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WO1998035166A1 - Paliers hydrostatiques auto-compensateurs - Google Patents

Paliers hydrostatiques auto-compensateurs Download PDF

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Publication number
WO1998035166A1
WO1998035166A1 PCT/IB1997/000081 IB9700081W WO9835166A1 WO 1998035166 A1 WO1998035166 A1 WO 1998035166A1 IB 9700081 W IB9700081 W IB 9700081W WO 9835166 A1 WO9835166 A1 WO 9835166A1
Authority
WO
WIPO (PCT)
Prior art keywords
grooves
bearing
shaft
pockets
collector
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/IB1997/000081
Other languages
English (en)
Inventor
Alexander H. Slocum
Kevin Wasson
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
AESOP Inc
Original Assignee
AESOP Inc
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by AESOP Inc filed Critical AESOP Inc
Priority to PCT/IB1997/000081 priority Critical patent/WO1998035166A1/fr
Priority to AU16143/97A priority patent/AU1614397A/en
Publication of WO1998035166A1 publication Critical patent/WO1998035166A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C32/00—Bearings not otherwise provided for
    • F16C32/06—Bearings not otherwise provided for with moving member supported by a fluid cushion formed, at least to a large extent, otherwise than by movement of the shaft, e.g. hydrostatic air-cushion bearings
    • F16C32/0629—Bearings not otherwise provided for with moving member supported by a fluid cushion formed, at least to a large extent, otherwise than by movement of the shaft, e.g. hydrostatic air-cushion bearings supported by a liquid cushion, e.g. oil cushion
    • F16C32/064—Bearings not otherwise provided for with moving member supported by a fluid cushion formed, at least to a large extent, otherwise than by movement of the shaft, e.g. hydrostatic air-cushion bearings supported by a liquid cushion, e.g. oil cushion the liquid being supplied under pressure
    • F16C32/0651—Details of the bearing area per se
    • F16C32/0659—Details of the bearing area per se of pockets or grooves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C32/00—Bearings not otherwise provided for
    • F16C32/06—Bearings not otherwise provided for with moving member supported by a fluid cushion formed, at least to a large extent, otherwise than by movement of the shaft, e.g. hydrostatic air-cushion bearings
    • F16C32/0681—Construction or mounting aspects of hydrostatic bearings, for exclusively rotary movement, related to the direction of load
    • F16C32/0685—Construction or mounting aspects of hydrostatic bearings, for exclusively rotary movement, related to the direction of load for radial load only

Definitions

  • the present invention relates to self compensating hydrostatic bearings for the shaft of a spindle or the like, being more particularly directed to self compensating hydrostatic bearing designs that do not require external sleeve mountings to contain the hydrostatic self-compensating fluid logic.
  • the bearings are supported by a thin film of pressurized fluid - liquid or gaseous - including among other fluids water and air, and hereinafter sometimes generically referred to as "fluid” and interchangeably as “fluidstatic” or by hydrostatic bearings.
  • cluster spindles multiple spindle assembly for simultaneous drilling or milling, using tools with a fixed center distance
  • the diameter of the shaft is a dominant function of the total system stiffness.
  • the present invention therefore, allows designers to use standard design methods developed for self compensating hydrostatic circuitry, taking advantage of the latest design advancements as disclosed in the above mentioned patents and applications, where the direction of the spindle rotation is primarily in one direction, so the fluid flow into the hydrostatic bearing compensator starts at one end and flows toward the other end
  • the collector grooves and the pockets are machined into the surface of the shaft, and then holes are drilled to establish flow channels across chords of the shaft or external shaft surface groove flow diameters are formed to connect corresponding oppositely located collector grooves and pockets. Because the fluid is collected at the training edge of the collector and injected at the leading edge of the pocket, the flow channels on or near the outer regions of the shaft are minimally affected by models centrifugal forces.
  • the invention is concerned with a system that includes a round, cylindrical, or arcuate bearing housing which guides the rotary motion of a round internal coaxial shaft that maintains its distances from the concentric surrounding cylindrical bearing housing surface sections by means of a thin pressurized film of fluid emanating from pockets in either the cylindrical bearing housing surfaces that surround the shaft or in the shaft bearing surfaces and are geometrically opposed to each other at opposing regions therof.
  • the flow of fluid to the pockets is regulated to allow a differential pressure to exist between the pockets in response to a force being applied to the shaft or bearing housing.
  • the mechanism by which this fluid flow regulation is accomplished involves compensated opposed pocket bearings that behave like two resistance in series with one another and in parallel with another series set.
