Staged bearing surface compliance for hydrodynamic fluid bearing

A hydrodynamic fluid bearing includes a pivoted shoe which can tilt in response to the hydrodynamic fluid forces generated between the bearing surfaces, and a locally conformable bearing layer attached to the shoe facing the opposing bearing surface. The pivoted shoe can pivot to conform to changes in the pitch of the opposing bearing surface and can tilt to the correct slope to generate the hydrodynamic supporting fluid wedge, and can also conform to the shape of the opposing bearing surface if it becomes thermally distorted or "dished". The compliant layer at the top surface of the shoe can conform to smaller area bearing irregularities, and in addition, can locally conform under hydrodynamic forces to assume the ideal configuration to generate the optimum hydrodynamic wedge. The support on which the pivoted shoe is mounted can itself be pivoted on a gimbal ring to provide large area compliance with the bearing surface to follow, if necessary, precession or other large scale runout of the bearing surface. Each stage of the compliant support can be designed to give the stiffness and damping best suited for that stage of support.

BACKGROUND OF THE INVENTION 
The field of gas bearings has long been considered theoretically ideal for 
large ultrahigh speed machinery. However, it has been disappointing in 
practice because the load carrying capacity of bearings of this nature, 
which theoretically should be great, has been found in practice to be 
disappointingly small. I have ascertained some of the causes for this 
disappointing performance and avoided them with a gas hydrodynamic bearing 
capable of supporting heavier loads at high speed. 
Hydrodynamic fluid bearings are essentially of two varieties: (a) hard 
surface types, and (b) compliant surface types. The hard surface types 
include (1) plane surface; (2) contoured surface types, such as spiral 
groove, tapered-land, step, pocket; and (3) tilt pad. Hydrodynamic 
supporting fluid films are generated over hard surface types by dragging 
the lubricating fluid, whether it be oil, water or air, by viscous shear 
forces, into a converging space between the two bearing members. The shape 
of that space is determined, in the case of plane and contoured surface 
types, by the profile to which the bearing surfaces are machined, and also 
by the position and orientation of the bearing surface of the rotating 
member. The tilt pad type employs a pad mounted on a pivot which enables 
the tilt pad to tilt under the influence of hydrodynamic forces acting on 
it to assume a slope that will cause a supporting hydrodynamic fluid film 
to be generated. 
The compliant surface type of bearing employs a compliant support layer 
which supports a flexible bearing sheet such as a thin sheet of stainless 
steel. Under the influence of relative movement between the opposing 
bearing surface and the bearing sheet, hydrodynamic forces are generated 
which depress the compliant support and the overlying bearing sheet to a 
profile that is conducive for generating a supporting hydrodynamic fluid 
film between the opposing bearing surface and the bearing sheet. 
The hydrodynamic supporting fluid film is created by the viscous or shear 
forces acting in the fluid parallel to the direction of relative movement 
between the two bearing surfaces. A rotating thrust runner, for example, 
drags its boundary layer of air with it as it rotates opposite to a tilt 
pad. The boundary layer, in turn, drags in the immediately adjacent layer 
of air, and so forth. In this way, an air velocity gradient is established 
in the gap between the thrust runner and the tilt pad. The pad is 
supported at about 58% of its length in the direction of rotation of the 
thrust runner and tilts away from the thrust runner at its leading edge. 
This creates a wedge-shaped gap between the thrust runner and the pad 
which causes the fluid being dragged into the wedge to increase in 
pressure toward the trailing edge of the tilt pad. This pressure increases 
gradually to a maximum at approximately 3/4 of the pad length from the 
leading edge. 
In the same way, compliant surface bearings also create a hydrodynamic 
supporting fluid wedge by viscous drag of the boundary layer exerted by 
the relative movement between the rotor bearing surface and the compliant 
bearing surface. A zone of high pressure fluid is created which provides 
the fluid support to maintain the separation between the opposing bearing 
surfaces. 
