JP5276414B2 - Followable hybrid gas journal bearings using an integral wire mesh damper - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a gas bearing having high load capability and damping properties. <P>SOLUTION: A damper bridge (36) of the follow-up hybrid gas journal bearing (10) has an axial length (38) smaller than an axial length (40) of each of a bearing pad (12) and an outside rim (34) to form a damper cavity (42) on each side of the damper bridge (36). An integrated wire mesh pattern (62) is located in the damper cavity (42) on each side of the damper bridge (36). Integrated centering springs (46, 48) are arranged between an inside rim (32) and the outside rim (34) for giving radial and rotational follow-up to the bearing pad (12). This oil-free bearing design copes with low damping and load capability properties specific in current follow-up aerofoil bearing design while holding follow-up to a change of a rotor shape. <P>COPYRIGHT: (C)2010,JPO&amp;INPIT

Description

  The present invention relates generally to bearings, and more particularly to a compliant hybrid gas bearing that uses an integral wire mesh damper.

  High speed equipment such as aircraft engine turbomachines, and aircraft diversion applications such as steam turbines, gas turbines, compressors, etc. must pass several natural or critical speeds before reaching the design operating speed. When the system operates at natural frequency or critical speed, the vibration amplitude of the system / rotor increases. These vibrations resulting from rotor imbalance can be destructive or even catastrophic if not properly damped. Bearings with appropriate damping characteristics limit or dampen synchronous rotational frequency vibrations so that equipment can safely pass critical speeds. Further, good bearing damping contributes to rotor dynamic stability at high speeds by suppressing subharmonic excitation of the rotor natural frequency.

  Fluid or oil film journal bearings have long been used to damp vibrations created by turbomachines. Aircraft gas turbine engines and industrial centrifugal compressor rotors often use oil-based squeeze film damper bearings supported by rotating cage-type retrocenter springs to control vibration amplitude. In a fluid film bearing, a buffer part is formed between a journal surface on which a thin fluid film rotates and a stationary bearing surface to attenuate vibrations from the rotor. In a squeeze film damper bearing, a thin film of fluid, usually in the form of oil, is squeezed by two non-rotating cylindrical surfaces. One surface is stationary, while the other is positioned by a retrograde spring structure and a track associated with the rotation of the rotor. The squeezing of the fluid film as a result of the rotor orbital motion attenuates the rotor vibration via the bearing support.

  The simplest squeeze film damper bearing design does not include a retrograde spring. The outer race of the rotating element bearing, or the outer bearing shell in the case of a fluid film bearing, can float and press in the clearance space between the bearing outer diameter and the housing inner diameter. The absence of a mechanical decentering spring in this design means that the damper journal will have reached the bottom upon activation. As the speed increases and the shaft begins to rotate, the journal of the damper (bearing shell outer surface) will lift off. The oil film in the squeeze film damper does not produce direct rigidity as in conventional fluid film bearings. However, this damper produces a direct stiffness-like behavior. This direct stiffness is due to a cross-coupled damping coefficient that exhibits (spring) properties such as stiffness.

  This unintentional damper is one of the most nonlinear of squeeze film damper designs. There are two basic mechanisms that cause this non-linear behavior. The first of these two non-linear mechanisms can be attributed to the non-linear characteristics caused by the cross-coupled damping coefficient. A second source of non-linear behavior present in this type of damper arises as a direct result of the damper journal reaching the bottom.

  The simplest means of providing a retrocentric spring in the squeeze membrane damper is through the use of an elastomeric O-ring. The advantages of this design are its simplicity, ease of manufacture, and the ability to incorporate a damper in a small enclosure. Some of the disadvantages associated with this design are attributed to the limited range of stiffness that can be achieved with elastomers. Due to material differences and the effects of temperature and time on their properties, it is difficult to predict stiffness with a good degree of certainty for elastomeric materials. This O-ring design is also susceptible to creep, allowing the damper to reach the bottom, which can lead to a bi-linear spring behavior as discussed above.

  The most commonly used squeeze membrane damper design, especially in aircraft engines, is a rotating cage support damper. The salient feature required for such a design is the relatively large axial space required compared to the damper length. This is one of the major drawbacks of this damper design. The rotating cage that forms the damper spring for the damper very often requires three to four times the axial space of the damper itself.

