Patent Abstract:
The present disclosure concerns the application of the “Wave Bearing Concept” to journal and thrust fluid film bearings to increase performance and reliability. The wave surface is present on whichever member is stationary or non-rotating. Some applications are: pressurized gas journal wave bearings for increased load capacity and dynamic stability; journal wave bearings with liquid lubricants for extreme load capacity and excellent thermal and dynamic stability under any load; thrust wave bearings for axial positioning and axial loads; journal bearings with an elastic wave sleeve that can be activated via actuators (“active/passive control fluid film bearing”) or may change by itself (“smart bearings”) to adapt the bearing performance to the applied bearing load and speed. Journal and thrust bearings incorporating the present invention are appropriate for either mono-directional or bi-directional rotation.

Full Description:
REFERENCES  
       [0001]     1. Dimofte, F., “Wave Journal Bearing with Compressible Lubricant; Part I: The Wave Bearing Concept and a Comparison to the Plain Circular Bearing,” STLE Tribology Trans. Vol. 38, 1, pp.153-160, (1995).  
         [0002]     U.S. Patent Documents:  
                                           5,593,230   Jan. 14, 1997   Tempest, Michael, C., and Dimofte, Florin       6,024,493   Feb. 15, 2000   Tempest, Michael, C., and Dimofte, Florin       6,428,211   Aug. 06, 2002   Murabe, et al.       6,402,385   Jun. 11, 2002   Hayakawa, et al.                  
 
         [0003]     Statement of Federal Sponsored Research/Development:  
         [0004]     Federal founds were use in certain testing of the wave bearings.  
       BACKGROUND OF THE INVENTION  
       [0005]     1. Field of the Invention  
         [0006]     The present invention concerns journal and thrust fluid film bearings which include a wave surface to optimize load capacity, thermal stability, and dynamic behavior for varying operating conditions.  
         [0007]     2. Description of Related Art  
         [0008]     High speed, high performance machines need stable, low friction bearings in order to operate smoothly and efficiently. Current standard journal bearings suffer from instabilities that can severely hinder operation of such machinery.  
         [0009]     The electronics industry has provided numerous new developments for high speed bearings, used, for example, in hard disc drives, laser printers, and other electronic equipment where speeds in excess of 10,000 rpm are needed. These bearings typically use a gas, specifically air, as a lubricant.  
         [0010]     Tempest and Dimofte in U.S. Pat. No. 5,593,230 disclose an air bearing having a non-circular form, which when developed into a normally flat plane has a shallow sinusoidal contour having three peaks, “wave peaks.” Each peak is arranged 120° to an adjacent peak. The top peak is formed with a groove which enhances dynamic stability of the bearing.  
         [0011]     Tempest and Dimofte in U.S. Pat. No. 6,024,493 disclose an air bearing which includes a static shaft wherein the shaft has a sinusoidal wave form, and a rotary polygon mirror device incorporating the air bearing.  
         [0012]     Murabe and Komura in U.S. Pat. No. 6,428,211 disclose a hydrodynamic gas bearing structure comprising a shaft with notches, “space enlarging portions,” located about the circumference of the shaft at equal distances. These notches are used to supply fluid to the bearing.  
         [0013]     Hayakawa, et al., in U. S. Pat. No. 6,402,385 disclose a dynamic pressure bearing that includes a rotary shaft and a centered oil-retaining bearing with pockets in the internal surface of the bearing to increase the pressure of the lubricating oil between the shaft and the oil-retaining bearing, for use in high rotational precision equipment, such as magnetic disc drives, polygon mirror rotary drives (laser printers), and the like.  
         [0014]     Such bearings as described in the prior art have not been shown to perform in applications where high temperatures in addition to high speed may be encountered. In particular, gas turbine engine manufacturers are seeking engine main shaft bearings capable of operating up to temperatures of 700° F. and 4 million DN (where DN is the speed parameter, the product of bearing bore diameter in mm and shaft rotative speed in rpm). Such operating conditions are beyond the capability of conventional ball and roller bearings. Under even less severe conditions, ball and roller bearings become unreliable, with reduced life cycle, increased maintenance problems and costs, and increased safety concerns.  