  • the resistance to fluid flow out of the bearing pocket of the load side increases, and decreases out of the bearing pocket of the other side.
  • the resistances to fluid flow into the bearing pockets behave in opposite relation to the resistances to fluid flow out of the bearing pockets; the resistance to fluid flow into the bearing pocket on the load side decreases, and increases into the bearing pocket on the other side. The result is that the pressure increases in the pocket on the side to which the load is applied until the load and the differential pressure generated between the two pockets balance.
  • the bearing therefore generates a restoring force and compensates for the applied load.
  • the resistance of the fluid flow into the pockets is herein referred to as compensation.
  • the fluid is routed from the inlet resistance features to the outlet resistance features, either by holes drilled through the housing or shaft to define flow channels and connect these points or grooves formed into the bearing surface section are used to provide such flow channels, thus providing a simpler, less costly, and more effective bearing.
  • a further objective is to provide self-compensated hydrostatic bearings integrated onto the spindle or other shaft.
  • Another object of the present invention is to provide a new and improved self compensating hydrostatic rotaiy motion bearing and method, void of prior art disadvantages, and that involves novel mechanisms to route the fluid from the inlet compensation resistances to the bearing pockets, whereby a more easily manufactured bearing with increased load-carrying capability is produced.
  • Still an additional object is to provide a novel bearing construction in which the fluid resistance, or compensation, into the bearing pocket is formed by a geometric pattern on the bearing housing surface itself, such that the nominal equilibrium position of the bearing, the resistance of the mechanism will be in desired proportion to the fluid resistance out of the bearing pocket regardless of the magnitude of the nominal equilibrium gap; thereby obtaining an easy-to-manufacture hydrostatic bearing that requires no special hand tuning of its performance.
  • the invention relates to self-compensating hydrostatic bearing fluid circuitry onto the shaft of a spindle or the like which enables the design to realize a self- compensating hydrostatic bearing design integrated with the spindles that does not require the system to contain the prior art self-compensating hydrostatic circuitry to be formed on the outside diameter of sleeve, requiring pressing into a bearing bore.
  • the design of the invention consists of a cylindrical bearing bore with easy to machine circumferential grooves that are each accordingly connected by fluid channel means in or along the shaft to pressure supply and drain systems.
  • the collections of grooves consist typically of two pressure supply grooves spaced a distance apart, typically 10-15 mm apart, and two drain grooves each typically 5 mm axially located from the said pressure supply grooves, and a drain groove typically one shaft diameter axially located from one of the other drain grooves.
  • the sets of grooves are arranged in two sets at each end of said bearing bore to act as fluid supply and fluid drain grooves; and a bearing rotor fits into the bearing bore with a radial clearance that allows for normal shaft deflection, while still allowing for radial bearing gap to provide hydrostatic support action between the bore and the shaft.
  • Circumferential collector groove means are machined or formed into both ends of the surface of the shaft; typically, four grooves equally spaced around the shaft with circumferential arc length on the order of 60 angular degrees, such that when the shaft is placed in the bore, the sets of grooves in the shaft are axially located between the set of two pressure supply grooves.
  • Single or multiple groove pocket shapes are circumferential ly spaced and equal in number to the number of collector grooves and axially displaced from the collector grooves, such that the pockets are located between the largely spaced drain grooves in the bore when the shaft is inserted in the bore.
  • Flow channels are provided either by holes drilled in the collector grooves through chords of the shaft traversing the same to connect to one corner of the pocket, or by surface grooves formed along the external shaft surface to traverse the same, such that when fluid flows axially from the pressure grooves across the shaft into the collector grooves, in proportion to the radial clearance between the shaft surface and the bore, it can flow to the pocket opposite to the collector, and this act to provide a restoring force in proportion to the radial displacement of the shaft.