  Assembling the rotating squirrel-cage spring and restoring the journal into the clearance space requires special tools and skills. This rotating cage spring also complicates the damper end seal design and assembly. It is also very difficult to offset the spring assembly to accommodate the gravitational load due to the shaft weight. Maintaining parallelism between the damper journal and the housing is another factor that adds uncertainty and complexity to this design.

  Another oil based squeeze membrane damper design includes an integral damper retrocenter spring. In this design, the cantilevered support ribs form a retrograde spring element with the sectors that they support at both ends. A small gap between the sector and the outer ring forms a squeeze film damper clearance space. Unlike the squirrel cage spring design, this integral damper retrocenter spring design does not occupy any additional axial space beyond the existing length occupied by the bearing. The completed assembly can include any number of sectors depending on the load and the required stiffness and damping for a particular application. Wire electrical discharge machine (EDM) provides excellent means for obtaining the desired clearance with very high accuracy and repeatability that maintains excellent parallelism between the damper journal and the housing.

  Despite the advantages provided by these squeeze membrane bearing structures, oil-lubricated bearings pose much higher cost and maintenance burdens and reliability problems with oil leakage, filtration and conduits. These and other shortcomings of oil-lubricated bearings have led the industry to pursue the development of compliant airfoil bearings.

US Pat. No. 5,603,574 US Pat. No. 5,743,654 US Pat. No. 6,379,046 US Pat. No. 6,630,761 US Pat. No. 6,695,478

Today, however, airfoil technology has been largely limited to small lightweight rotors and machines, ie, aircraft air cycle machines (ACMs). Thus, while the benefits of incorporating air bearings into an aircraft are well understood, significant technical challenges must be overcome in order to develop an actual design. These challenges include developing gas bearings with significantly higher load capacity and damping than today's airfoil bearing technology.

  Briefly, a compliant hybrid gas journal bearing is radially and concentric between a plurality of compliant hybrid bearing pads and an inner rim, outer rim, and inner and outer rims adjacent to the plurality of bearing pads. A damper bridge to be inserted having a plurality of bearing pads and an axial length shorter than the axial length of the outer rim, thereby forming a damper cavity on each side of the damper bridge; And an integral wire mesh damper located in the damper cavity on each side of the.

  In another aspect of the invention, the compliant hybrid gas journal bearing includes at least one bearing pad including a hydrostatic recess and a capillary restrictor for supplying a flow of pressurized gas. A plurality of compliant hybrid bearing pads, an inner rim adjacent to the plurality of bearing pads, an outer rim, and a damper bridge inserted radially and concentrically between the inner and outer rims, wherein the plurality of bearing pads And a damper bridge having an axial length shorter than the axial length of the outer rim, thereby forming a damper cavity on each side of the damper bridge.

  In yet another aspect of the present invention, the compliant hybrid gas journal bearing has a plurality of compliants, wherein at least one bearing pad includes a hydrostatic recess and a capillary restrictor for supplying a flow of pressurized gas. Hybrid bearing pad and an inner rim, an outer rim adjacent to the plurality of bearing pads, and a damper bridge inserted radially and concentrically between the inner rim and the outer rim, the plurality of bearing pads and the outer rim A damper bridge having an axial length shorter than the axial length of the damper bridge thereby forming a damper cavity on each side of the damper bridge, and an integral wire mesh damper located in the damper cavity on each side of the damper bridge And a plurality of integral decentering springs disposed between the inner rim and the outer rim.

  These and other features, aspects, and advantages of the present invention will be better understood when the following detailed description is read with reference to the accompanying drawings, in which like characters indicate like parts throughout the several views. Will.

1 is a partial cross-sectional perspective view of a compliant hybrid gas journal bearing assembly according to an embodiment of the present invention. FIG. FIG. 2 is an end view of the followable hybrid gas journal bearing housing of FIG. 1. FIG. 3 is a cross-sectional view of the compliant hybrid gas journal bearing housing taken along line 3-3 of FIG. FIG. 3 is an enlarged cross-sectional view of a bearing housing span central fan portion of the followable hybrid gas journal bearing housing of FIG. 2. FIG. 5 is an enlarged cross-sectional view of a hydrostatic pressure recess and a capillary restrictor in the central section of the bearing housing span of FIG. 4.