         [0015]     Conventional circular journal bearings are disadvantaged in high performance applications due to tendencies to promote shaft instabilities at high speeds and low load conditions. More recently, non-circular types of journal bearings which provide more stability have been developed; some are disclosed, for example, in U.S. Pat. Nos. 5,593,230; 6,024,493; and 6,428,211.  
         [0016]     Gas lubricated journal wave bearings without any supply of lubricant are disclosed and have been described, in Dimofte, F., “Wave Journal Bearing with Compressible Lubricant-Part I: The Wave Bearing Concept and a Comparison to the Plain Circular Bearing,” STLE Tribology Transactions, Vol. 38(1), pp. 153-160 (1995).  
         [0017]     The journal wave bearing is a journal bearing which features a non-circular or wave configuration on the bearing sleeve. (Ref. 1) There is a slight, but precise variation in the circular profile such that a wave profile is circumscribed on the diameter of the stationary part, having an amplitude equal to a fraction of the bearing clearance. The rotating member has a circular configuration.  FIG. 1  shows a journal wave bearing having three waves in the bearing sleeve, and a circular rotating journal or shaft. The “radial clearance” is the difference between the sleeve and shaft radii. The sleeve radius is the radius of the mean circle of the wave ( FIG. 1 ). The shaft can rotate in either direction. The waves have a starting point ( FIG. 1 ) which is the maximum outside point of the wave profile closest to the load position, and can be located by the wave position angle. In  FIG. 1  the wave height and clearance are greatly exaggerated. Typically, the wave height and the clearance are about one thousandth the size of the radius.  
         [0018]     The journal wave bearing has several unique advantages when compared to either the plain journal bearing or other types of non-circular journal bearings such as a lobed, fixed pad, or tilting pad. The plain journal bearing has the highest load capacity, but shafts supported in it are subject to instabilities known as fractional frequency, whirl which can lead to failures. The occurrence of fractional frequency whirl makes journal plain bearings unsuitable for lightly loaded, high speed applications. Non-circular types of journal bearings can provide stable shaft operation and their use is obligatory in applications where “shaft whirl” is a problem. The journal wave bearing has two advantages over other known types of non-circular journal bearings: it has the highest load capacity of all the types of non-circular journal bearings, and it is the least expensive bearing to fabricate.  
         [0019]     Journal wave bearing technology has been demonstrated with compressible fluid (gas) lubrication. With gas lubrication, the bearing is typically surrounded by the gas so that supplying the bearing with lubricant is not a problem; it does not require any sophisticated design features. The surrounding gas at the bearing edges is absorbed into the bearing where the distance between the shaft and the sleeve is large and it is exhausted where the shaft and sleeve surfaces are very close to each other.  
         [0020]     There remained a need: to combine the wave shape advantages to raise the performance of the pressurized gas journal bearings; to extend the performance of the liquid lubricated journal bearings beyond their current limits by including the wave shape; to develop new, simple, and efficient thrust bearings that use the wave shape; and to open another avenue for developing active control and smart bearings based on wave bearing technology. All these create methods of operating high performance rotating machinery at higher speeds, higher temperatures, and higher efficiency, with extremely precise rotation and reliable performance. The present invention meets this need.  