A pressure curve showing the distribution of pressure over a hard, tilt pad 
bearing surface reveals that the pressure increases gradually from the 
leading edge of the wedge shaped gap between the opposing bearing surfaces 
to a maximum at about 3/4 of the length, then falls off steeply. This 
pressure curve inherently results in a peak. Although the maximum pressure 
in the supporting fluid film might be high, the total supporting force is 
not high because of the nonuniform pressure distribution over the surface 
of the bearing. The pressure curve over a compliant bearing module reaches 
an early maximum pressure which is maintained over a substantial portion 
of the module surface area before falling off steeply at the end of the 
module. The total supporting force over the compliant bearing module is 
higher than that over the rigid tilt pad because the area under the 
compliant bearings's pressure curve is considerably greater than the area 
under the tilt pad's pressure curve, even though the maximum pressure over 
the compliant bearing may be less. 
One of the advantages of gas bearings is their ability to operate in high 
temperature environments. However, it is common for a thrust bearing to 
operate in a very uneven temperature condition, with the thrust runner 
being hotter at its bearing face than on the opposite face, establishing a 
temperature gradient across the bearing member in an axial direction. The 
higher temperature zone at the bearing face causes the material of the 
bearing to expand nonuniformly and produce a convex shape of the bearing 
face. This thermal distortion of the bearing surface has an adverse effect 
in the hydrodynamic action at the bearing interface which no longer has 
the optimum wedge for generating the hydrodynamic supporting fluid film. 
In effect, the distortion reduces the bearing surface and transfers the 
entire load to the portion of the bearing surface which has not been bowed 
away from the opposed bearing surface by thermal distortion. The smaller 
effective bearing surface now must carry the same load, and the resulting 
greater pressure may exceed the load bearing capacity of the bearing. This 
condition may be further exacerbated by misaligned loads on the thrust 
bearing. Where the shaft tilts or precesses under the misaligned load, one 
edge of the thrust runner lifts away from the thrust plate and the entire 
axial load must then be borne by the opposite edge. 
The damping effect of hydrodynamic fluid bearings is theoretically ideal 
for ultrahigh speed applications. However, the runout excursions which 
should be damped by the bearing, may be of a nature which do not coincide 
with the bearing's damping characteristics. Thrust runner excursions 
exerted on a thrust bearing can vary widely in frequency amplitude and 
direction. In addition, fluid effects, such as half speed whirl, can exist 
in the lubricating fluid itself to complicate the situation. Failure of a 
bearing can occur when the power of the rotor runout excursions exceeds 
the damping capacity and misalignment tolerance of the bearing, such as 
when a rotor, passing through its critical velocity, experiences runout of 
such amplitude that the hydrodynamic fluid film is breached and the 
bearing surfaces contact each other at high relative speed. 
Many prior art hydrodynamic bearings have been unsuccessful because of the 
lack of understanding of what occurs in the bearing interface and how the 
complex forces acting on the bearing interact and affect the hydrodynamic 
action of the bearing. When solutions are proposed, they are usually 
focused on a single, perceived problem in the bearing and fail to account 
for the other effects. I have designed a bearing which accounts for the 
factors which are most influential in the operation of the hydrodynamic 
fluid bearing and which solve the problems which most seriously affect its 
load bearing capacity. 
SUMMARY OF THE INVENTION 
Accordingly, it is an object of this invention to provide a hydrodynamic 
fluid bearing having a high load carrying capacity. This bearing can be 
used in high speed heavy machinery applications with long and reliable 
operation. It is tolerant of misalignment, vibration, and thermal 
distortion of the thrust runner, and it can function in extremely hostile 
environments such as high temperature, low temperature, steam and 
corrosive conditions which would quickly destroy conventional bearings. It 
is quiet running and its damping characteristics can be matched to 
correspond to the damping requirements of the application. 
Although this bearing is inherently cool running and efficient because it 
absorbs little energy from the rotor which it supports, its design 
facilitates easy supplemental cooling of the bearing members if needed. 
Although this bearing gives superior performance, it is actually more 
economical to manufacture than bearings far inferior in performance. It 
does not require high precision machining or exotic bearing materials and 
indeed it is amenable to high volume production techniques. In addition to 
ease and speed of construction, the inspection, replacement, and repair of 
this bearing is easy and quick and does not require precision alignment or 
extensive disassembly of the machine in which it is being used. 