  Referring to the drawings in which the same reference numerals indicate the same elements throughout the various figures, FIGS. 1 to 5 show a compliant hybrid gas journal bearing shown generally at 10 according to one embodiment of the present invention. The bearing 10 includes a plurality of followable hybrid bearing pads 12. Each bearing pad 12 includes a pad surface 14 proximate to a rotating rotor or shaft 16 that normally rotates in the direction of arrow 18. In one embodiment, the bearing pad 12 comprises a tilted pad with an integral spring that is offset (FIG. 2). Each bearing pad 12 also includes an axial slot 20 for minimizing the weight of the bearing pad 12. The pad surface 14 of each bearing pad 12 also includes a hydrostatic recess 22 having a depth 24 and a width 26 (FIG. 5). For example, the hydrostatic recess 22 can have a depth of 2 to 10 times the gas film thickness between the bearing surface 14 and the rotor surface. In the illustrated embodiment, the ratio of the hydrostatic recess 22 area to the bearing surface 14 area is about 0.027. In general, the smaller the hydrostatic recess 22 area compared to the bearing surface area, the greater the room for bearing stability.

  In one embodiment, the hydrostatic recess 22 is centrally disposed on each bearing pad 12 and is therefore symmetrically disposed about the shaft 16. In another embodiment, the hydrostatic recess 22 is off-center on the bearing pad 12. However, it is still possible to arrange the hydrostatic recess 22 symmetrically around the shaft 16 even if the recess 22 is off-center on the bearing pad 12. Each hydrostatic recess 22 is in fluid communication with a capillary restrictor 28 for supplying a flow of pressurized gas, such as air, to the bearing pad 12. Each capillary restrictor 28 is connected to a pressurized gas source 30 such as a compressor bypass flow (FIG. 1). In one embodiment, the capillary restrictor 28 is disposed asymmetrically with respect to the hydrostatic recess 22. In another embodiment, the capillary restrictor 28 is symmetrically disposed with respect to the hydrostatic recess 22. This pressurized gas provides a hydrostatic lifting force so that the shaft 16 does not reach the bottom during activation. In other words, the pressurized gas supplies a lifting force so that the shaft 16 can be lifted away from the bearing pad 12 even when the shaft 16 is not rotating. In one embodiment, the gas has a pressure in the range of between about 206.8 kPa (30 psig) to about 1379 kPa (200 psig).

  In another embodiment, the bearing 10 does not have a hydrostatic recess 22 but has only a capillary restrictor 28 relative to the surface 14 of the bearing pad 12. The capillary restrictor 28 in this configuration can be centered around the axes 54, 58, or the arrangement can be asymmetric with respect to the axes 54, 58 and offset toward the leading edge of the bearing pad 12. Can do.

  In the illustrated embodiment, the bearing 10 has four bearing pads 12 that are symmetrically disposed about the shaft 16, each bearing pad 12 having a single hydrostatic recess having a capillary restrictor 28. 22 is included. However, it will be appreciated that the invention is not limited by the number of bearing pads, recesses and capillary restrictors, and that the invention can be practiced with any desired number of bearing pads, recesses and restrictors.

  Referring to FIGS. 3 and 4, the bearing 10 includes an inner rim 32 adjacent to the bearing pad 12, an outer rim 34, and a damper bridge 36 that is inserted radially and concentrically between the inner rim 32 and the outer rim 34. . The gap 35 divides the inner rim 32 into a plurality of sector portions 50 corresponding to the plurality of bearing pads 12 as a whole. The damper bridge 36 has an axial length 38 that is shorter than the axial length 40 of the bearing pad 12 and outer rim 34, thereby forming a damper cavity 42 having a width 44 on each side of the damper bridge 36.