       SUMMARY  
       [0021]     The object of this invention is to provide bearings having a wave surface on the stationary bearing part while the rotating member has a plain configuration. In particular the present invention provides a pressurized gas journal bearing having a wave surface that adds an improved hydrodynamic effect when the shaft rotates, in conjunction with the pressure supplied externally. The shaft can rotate in both directions. The bearing load capacity, stiffness, and stability can be significantly improved as compared to either a pressurized plain bearing or an aerodynamic wave bearing. The present invention also provides a liquid lubricated journal wave bearing having a wave surface circumscribed on the diameter of the stationary part. The position of the waves and the lubricant supply ports position is optimized for the specific application. Any liquid, such as, for example, cryogenics, mineral and synthetic hydrocarbon oils, fuels, water, polyphenylethers (PPE), and perfluoropolyethers (PFPE), can be used. The bearing can run at any temperature at which the lubricant remains stable. Another object of the present invention is to provide a bidirectional double thrust wave bearing consisting of an axial disk located between a pair of thrust plates. In addition, the present invention provides a mono-directional singular thrust wave bearing consisting of an axial disk that faces a thrust plate. Either the disk or the thrust plate rotates. The stationary part of this bearing (either the thrust plate or the disk) has a wave surface incorporated into its active face. The interaction of the stationary wave surface and the plain running surface generates hydrodynamic pressures that allow the bearing to carry thrust loads. These thrust wave bearings can be lubricated with any gas or liquid and can run at any temperature (assuming lubricant stability). Finally, this invention provides wave bearings with an elastic stationary part. The elastic part has a wave surface that can be distorted to adapt the bearing performance to the applied loads and speeds. The distortions are made by actuators (as an “Active/Passive Control Fluid Film Bearing”) or by the hydrodynamic pressures between the stationary and rotating parts (as a “Smart Bearing”). 
     
    
     BRIEF DESCRIPTION OF THE FIGURES  
       [0022]      FIG. 1  shows the journal wave bearing concept. Wave height and clearance are greatly exaggerated.  
         [0023]      FIG. 2  shows a pressurized, gas lubricated wave bearing according to the present invention.  
         [0024]      FIG. 2A  shows a 3D view of the pressurized, gas lubricated, wave bearing sleeve according to the present invention.  
         [0025]      FIG. 3  shows a liquid lubricated journal wave bearing according to the present invention.  
         [0026]      FIG. 3A  shows a 3D view of the liquid lubricated journal wave bearing sleeve according to the present invention.  
         [0027]      FIG. 3B  shows a pressure distribution in the fluid film of the liquid lubricated journal wave bearing according to the present invention.  
         [0028]      FIG. 3C  shows the profile of a transmission gear which acts as the bearing sleeve, distorted by the applied forces, and a stationary wave shaft, according to the present invention.  
         [0029]      FIG. 4  shows a double thrust wave bearing according to the present invention.  
         [0030]      FIG. 4A  shows a 3D view of a thrust plate according to the present invention.  
         [0031]      FIG. 4B  shows a 3D view of a thrust plate with holes for pressurized gas according to the present invention.  
         [0032]      FIG. 4C  shows a 3D view of a thrust plate of a liquid lubricated thrust bearing according to the present invention.  
         [0033]      FIG. 5  shows a journal wave bearing with an elastic sleeve that is distorted by actuators according to the present invention.  
         [0034]      FIG. 5A  shows a one wave elastic element according to the present invention.  
         [0035]      FIG. 6  shows a bidirectional smart journal bearing with an elastic wave surface according to the present invention.  
         [0036]      FIG. 6A  shows an unloaded smart journal wave bearing according to the present invention.  
         [0037]      FIG. 6B  shows a smart journal wave bearing under half the maximum load, according to the present invention.  
         [0038]      FIG. 6C  shows a smart journal wave bearing under the maximum load, according to the present invention. 
     
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT  
       [0039]     A pressurized gas journal wave bearing  10  according to the present invention is illustrated in  FIG. 2 . The journal bearing  10  supports a rotating shaft  50 . A vertical load  90  is applied to the shaft  50 .  