  • the invention further embraces in a fluidstatic bearing having opposed bearing surface sections in a cylindrical shaft surrounded concentrically by a cylindrical housing in which the shaft extends coaxially therealong and therebetween, with each bearing surface section having similar and symmeti ical pockets and grooves in the surface through which pressure fluid travels to provided a thin film of fluid inteiposed in the gaps between the shaft and the housing surfaces, appaiatus for self-compensating for load variation on either side of the bearing, that comprises, a common pressure annulus from which fluid is fed over a compensation resistance region to a groove flow channel which routes the fluid longitudinally around the shaft on its surface to a pocket, and with the resistance to fluid flow to these grooves being adjusted to equal a proportion of resistance to fluid flow out of the opposite surface pocket when the bearing is at nominal equilibrium position and gap, unloaded by external forces, whereby as external forces are applied, the fluid flow is regulated to self-compensate for the load proportionately to variation in the bearing gap caused by the applied load, with a differential pressure being established
  • Fig. 1 is an end view of a spindle assembly, showing the housing and four spindles that form a cluster;
  • Fig. 2 is a cross section through the cluster assembly using flow channels along chords within the shaft connecting oppositely located collector and pocket grooves;
  • Fig. 3 is a cross section through one of the shafts;
  • Fig. 4 is a flat projection (unwrapped) of the arcuate cylindrical surface sections of a bearing constructed in accordance with the present invention, illustrating the compensation resistance regions, fluid distribution grooves, and pockets and the use of flow channel grooves on the external surface of the shaft to connect collector grooves and pockets;
  • Fig. 5 is a partial isometric view of the bearing system of Fig. 4 which, for the pote of clarity, has an exaggerated gap between the shaft and housing, and a section of the housing cut away to reveal the geometry on the surface of the shaft;
  • Fig. 6 is a flat projection (unwrapped) of the arcuate cylindrical surface sections of a modified bearing with pockets which have the same order-of-magnitude width as the fluid distiibution pockets; and Fig. 7 is a flat projection (unwrapped) of the arcuate cylindrical surface sections of a further modified bearing with drainage grooves between the pockets so as to prevent circumferential leakage flow between the pockets.
  • a precision hydrostatic bearing supported spindle is manufactured by placing a smooth rough shaft into a bearing bore that either contains pockets machined into the bore, or sleeves that contain the pockets and hydrostatic circuitry that is pressed into the bore.
  • a housing 20 is shown containing multiple spindle shafts la, lb, lc, and I d therein. These spindles may be used, for example, to bore or drill fixed spacing holes in a part. Such applications are extremely common in the automotive industry. Unfoitunately because the stiffness of a shaft is proportional to the fourth power of the shaft diameter, it is extremely important to be able to maximize the shaft diameter in a cluster spindle.
  • circumferential collector grooves 4a, 4b, 4c, and 4d are integrally machined into surfaces at both ends of the shaft la (and correspondingly into all the other shafts) as shown in Figs. 2 and 3.
  • four grooves are equally spaced around the shaft with circumferential arc length on the order of 60 angular degrees, such that when the shaft is placed in the bore of the housing 20, sets of collector grooves are axially located between sets of two pressure supply grooves 1 1 and 12. Flow is then regulated to the collector grooves according to the load and resulting gap changes between the shaft and the bore.
  • Drain groove 14 acts to drain fluid from the other end of the bearing. Radial holes 15, 16, 17, 18, and 19 connect the drain and pressure grooves to external sources, as is well known to those skilled in the art. To support the load, the pressure compensated fluid collected in the collectors 4a,
  • Fig. 2 exemplarily shows the pockets in a preferenced rhombus shape, as taught in the before described patent application.
  • Between the pockets are helical axial drain grooves 7a, 7b, 7c and 7d that maximize pocket pressure differentials and hence also the load capacity and stiffness.
  • the ends of the helical axial drain grooves have short circumferential segments 8a, 8b, 8c, and 8d which help to couple the drain groove ends to the circumferential drain grooves 13 and 14 in the bore of the housing 20.
  • the holes 6a, 6b, 6c, and 6d that connect the collectors to the opposing pockets are located at the deep ends of the collectors, such that as the fluid flow rate into the collector increases, more room is provided for it with less danger of cavitation; and at the end, it is encouraged by the inclined hole to flow into the hole and chordially across said shaft into the leading edge of the pockets.
  • a further significant advantage of this design is that the passages can easily be cleaned simply by removing the shaft.
  • the passages can only be cleaned with the use of a high pressure solvent.
  • the nature of the self compensating gap to act as a filter will prevent any large particles from causing an unremovable obstruction.
  • the drawings thus illustrate apparatus for supporting a cylindrical shaft using a pressurized thin film to provide accurate motion capability.
  • the cylindrical bearings would be spaced axially along the shaft.
  • a single rotational bearing may be combined with a thrust bearing that will resist the tilt motions.
  • the latter option will result in the formation of a veiy accurate rotaiy table.
  • the former option combined with a pure thrust bearing which may also be self-compensating, will result in the formation of an accurate spindle for use, for example, in machine tools and grinding machines and the like.
  • flow channels 6a, 6b, etc. may be provided by locating the flow channels on the external surface of the shaft to traverse the same as shown at 52A, 52B, etc. in the embodiment of Figs. 4 and 5, as 62A in Fig. 6, and as 82A, 82B in the modification of Fig. 7.