  One aspect of the present invention is that the bearing 10 includes a plurality of integral decentering springs 46, 48 disposed between the inner rim 32 and the outer rim 34. In one embodiment, each sector 50 of the bearing 10 includes a pair of integral decentering springs 46, 48, where the spring 46 is a leading edge integral spring and the spring 48 is a trailing edge integral spring. In the illustrated embodiment, the integral retrograde springs 46, 48 have a generally “S” -shaped cross-sectional shape. The integrated retrocentric springs 46, 48 can be formed using wire electrical discharge machine (EDM) technology. The gap 60 between the bearing pad 12 and the outer rim 34 provides a clearance space for the springs 46, 48 to adapt to the effects of centrifugal force and heat generated for high speed, high temperature applications. Similarly, this clearance space 60 reduces the bearing pad weight and the overall weight of the bearing assembly. The integral retrocentric springs 46, 48 provide a linear behavior, unlike the non-linear behavior exhibited by conventional bump foil bearings. This integral retrocentric spring design differs from conventional rotating cage designs that require an additional axial length in any additional axial space beyond the axial length 40 of the bearing pad 12. Not even occupy. An example of an oil-based, integrated wire squeeze membrane damper is available from KMC, Inc. of West Greenwich, RI. (Www.kmcbearing.com) is commercially available.

  As shown in FIG. 4, the pair of springs 46, 48 for each sector 50 has been shown to greatly enhance bearing stability and load capacity to produce tilted pad movement or pad rotation. The bearing pad 12 is offset. In other words, the springs 46 and 48 are disposed asymmetrically with respect to the bearing pad 12. Specifically, the center line of the spring 46 forms an angle 52 with respect to the horizontal axis 54 of the sector 50, and the center line of the spring 48 is about a vertical axis 58 that is substantially perpendicular to the horizontal axis 54. Form a smaller angle 56 with respect to it. Having an angle 56 that is less than angle 52 creates an off-center moment on the bearing pad 12, thereby creating a wide range of tilting or rotation of the bearing pad 12. In one embodiment, the difference between the angles 52, 56 can range between about 10 degrees and about 25 degrees. For example, angle 52 can be about 28 degrees, and angle 56 can be about 18 degrees, or a difference of about 10 degrees. Moreover, the difference between the angles 52, 56 allows the bearing pad 12 of the bearing 10 to be radially oriented so as to allow radial elongation of the bearing bore resulting from centrifugal forces and thermal effects generated for high speed, high temperature applications. And allows it to be rotationally followable. In the illustrated embodiment, the bearing 10 has four sectors 50 with a pair of springs 46, 48 in each sector. However, it is understood that the present invention can be practiced with any desired number of sectors and any desired number of springs within each sector, depending on the load for the particular application and the required stiffness. Will.

  As explained above, the asymmetrical arrangement of the integral springs 46, 48 and the asymmetrical arrangement of the hydrostatic recess 22 cause extensive tilting and rotation of the bearing pad 12. However, extensive tilting and rotation of the bearing pad 12 can also be achieved by providing the leading edge integral spring 46 with a relatively lower radial stiffness than the trailing edge integral spring 48. For example, the leading edge integral spring 46 can have a radial stiffness of about 75,000 pounds / inch, while the trailing edge integral spring 48 can have a radial stiffness of about 85,000 pounds / inch. . This difference in radial stiffness between the springs 46, 48 causes extensive tilting and rotation of the bearing pad 12.

  Another aspect of the present invention is that the bearing 10 includes an oil-free integrated wire mesh damper (IWMD) 62 positioned within the damper cavity 42 on each side of the damper bridge 36. In other words, the IWMD 62 is located between the bearing pad 12 and the outer rim 34 as shown in FIG. This IWMD 62 is a knitted wire mesh comprising metal wires or plastic wires knitted into a net structure. This knitting process produces a mesh of interlocking loops. These loops can move relative to each other in the same plane without deforming the net, giving the knitted net an extension in two directions. When exposed to tensile or compressive stress, each loop behaves as a small spring so that the knitted metal has inherent elasticity. Woven metal also provides high mechanical oil-free damping properties and non-linear spring constants. Vibrations and mechanical shocks can be effectively controlled to eliminate severe resonance conditions and provide sufficient protection from dynamic overload. Studies have shown that IWMD 62 provides at least 30 times more damping than conventional airfoil bearings. The IWMD 62 can be made from a variety of materials such as steel, inconel, aluminum, copper, tantalum, platinum, polypropylene, nylon, polyethylene, and the like. The density and dimensions of the IWMD 62 can be adjusted to suit specific design applications. An example of a wire mesh damper is commercially available from Metex Corporation (www.metexcorp.com) of Edison, NJ.