         [0040]     The bearing sleeve  15  has a wave surface  18  circumscribed on its inner diameter. If the shaft is stationary and the sleeve is rotating, the wave profile is circumscribed on the shaft diameter (not illustrated). The profile of the wave surface  18  shows a “mean circle”  19 . The radius  20  of the mean circle  19  is also the radius of the bearing sleeve. The wave surface has a starting point  22 . The wave has an amplitude  25  which is the distance from the mean circle  19  to the maximum outside point of the wave  26 . The position of the wave relative to the applied load direction  90  is defined by the wave position angle  30 . The wave surface has a plurality of waves (three are illustrated here). The wave surface  18  is made either through a manufacturing process (such as grinding, lapping, honing, pressing, etc) or through elastic deformation of the sleeve  15  when it is mounted in its housing.  
         [0041]     The bearing is supplied with gas (air) through holes  35  which can be designed with either inherent or orifice restrictors. In  FIG. 2A , a 3D illustration of the bearing sleeve  15  is shown. Any number of supply holes  35  can be used (24 are illustrated). The holes  35  can be located in several supply planes (only two are illustrated in  FIG. 2A ).  
         [0042]     The shaft has a radius  55  and an axis of rotation  57 . Without a load, the axis of rotation  57  will be in the center of the bearing sleeve  11 . When a load  90  is applied, the shaft axis  57  moves in an offset position. The distance  12  between the center of the sleeve  11  and the axis of the shaft  57  is the “eccentricity.” The difference between the bearing sleeve radius  20  and the shaft radius  55  is the bearing radial clearance. The ratio of the wave amplitude  25  to the radial clearance is the “wave amplitude ratio.” 
         [0043]     In most machinery, loads are built up as the shaft is rotating. At rest the load applied to the bearings is the weight of the rotating part only. Therefore, the gas (air) supplied through the holes  35  is enough to levitate a “non rotating” shaft  50 . When the shaft starts rotating the pressure around the shaft is amplified by the hydrodynamic effect of the plurality of convergent regions of the fluid film thickness between the shaft surface  58  and the wave surface  18 . According to the present invention, in  FIG. 3 , the fluid film between the shaft surface  58  and the wave surface  18  shows minimum thickness in several locations  40  (three here). Convergent regions of the fluid film are developed upstream of these locations  40  when the shaft rotates either clockwise or counterclockwise  51 . These convergent regions help create hydrodynamic pressures, that in conjunction with the supplied pressurized gas increases the bearing load capacity beyond the limits of the load capacity of the hydrodynamic plain and wave bearing, or the pressurized plain bearing. The waves also improve the bearing stability. The bearing dynamic stiffness and damping can be adjusted to the values required in conjunction with the dynamic behavior of the rotor that is to be supported, by varying the wave amplitude  25 . Thus, the rotor&#39;s critical speeds can be avoided and greater dynamic amplitude suppressed when the rotor runs at specific rotation speeds.  
         [0044]     The sleeve  18  and the shaft  50  are made from: solid ceramic materials such as silicon nitride or silicon carbide; solid hard alloys with superficial coatings (such as physical vapor deposition, PVD, or diamond like carbon, DLC, coatings); or metallic materials with plasma spray ceramic coatings.  
         [0045]     The pressurized wave bearing can be used (for example) in any high precision machinery, such as high precision tools, centrifuges, and inspection machines, as well as in small or medium sized turbo-machinery, compressors, fans, air-breathing machines, and auxiliary power units.  
         [0046]     A journal wave bearing lubricated with liquids  10  according to the present invention, is illustrated in  FIG. 3 . The journal bearing  10  supports a rotating shaft  50 . A vertical load  90  is applied to the shaft  50 .  
         [0047]     The bearing sleeve  15  has a wave  18  circumscribed on its inner surface. If the shaft is stationary and the sleeve rotates the wave surface is circumscribed instead on the shaft (not illustrated). The profile of the wave surface  18  shows a mean circle  19 . The radius  20  of the mean circle  19  is also the radius of the bearing sleeve. The wave surface has a starting point  22 . The wave has an amplitude  25  which is the distance from the mean circle  19  to the maximum outside point of the wave profile  26 . The position of the wave surface relative to the applied load direction  90  is defined by the wave position angle  30 . The value for this position angle  30  is optimize for the specific application and can be in a range from 0 to 60 degrees. The wave surface has a plurality of waves (three, for example, are illustrated). The wave surface  18  is produced either through a manufacturing process (such as grinding, lapping, honing, pressing, etc) or through elastic deformation of the sleeve  15  when it is mounted in its housing.  