  • a cylindral shaft 40 of the invention moves coaxially inside a concentric outer cylindrical bearing house 41 (partially cut away in the figure to show more clearly the grooves and pockets in the shaft).
  • the motion of the shaft 40 about (or along) the horizontal or X axis is guided by arcuate or cylindrical (or portions of a cylinder herein termed cylindrical) longitudinally extending bearing pad surface section 53B, of which there can be multitude but at least three.
  • fluid enters the bearing at high pressure through a hole drilled radially through the housing 41 and into the supply pressure annulus 50. Since the dimensions of the annulus 50 are large compared with the other flow dimensions of the bearing, the fluid flows freely circumferentially through the annulus 50 to create a region of uniformly high pressure.
  • the fluid then flows axially across the compensation resistance regions 51A and 5 IB to the fluid routing flow channel grooves 52A and 52B, formed in the external surface of the shaft rather than traversing chordally internally through the shaft as in Figs. 2 and 3, before-described, and whose depth radially into the shaft is large (by at least a factor of about 5) compared with the clearance between the shaft 40 and the housing 41. If so dimensioned, the fluid will flow freely to the bearing pocket regions 53A and 53B whose depths are comparable to that of the grooves 52A and 52B so as to provide regions of relatively uniform pressure.
  • Leakage reduction regions 56A, 56B, 57A, and 57B are at a small distance radially from the inner surface of the bearing housing 41 (typically 0.01 mm to 0.03 mm). This region acts to substantially reduce leakage flow circumferentially between the routing grooves 52A and 52B, and circumferentially between the pockets 53A and 53B. The flow leaves the bearing pockets 53A and 53B across the pocket resistance regions 54A and 54B and into the fluid collection annuli 55A and 55B, whose dimensions are similar to those of the supply pressure annulus 50. After freely flow circumferentially through the fluid collection annuli 55A and 55B, the fluid exits the bearing to low pressure (typically atmospheric pressure) through holes 10A and 10B drilled radially through the housing 41.
  • low pressure typically atmospheric pressure
  • the flow channel routing grooves 52A and 52B are typically dimensioned to extend between 90 degrees and 180 degrees around the shaft with respect to the centers of the pockets 53A and 53B. They are so dimensioned using fluid flow resistance calculations methods known to those skilled in the art to obtain a different response of the compensation resistance regions 51A and 5 IB compared to the pocket resistance regions 54A and 54B as the shaft is moved out of concentricity by a load applied to the shaft or housing.
  • the changes in the resistances of the compensation regions 51A and 5 IB and the resistances of the pocket regions 54A and 54B will induce a pressure differential between the pocket regions 53A and 53B that are on opposite sides of the shaft longitudinally to create a restoring force to counter the applied load, as earlier described.
  • the improvement in this modification is the use of external shaft flow channel routing grooves 52A and 52B.
  • Fig. 5 indicates that the pockets 53 A and53B and flow channel routing grooves 52A and 52B are machined or otherwise formed into the surface of the shaft 40 (such as by casting or molding, if appropriate slab materials are used). This is done for ease of manufacture. The effect intended by the present invention could also be achieved by machining or otherwise forming the pockets 53A and 53B and flow channel routing grooves 52A and 52B into the inner surface of the bearing house 41, if desired.
  • Fig. 6 Shown in Fig. 6 is a bearing similar to that previously described except the pockets have been reduced in dimension to such an extent that their width is comparable to the width of the routing grooves and shown in an obtuse angle relative relationship. Since this design invites the use of a plurality or plethora of inherently parallel grooves and pockets, in many bearing applications it can be designed to have higher load capacity than the bearing of Fig.4, and since it has a plurality of pockets, it will obtain more precise rotary motion than the bearing of Fig. 4.
  • the fluid After entering the fluid supply pressure annulus 60, the fluid flows axially across the compensation resistance regions 61 A and 6 IB to the fluid routing flow channel grooves 62A and 62B, whose depth radially into the shaft is large (by at least a factor of 5) compared with the clearance between the shaft 40 and the housing 41. If so dimensioned the fluid will flow freely to the bearing pocket regions 63 A and 63 B whose depths are comparable to that of the flow channel grooves 62A and 62B so as to provide regions of relatively uniform pressure. Leakage reduction regions 66A, 66B, 67A., and 67B are at a small distance radially from the inner surface of the bearing housing 41 (typically 0.01 mm to 0.03 mm).
  • This region acts to substantially reduce leakage flow from circumferentially between the routing grooves 62A and 62B, and circumferentially between the pockets 63A and 63B.
