturing accuracy to yield acceptable radial shaft clearances. Pad pivot friction under high radial load can also lead to premature wear, or even fatigue failure, which can enlarge clearances and increase rotordynamic unbalance response. All of these requirements combine to make the tilt-pad bearing one which demands maximum attention to design, manufacturing, and materials.
Many of today's modern turbomachines, especially those running at high speeds and low bearing loads, require the superior stability characteristics of tilt-pad journal bearings to prevent rotordynamic instabilities. Until now, the design complexity of tilt-pad bearings has precluded their use in many small, high-volume applications where cost and size are important.
The present inventor has developed an improved, less complicated moving pad bearing construction. For example, U.S. Pat. No. 4,496,251 a pad which deflects with web-like ligaments so that a wedge shaped film of lubricant is formed between the relatively moving parts.
U.S. Pat. No. 4,515,486 discloses hydrodynamic thrust and journal bearings comprising a number of bearing pads, each having a face member and a support member that are separated and bonded together by an elastomeric material.
U.S. Pat. No. 4,526,482 discloses hydrodynamic bearings which are primarily intended for process lubricated applications, i.e., the bearing is designed to work in the working fluid. The hydrodynamic bearings are formed with a central section of the load carrying surface that is more compliant than the remainder of the bearings such that they will deflect under load and form a pressure pocket of fluid to carry high loads.
It has also been noted, in Ide U.S. Pat. No. 4,676,668, that bearing pads may be spaced from the support member by at least one leg which provides flexibility in three directions. To provide flexibility in the plane of motion, the legs are angled inward to form a conical shape with the apex of the cone or point of intersection in front of the pad surface. Each leg has a section modulus that is relatively small in the direction of desired motion to permit compensation for misalignments.
U.S. Pat. No. 5,054,938 also to Ide discloses a number of bearings particularly well-suited for high speed equipment. The bearings include fluid dampened support structures.
Such deflection pad bearings offer exceptional damping characteristics. In addition to the damping typically associated with tilt pad bearings, the support structure and fluid located between the webs also provide damping. It is even possible to provide an oil filled diameter membrane to increase damping. Moreover, because these bearings function without contact between moving parts they offer the possibility of virtually infinite life.
Despite the advantages offered by these bearing constructions, they have not yet been universally accepted. This can be attributed, at least in part, to the revolutionary nature of these bearings and the fact that they are a radical departure from "conventional" thinking in the field of rotordynamics. In addition, when a fluid film bearing fails it often can completely seize without warning. The results could be catastrophic in a jet engine, for example. On the other hand, failure of a rolling element bearing is usually gradual and indicated by the increasing noise generated by the bearing. Moreover, rolling element bearings work, to some extent, even without lubricant. This certainly accounts for the continued use of rolling element bearings in jet engines, but does not explain the continued use of rolling element bearings in applications where loss of lubricant is less catastrophic, e.g., aeroderivative applications. For whatever reason, there remains a preference among some in the field for rolling element bearings. There is a need, therefore, for a simple inexpensive reliable system which provides good damping characteristics for rolling element bearings.
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