         [0048]     The bearing is supplied with a liquid lubricant through a plurality of holes  135  (only three are illustrated), one for each wave. These holes  135  feed the supply pockets with lubricant  136 , as seen in  FIG. 3A . In  FIG. 3A , a 3D illustration of the bearing sleeve  15  is shown. The locations of the holes  135  and the pockets  136  relating to the wave profile  18  are defined by the “supply location angle”  140  between the supply hole axis  137  and the starting point of the waves  22 . If this angle is zero (not illustrated in  FIG. 3 ) the shaft can rotate in either a clockwise or a counterclockwise direction and the bearing is appropriate for bi-directional journal rotation. According to the present invention, the location of the holes  135  and the pockets  136  defined by the angle  140  can be optimized to maximize bearing load capacity while running at the lowest temperature. A frequent value is 20 degrees but can have various values for a specific application. In this case the journal bearing is appropriate for mono-directional rotation.  
         [0049]     The shaft has a radius  55  and an axis of rotation  57 . Without a load the axis of rotation  57  will be in the center of the bearing sleeve  11 . When a load  90  is applied, the shaft axis  57  moves to an offset position. The distance  12  between the center of the sleeve  11  and the axis of the shaft  57  is the eccentricity. The difference between the bearing sleeve radius  20  and the shaft radius  55  is the bearing radial clearance. The ratio of the wave amplitude  25  to the radial clearance is the wave amplitude ratio.  
         [0050]     When the shaft starts rotating, hills of pressure are created between the shaft  50  and the sleeve  15  due to the hydrodynamic effect of the plurality of convergent regions of the fluid film thickness between the shaft surface  58  and the wave profile  18 . According to the present invention, in  FIG. 3 , the fluid film between the shaft surface  58  and the wave surface  18  shows minimum thicknesses in several locations  40  (three are illustrated). Convergent regions of the fluid film are upstream of all locations  40  when the shaft rotates either clockwise or counterclockwise  51 . These convergent regions help create hydrodynamic pressures in any position of the shaft  50  inside the bearing sleeve  15 . Thus, if the shaft  50  is unloaded and the eccentricity  12  is zero, the axis of the shaft  57  takes a concentric position in the center of the sleeve  11 , and hills of pressure are still present—unlike the case of a plain journal bearing which cannot create any hydrodynamic pressure when it is unloaded. According to the present invention, the permanent presence of the hills of pressure inside the wave bearing as soon as the shaft rotates stabilizes the bearing at all loads. The wave position angle  30  can be selected so that the applied load to the bearing is supported by two hills of pressure.  FIG. 3B  shows the pressure distribution in an unwrapped bearing. The position of the load is in between two hills of pressure. A supply hole and pocket are inserted in between the pressure hills and fresh lubricant at supply temperature is injected into the bearing just before the next hill of high pressure. This configuration allows the bearing to run thermally stable at any load and temperature avoiding the situation of when the lubricant viscosity could collapses and bearing fails.  
         [0051]     According to the present invention, wave journal bearings are appropriate for use when the rotating bearing part, either the bearing sleeve or the shaft, deforms under the applied load. A bearing with a rotating elastic sleeve is illustrated in  FIG. 3C . Pressure distribution with multiple hills due to the wave profile (two are illustrated in  FIG. 3B ) with lubricant supply ports between the pressure hills supports deformation of the bearing sleeve. An example of such a case is a wave bearing used to support planetary gears in transmissions. In this case the shaft is stationary and the bearing sleeve, the actual planetary gear is rotating. Due to the gear loads the gear sleeve deforms and its shape varies from that of a rigid gear, as illustrated in  FIG. 3C . The wave profile is circumscribed on the stationary shaft&#39;s outer diameter. The location of the waves are properly selected and the pressure hills, such as illustrated in  FIG. 3B , support both radial Fr and tangential Ft loads shown in  FIG. 3C .  FIG. 3C  also shows that an elastic gear sleeve supported by a waved shaft can handle heavy loads better than a rigid gear sleeve. The minimum lubricant film thickness of the elastic gear sleeve that occurs at position  1  is greater than the minimum film thickness of the rigid gear-sleeve that occurs at position  2 . Thin film thicknesses such as at position  2  cause the bearing to fail.  