  There are significant differences between the compliant hybrid gas journal bearing 10 of the present invention and a conventional bearing assembly. As noted above, conventional oil-based integrated wire squeeze membrane dampers are available from West Greenwich, RI's KMC, Inc. Commercially available. This KMC integrated wire squeeze membrane damper (ISFD) using an “S” shaped spring is a bearing support that results from the damping moving fluid from one control space to another. This movement of fluid through the clearance space or orifice creates a “fluid-based” viscous dissipation that is realized as damping in the bearing support system. Typically, this squeeze film fluid is the same as the lubricating fluid of a rotating element bearing or journal bearing in which ISFD is incorporated. In conventional ISFD bending pivot bearing assemblies, symmetrically positioned “S” shaped springs allow radial movement of each inner rim quadrant with bending pivot bearing pads. This bending pivot allows the bearing pad to pivot or rotate around the base of the bending pivot beam. Thus, this “S” shaped spring provides radial followability, while rotational followability or bias in the bearing pad is achieved via this bending pivot. Each mechanism, bending pivot and “S” shaped spring has a single function.

  In contrast, a compliant hybrid journal bearing 10 using an integral wire mesh damper 62 does not generate damping via squeezing or displacing fluid, but damping is not a mechanical structure, ie a wire mesh. It is generated via the damper 62. The lubricating fluid in the bearing 10 of the present invention is a gas and does not provide any viscous dissipation or damping in the bearing support, which results from hysteretic structural damping combined with Coulomb friction. This structural damping results from thousands of individual wire segments that bend together, and Coulomb friction results from minor slipping of the wires on each other. This combination provides excellent oil free source damping, especially for machines that require or can benefit from oil free operation. One advantageous feature of a compliant hybrid journal bearing 10 that uses an integral wire mesh damper 62 is that radial and rotational responsive functions are achieved only through "S" shaped springs 46,48. is there. This is because the “S” shaped springs 46, 48 of the present invention are not symmetrically arranged around each quadrant or sector 50 of the inner rim 32. This asymmetric positioning of the “S” shaped springs 46, 48 provides a rotational bias within the individual bearing pads 12 that allows rotation in addition to radial movement. In addition to having one mechanism that performs a dual function, another advantage is that the radial space occupied is reduced because there is no bending pivot, which reduces the weight of this bearing. Let This is unsurpassed in space applications because weight is a major driving factor for performance. Further bearing weight reduction results from EDM machining of the material between the “S” shaped springs 46, 48.

  This bearing 10 forms damper compression lands 68, 70, respectively, that can be used to prevent unwanted axial movement of the IWMD 62 and also to tune the stiffness and damping factor depending on the level of end plate compression. A pair of end plates 64, 66 are also included.

  The compliant hybrid gas journal bearing 10 of the present invention can be used in several modes of operation. One mode of operation is a purely hydrodynamic mode of operation. In this simplest mode of operation, this compliant gas bearing using an integral wire mesh damper has a compliant pad with no hydrostatic capability. This mode of operation may be applicable to situations where excess pressurized gas was not feasible. This mode of operation creates pressure via fluid dynamics between the compliant pad and the rotor. In this mode of operation, the bearing pad will lift off the rotor surface at the corresponding rotor speed for a given load. However, there is a transient rubbing zone before the fluid dynamics take effect and the bearing pad lifts.

  Another mode of operation consists of a transition period from hydrodynamic operation to hybrid (hydrodynamic and hydrostatic) operation. This mode may exist, for example, when external fluid hydrostatic pressure is not available during start-up and this initial mode of operation is governed by fluid dynamics. In this situation, there is a transient rubbing area where there is sliding friction between the rotor surface and the bearing pad surface. Eventually, the bearing pad lifts away due to fluid dynamics. The hydrostatics for each bearing pad will begin at a certain number of revolutions and will most likely be powered via the pressure gas released from the main working fluid in the turbomachinery. In this respect, the bearing is operating in a hybrid mode.