These objectives are achieved by one embodiment of this invention which 
employs a series of tilt pads, each of which includes a compliant bearing 
layer. The pads themselves will self-adjust and conform to the slope of 
the opposing bearing surface, and the compliant bearing surface on the pad 
will self-adjust to the smaller scale local bearing surface and 
automatically assume the ideal profile to produce a uniform, high capacity 
hydrodynamic supporting fluid film. 
Another embodiment contemplates mounting the thrust plate, on which the 
tilt pads are mounted, on a gimbal to enable the plate itself to follow 
precession and other large scale deflections of the thrust runner, and to 
flex axially in response to axial shocks exerted by the rotor on the 
bearing. 
The stiffness of each support stage in the two embodiments is designed to 
properly contribute to the desired overall stiffness of the bearing 
combination. Thus, for example, the stiffness of the large scale support 
stage such as the gimbal arrangement for the thrust plate must be selected 
to complement the stiffness of the smaller stages, otherwise the large 
scale support stage will act like a rigid support, or will make the 
smaller scale stages act like rigid supports. Thus, this bearing provides 
the maximum hydrodynamic supporting fluid film generating capability with 
minimum heat generation and energy losses, while permitting damping 
characteristics for the application involved.

DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring now to the drawings, wherein like reference characters designate 
identical or corresponding parts, and more particularly to FIG. 1 thereof, 
a bearing assembly having a series of tilt pad assemblies 10 mounted on a 
thurst plate 12, is shown facing a thrust runner 14 mounted on a rotating 
rotor 16. 
Looking now at FIGS. 2 and 3, each tilt pad assembly 10 includes a tilt pad 
body 15 supported on a flexure support 18 mounted on the thrust plate 12. 
The flexure support 18 includes a single mounting stalk having lower 
cylindrical base portion 20 and an upper reduced portion 22 attached to 
the tilt pad body 15 by a weld flange 24 formed at the top of the reduced 
portion 22. The lower cylindrical portion 20 fits into a cylindrical bore 
26 formed in the thrust plate 12 and is welded at 28 to the thrust plate 
12 to secure it permanently in position. To insure the proper spacing 30 
and angular orientation between the tilt pad 10 and the thrust plate 12, 
and to insure that the several tilt pads in the thrust bearing assembly 
are exactly coplanar, an alignment jig (not shown) may be inserted between 
the underside of the tilt pad 10 and the top surface 32 of the thrust 
plate prior to welding the cylindrical portion 20 to the underside of the 
thrust plate 12 at 28. The alignment jigs are then removed after the 
welding has been completed. 
The particular form of flexure support illustrated is employed where 
conventional lubrication is impracticable, for example, because of high 
temperature. In situations where lubrication of a ball and socket support 
for the tilt pad 10 is possible, it may be substituted for the flexure 
support 18. 
The bearing surface of the tilt pad assembly 10 is provided by a flat 
flexible sheet 34 of bearing material such as stainless steel of Iconal, 
which may also be coated with an antifriction coating such as molybdenum 
disulfide or, in high temperature applications, a proprietary coating 
known by the trademark "HL-800" sold by Mechanical Technology 
Incorporated, of Latham, New York. "HL-800" includes a mixture of cadmium 
dioxide and graphite in a thin flexible ceramic matrix. Its composition 
and method of application are disclosed in the co-pending application of 
Bharat Bhusan Ser. No. 974,264 entitled "High-Temperature Low-Friction 
Surface Coating" filed concurrently herewith. The flexible sheet 34 of 
bearing material is attached to the leading edge (in the sense of the 
rotation direction of the thrust runner 14) by spot welding, seam welding 
or the like at 36 to a spacer block 38. The sheet 34 of bearing material 
is resiliently supported by a resilient supporting structure 40, shown 
here as a sheet of Inconal X-750 formed in a corrugated pattern to produce 
a series of regular wave forms which can yield locally or over a wide area 
to give resilient and conformable support for the sheet 34 of bearing 
material. 
The spacer block 38 may be welded to the tilt pad body 15 over the leading 
edge of the corrugated sheet to secure the bearing layer assembly together 
and in place on the tilt pad. 