         [0052]     To preserve the wave bearing performance, the bearing geometry must be unchanged during the wave bearing&#39;s life. The shaft and the sleeve is made from hard materials, with a hardness greater than 60 HRc. Any steels and alloys that can be hardened or case-hardened greater than 60 HRc may be used.  
         [0053]     Coatings (such as physical vapor deposition, PVD, or diamond like carbon, DLC, coatings) are applied to both shaft and sleeve surfaces to avoid damage to the wave bearing surfaces when the bearing starts and stops, and to make the wave bearing less sensitive to lubricant interruption.  
         [0054]     The wave bearing, according to the present invention, can be used in heavily loaded applications with specific loads up to 24 MPa (3500 PSI). The wave bearing is also very appropriate for use in any medium-sized loaded application with specific loads up to 5.5 MPa (800 PSI) where stable motion is requested at all loads. Journal wave bearings, according to the present invention, are appropriate for either mono-directional or bi-directional journal rotation. The wave bearings have stiffness and damping properties that can be adjusted to the needs of the machinery in which they are being used. In particular, their damping characteristics are useful to attenuate the noise and vibration level of any machinery and particularly in mechanical aero and terrestrial transmissions. Their thermal stability makes the wave bearings very suitable for high temperature application. When lubricated with polyphenylethers (PPE) and perfluoropolyethers (PFPE) the wave bearing runs at temperatures over 350° C. (662° F.).  
         [0055]     A bidirectional thrust wave bearing  200  lubricated with a fluid (gas or liquid) according to the present invention is illustrated in  FIG. 4 . A rotating shaft  201  having a disk  202  is supported in the axial direction  203  by two stationary thrust plates  204  separated by a spacer  205 . The thrust plates provide a bidirectional axial effect in the region  206  that positions the shaft in the axial direction  203  or carries loads in both axial directions  207  and  208 . If the axial load is permanent in only one direction and no axial positioning is required, only one thrust plate  204  is used for a mono-directional thrust wave bearing (not illustrated). The shaft rotates around its axis  203  either in clockwise or counterclockwise directions  209 .  
         [0056]     A thrust plate  204  is illustrated in  FIG. 4A . According to the present invention, both gases and liquids can be used as lubricants. The thrust plate  204  has an inner radius  230  and an outer radius  235 . The active face of the thrust plate  204  has a wave surface  240  with a “middle plane”  250 . The middle plane  250  is tilted from the horizontal plane  255  with a tilt angle  257 . The tilt angle is positive (as illustrated), or can be negative or zero. The wave surface  240  has a plurality of waves (four are illustrated). Each wave has an amplitude  245  which is constant along the radial direction as illustrated in  FIG. 4A , or variable along the radial direction (not illustrated). The active wave surface  240  of the thrust plate  204  faces the disc&#39;s active surface  210 . If the shaft is stationary and the thrust plate(s)  204  rotate, the wave surface is made on the disc&#39;s active surface  210 . The wave surface  240  or  210  is produced either through a manufacturing process (such as grinding, lapping, honing, pressing, etc) or through elastic deformation of the thrust plate  204  when it is mounted in his housing.  