  Another mode of operation is the transition from hydrostatic mode to hybrid mode. This situation may exist, for example, when pressure gas is available throughout the entire operational sequence of the machine. Having pressure gas before rotor rotation is advantageous because this gas can prevent rubs during transition, reduce starting torque, and allow safe shutdown in emergency situations such as engine stalls. . In this mode of operation, the bearing pad will raise the rotor without rotation, which is pure hydrostatic. As the rotor increases in speed, the contribution from fluid dynamics becomes more prevalent. Hybrid operation is achieved at high speeds.

  As explained above, the compliant hybrid gas journal bearing 10 of the present invention provides an oil-free bearing design that addresses the low damping and low load capability characteristics inherent in today's airfoil bearing designs. Moreover, the compliant hybrid gas bearing of the present invention provides much higher damping (30 times or more) than conventional foil bearing technology, significantly greater load capacity and significantly higher manufacturing repeatability. Furthermore, the oil-free gas bearing of the present invention provides significant cost savings when compared to fluid film or oil-based damper designs.

  This written description uses examples to disclose the invention, including the best mode, and also to enable any person skilled in the art to make and use the invention. The patentable scope of the invention is defined by the claims, and may include other examples that occur to those skilled in the art. Such other embodiments include only structural elements that have structural elements that do not differ from the literal language of the claims, or that they have negligible differences from the literal language of the claims. If not, it is within the scope of the claims.

10 followable hybrid gas journal bearing 12 bearing pad 14 pad surface 16 rotor or shaft 18 arrow 20 axial slot 22 hydrostatic recess 24 depth 26 width 28 capillary restrictor 32 inner rim 34 outer rim 35 gap 36 damper bridge 38 Axial length (damper bridge)
40 Axial length (bearing pad)
42 Damper cavity 44 Width (damper cavity)
46 Integral Decentering Spring 48 Integral Decentering Spring 50 Fan 52 Angle 54 Horizontal Axis 56 Angle 58 Vertical Axis 60 Clearance 62 Integrated Wire Mesh Damper (IWMD)
64 End plate 66 End plate 68 Damper compression land 70 Damper compression land

Claims (10)

A plurality of followable hybrid bearing pads (12);
An inner rim (32) adjacent to the plurality of bearing pads (12), an outer rim (34), and a damper bridge inserted radially and concentrically between the inner rim (32) and the outer rim (34) (36) having an axial length (38) that is shorter than an axial length (40) of the plurality of bearing pads (12) and the outer rim (34), whereby the damper bridge (36 ) A damper bridge (36) forming a damper cavity (42) on each side of
A compliant hybrid gas journal bearing (10) comprising an integral wire mesh damper (62) located in the damper cavity (42) on each side of the damper bridge (36). The bearing of any preceding claim, further comprising a plurality of integral retrocenter springs (46, 48) disposed between the inner rim (32) and the outer rim (34). The bearing of claim 2, wherein the plurality of integral retrograde springs (46, 48) have an “S” shaped cross-sectional shape. The bearing of claim 2, wherein the gap (35) partitions the inner rim (32) into a plurality of sectors (50) that generally correspond to the plurality of bearing pads (12). A bearing according to claim 4, wherein each sector (50) comprises a pair of integral retrograde springs (46, 48). One (46) of the pair of integral decentering springs forms a first angle (52) with respect to the horizontal axis (54) of the bearing (10) and the pair of integral springs. The other of the decentering springs (48) forms a second angle (56) with respect to the vertical axis (58) of the bearing (10), and the first angle (52) is the second angle. 6. A bearing according to claim 5, wherein the bearing is different from the angle (56). One of the pair of integral decentering springs (46, 48) has a first radial stiffness and the other of the pair of integral decentering springs (46, 48) is The bearing of claim 5, wherein the bearing has a second radial stiffness that is different from the first radial stiffness. The bearing of claim 1, wherein each bearing pad (12) comprises a pad surface (14) with a hydrostatic recess (22) having a depth (24) and a width (26). The bearing of claim 8, further comprising a capillary restrictor (28) in fluid communication with the hydrostatic recess (22) for supplying a flow of pressurized gas to the bearing (10). The bearing according to claim 9, wherein the capillary restrictor (28) is arranged asymmetrically with respect to the hydrostatic recess (22).

Images