The underside of the sheet 34 may be treated with friction enhancing 
material or may be etched, sand blasted or otherwise treated or machined 
to provide a friction enhancing texture to the surface. Likewise both 
surfaces of the resilient supporting structure 40 and the top surface of 
the tilt pad body 15 may be similarly treated to enhance the frictional 
characteristics of the surfaces. The purpose of this treatment is to 
provide optimal frictional or coulomb damping between these surfaces. 
Vibrations transmitted to the sheet 34 of bearing material from the thrust 
runner 14 will cause the resilient support member 40 to flatten slightly, 
which has the effect of spreading the area of the support member 40. The 
peaks and valleys of the corrugations thus travel laterally in small 
displacements with respect to the underside of the sheet 34 of bearing 
material and the topside of the tilt pad body 15. This slight rubbing of 
surfaces over a relatively wide area provides an ideal mechanism for 
damping high frequency vibrations. 
The angular stiffness of the flexure support of the tilt pad 10 is selected 
to complement the stiffness of the support element 40. It is desirable 
that the angular stiffness of the flexure 22 be about one-half to 
one-tenth of the angular stiffness of the support element 40. The lower 
limit of twice the angular compliance of the element 40 is based on the 
need to "save" the element 40 for small scale deflections. Otherwise the 
deflections necessary to produce the gross or large scale deflections of 
the tilt pad would be partially contributed by the element 40, so the full 
range of deflection of the element 40 would not be available for the small 
scale adjustments in bearing sheet height that the element 40 is ideally 
suited to provide. The greater angular compliance of the flexure 22 
enables the large scale deflections of the pad in following the runout 
excursions of the rotor to be carried out by the flexure instead of the 
support element 40 whose range of motion is much smaller than that of the 
flexure 22. The upper limit of the flexure angular compliance is actually 
imposed by its minimum vertical stength. It must be strong enough to 
support the axial force exerted on the pad 10 by the thrust runner 14. 
In addition to the two stages of bearing surface compliance and damping 
shown in FIGS. 1-5, it may be desirable to add one or more additional 
stages of damping and/or compliance in some applications. A third stage 
providing large scale surface compliance is provided by gimbaling the 
thrust plate 12 in gimbal ring 50. The gimbal ring 50 includes a ring 52 
which is supported for rotation about an axis defined by the axis of 
pivots 54 supported by the machine on which this bearing is used. The ring 
52 carries a second set of pivots 56 mounted in the same plane as the ring 
52 and orthogonally disposed to the axis of the pivots 54 so that the 
thrust plate 12, which is carried by the pivots 56, can rotate about the 
axis of the pivot 56 and can also rotate, by virtue of the pivotal support 
of the ring 52, about the axis of pivots 54 at right angles to the axis of 
the pivot 56. Therefore, the thrust plate 12 can tilt in any direction to 
follow the tilting or precession of the thrust runner 14. In addition, the 
pivots 54 and 56, and the ring 52 can be manufactured with a stiffness 
designed to give the optimum resilience in an axial direction to enable 
the bearing to flex axially in the event of axial shocks to the rotor. The 
rotor is thus protected from permanent damage which could occur in the 
event of a rigid mounting. 
The pivots 54 and 56 can be provided with a damping mechanism such as 
coulomb damping. In addition, damping may be provided to the thrust plate 
12 by other means, such as an elastomer pad or other energy absorbing 
material positioned on the back side of the thrust plate 12 between it and 
its support. 
If the maximum vibration excursions which the thrust plate 12 will 
experience are small enough that the gimbal ring mounting shown is 
unnecessary, the thrust plate 12 may be mounted directly on a resilient 
support such as an elastomer block, which would permit the thrust plate to 
rock in response to the large scale vibration which it experiences. The 
material used will depend on the desired damping characteristics and axial 
load which the thrust plate 12 will carry. In addition, environmental 
factors such as temperature, chemicals present and life expectancy will 
influence which material is to be used. Obviously, numerous modifications 
and variations of the disclosed embodiments will occur to those skilled in 
the art.