         [0057]     According to the present invention, a gas thrust wave bearing could be also supplied with pressurized gas through holes with restrictors as illustrated in  FIG. 4B . The holes  250  are located at a radius  255  greater than inner radius  230  and less than outer radius  235 . The pressurized gas provides a smooth start. In addition, according to the present invention, when the shaft rotates the pressurized gas is supplied through holes  250  into the clearance between the active surface  210  of the disk  202  and the active surface  240  of the thrust plate  204 ; in conjunction with this, the hydrodynamic effect of the wave surface  240  increases the bearing performance beyond the limits of either the pressurized thrust bearing with plain surfaces or a non-pressurized thrust wave bearing.  
         [0058]     When a liquid lubricant is used, according to the present invention, the thrust plates  204  could have radial grooves  260  at the start of each wave, as illustrated if  FIG. 4C . These radial grooves allow the lubricant to easily enter between the active surface  210  of the disk  202  and the active surface  240  of the thrust plate  204 . The liquid lubricant can also supply the thrust bearing through holes and pockets similar to the holes  135  and pockets  136  illustrated in  FIG. 3A . These holes and pockets are located at the start of each wave, replacing the grooves  260 . The wave surface  240  has the middle plane  250  horizontal with a zero tilt angle and the wave amplitude  245  is constant along the radius. The wave amplitude  245  can vary along the radius (not illustrated in  FIG. 4C ). Positive or negative tilt angle  257  can be also used but not illustrated in  FIG. 4C .  
         [0059]     According to the present invention, both the disk  206  and the thrust plates  204  are made from hard materials. For gas lubricated thrust bearings the disk and the thrust plate are made from: solid ceramic materials such silicon nitride or silicon carbide; solid hard alloys with a superficial coating (such as physical vapor deposition, PVD, or diamond like carbon, DLC coatings) on the active faces  210  and  240 ; or hard stainless steels with plasma spray ceramic coatings on the active faces  210  and  240 . For liquid lubricated thrust bearings, steels and alloys that can be hardened or case-hardened over 60 HRc can be used. Coatings (such as physical vapor deposition, PVD, or diamond like carbon, DLC, coatings) are applied on the active faces  210  and  240  to avoid damage to the bearing surfaces when the bearing starts and stops and to make the bearing less sensitive to lubricant interruption.  
         [0060]     A controllable journal wave bearing  300 , according to the present invention, is illustrated in  FIG. 5 . The controllable journal wave bearing  300  supports a rotating shaft  50 . The shaft  50  can rotate clockwise or counterclockwise  51 . The bearing housing  310  includes an elastic shell  315  that has a wave surface  18  with a mean radius  20  and amplitude  25 . The wave surface has a plurality of waves (six are illustrated). A portion of the elastic shell  315  that corresponds to one wave is illustrated in  FIG. 5A . This portion has a length  330  (called L) and a width  335  (called B). The ratio of B/L should be close to 1/2. The mean radius of the waves  20  is called R m . The number of waves is approximated as 2πR m /L, but is not less than three. Large diameter bearings with a length to diameter ratio of less than 1/2 need more than 3 waves. The elastic shell  315  is made as one piece or from a number of pieces, one for each wave. They are assembled together at the locations of wave ends.  
         [0061]     According to the present invention, the amplitude  25  of the wave is controlled by the actuators  320 . Any type of actuator can be used, for example, mechanical, electromagnetic, piezoelectric, hydraulic, or pneumatic. The actuators are connected to an active or passive control system that adjusts the wave amplitude  25  to shaft speed, shaft vibration level, and load. Enlarging the wave amplitude  25  causes the bearing to run stably and increases the bearing stiffness. Under heavy loads the bearing is stable and the wave amplitude should be diminished to approach the plain journal bearing geometry; the bearing can then carry a heavy load better than any type of fluid film bearing.  
         [0062]     The bearing  300  is lubricated with a liquid lubricant. Both oils and fuels are can be used. The lubricant is supplied to the bearing through holes  135  and pockets  136  shown in  FIG. 5  and  FIG. 5A . The holes and pockets are located at the beginning of each wave. According to the present invention, this location of the pressure holes and pockets permits a supply of fresh lubricant near the hot spots of the fluid film which keeps the bearing running thermally stable, especially at high speeds or heavy loads.  
         [0063]     Both the shaft  50  and the elastic shell  315  are made from hard materials, with hardness over 60 HRc. Any steels and alloys that can be hardened or case-hardened over 60 HRc can be used. Coatings (such as physical vapor deposition, PVD, or diamond like carbon, DLC, coatings) are applied to both shaft and elastic shell surfaces to avoid damage to the controllable bearing surfaces when the bearing starts and stops, and to make the controllable bearing less sensitive to lubricant interruption.  
         [0064]     According to the present invention, the controllable bearing  300  can be used in high performance rotating machinery which needs high precision rotation, or safe rotation with levels of vibration under fixed limits. Rotating machinery which is heavily loaded but starts and stops under low loads will benefit from the use of the controllable wave bearing  300 .  
         [0065]     According to the present invention, a self-acting (smart) wave bearing  400  is illustrated in  FIG. 6 . The smart wave bearing  400  supports a rotating shaft  50 . The shaft  50  rotates clockwise or counterclockwise  51 . The smart wave bearing has an elastic shell  410 . The elastic shell has initial shape as a wave surface with a mean circle  19 . The wave surface has a plurality of waves (three are illustrated). The elastic shell  410  is supported by the bearing housing  420 . If the bearing is lubricated with a liquid, holes  135  and pockets  136  are located at the beginning of each wave. The shell is free to deform under the pressure in the fluid film and to change the position of its inside sections  430  that are closer to the shaft surface  450  than the mean circle  19 .  FIGS. 6A  to  6 C show haw the smart bearing works. If the shaft  50  rotates and is unloaded ( FIG. 6A ), its axis  57  is concentric to the shell center  11 . The shell wave surface is uniform around the circumference and has equal amplitudes  25  in all locations. According to the present invention, the shaft is running stably due to the wave shape of the shell when it is unloaded.  
         [0066]     If a vertical load is applied to the shaft, the pressure in the fluid film opposite the load increases and distorts the shape of the elastic shell in that region.  FIG. 6B  illustrates a case when a vertical load  90 ′, equal to one half of the maximum load that the bearing can carry, is applied to the shaft  50 . According to the present invention, when the load  90 ′ is applied, the axis  57  of the shaft moves into an eccentric position relative to the center  11  of the elastic shell; the pressure increases in the bottom side of the shaft, and the elastic shell  410  diminishes its amplitude  25 ′ at the bottom of the bearing (compared to the initial amplitude  25  of the wave surface). This makes the bearing better able to carry the applied load  90 ′, while still running stably, due to the wave shape of the elastic sleeve  410  ( FIG. 6B ) which still shows a three wave shape.  
         [0067]     According to the present invention, if the vertical load increases to the maximum load  90 ″ that the smart bearing  400  can carry, the amplitude  25 ″ of the bottom wave goes to zero, approaching a shape similar to that of a plain bearing on the bottom side, as illustrated in FIG.  6 C. The elastic shell  410  superimposed over the mean circle in the bottom side of the smart bearing allows the bearing to carry a higher maximum load than a rigid wave bearing.  
         [0068]     According to the present invention, any fluid (gas or liquid) can be used to lubricate the smart bearing. The smart bearing runs very stably dynamically and thermally at any speeds and loads and can carry a maximum load greater than any fluid film bearing including a plain journal bearing. The mart bearing can approach a shape similar to the plain bearing in the region that carries the load as the load increases (see  FIGS. 6A  to  6 C), but it is better lubricated than the plain bearing, running more thermally stable than the plain bearing.  
         [0069]     Both the elastic shell and the shaft are from a hard metallic alloy. Coatings (such as physical vapor deposition, PVD, or diamond like carbon, DLC, coatings) are applied to both shaft and elastic shell surfaces to avoid damage to the controllable bearing surfaces when the bearing starts and stops, and to make the smart bearing less sensitive to lubricant interruption.

Technology Classification (CPC): 5