  • the flow leaves the bearing pockets 63 A and 63 B across the pocket resistance regions 64 A and 64B and into the fluid collection annuli 65A and 65B, whose dimensions are similar to those of the supply pressure annulus 60.
  • the fluid exits the bearing to low pressure (typically atmospheric pressure) through holes 10A and 10B drilled radially through the housing 41.
  • Fig. 7 Shown in Fig. 7 is a bearing similar to those previously described except drainage grooves have been added between the pockets, and the shape of the supply pressure annulus has been changed.
  • the supply pressure annulus 80 has been shaped so as to reduce the flow resistance to the routing flow channel grooves 82A and 82B, and in this case would preferably be machined into the shaft if the pockets are machined into the shaft or into the housing if the pockets are machined into the housing.
  • the drainage grooves 85A and 85B have been included between the pocket resistance regions 84A and 84B to eliminate circumferential fluid flow between the pockets that acts to reduce the load-ca ⁇ ying capacity of the bearing. Adding the drainage grooves 85A and 85B increases the flow rate of the fluid through the .
  • the fluid After entering the fluid supply annulus 80, the fluid flows axially across the compensation resistance regions 81 A and 8 IB to the fluid routing grooves 82 A and 82B, whose depth radially into the shaft is large (by at least a factor of 5) compared with the clearance between the shaft 40 and the housing 41. If so dimensioned the fluid will flow freely to the bearing pocket regions 83A and 83B whose depths are comparable to that of the grooves 82A and 82B so as to provide regions of relatively uniform pressure. Leakage reduction regions 86A and 86B are at a small distance radially from the inner surface of the bearing housing 41 (typically 0.01 mm to 0.03 mm).
  • These regions act to substantially reduce leakage flow circumferentially between the routing grooves 82A and 82B.
  • the flow leaves the bearing pockets 83A and 83B across the pocket resistance regions 84A and 84B and into the drainage regions 85A and 85B, where the pressure is low (typically atmospheric pressure).
  • the fluid may then be fed similarly to the previously described bearings to fluid collection annuli and out of the bearing.
  • Figs. 4-7 show the preferred direction of shaft rotation for the bearings contained in this application.
  • the movement of the shaft relative to the housing induces a flow of fluid generally in a circumferential direction and induces a gradient in pressure from the entrance of, for example, the flow channel routing grooves 52A and 52B to the exit of the pockets 53A and 53B, the pressure being lowest at the entrance of the routing grooves and highest at the trailing edges of the pockets.
  • the bearing has been so designed that, when the shaft is rotated in this direction at a high rate of rotational speed, the lowest pressure in the bearing will occur at the entrance of the routing grooves 52A and 52B.
  • the present invention can be used to replace rolling element bushings to obtain a substantial increase in precision guiding performance in applications where the guide rails are cylindrical or arcuate in geometry.
  • the objectives of the invention are thus attained, generally, in a mechanism providing smooth accurate rotaiy motion by means of one or more bearing pad pairs which guide the motion of a shaft 40 about and/or along an axis.
  • the bearings are kept from making mechanical contact with the housing 41 by the thin film of pressurized fluid that flows from sets of opposed recess pockets in selected cylindrical surfaces of the shaft 40 that are surrounded by the housing 41.
  • the flow of fluid to the pockets is regulated along novel flow channels to allow a differential pressure to be established between the pockets, thereby compensating for changes in applied loads to the housing.
  • This fluid flow regulation is provided by applying pressurized fluid in the annular recess grooves from which fluid flows to across respective compensation resistance regions to pockets in which the differential pressures are generated to provide load restoring capacity.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Magnetic Bearings And Hydrostatic Bearings (AREA)

Abstract

Cette invention a trait à un ensemble intégré arbre à broche (1) - ou analogue - et palier hydrostatique auto-compensateur, lequel ensemble évite d'avoir à recourir à des structures externes de gaines de palier. Cet ensemble comporte un alésage cylindrique de palier possédant des gorges circonférentielles (4) en liaison avec des systèmes de mise sous pression (11 et 12) et d'évacuation. Les sources de mise sous pression distinctes et les gorges d'évacuation fluidique (10, 13) coopèrent avec des gorges circonférentielles de collecteur (4) et des poches, de sorte que, lorsque un fluide s'écoule des gorges servant à la mise sous pression (11 et 12) le long de l'arbre (1) pour se rendre dans les gorges du collecteur (4), cet écoulement s'effectue, en rapport avec l'espacement entre la surface de l'arbre (1) et l'alésage, vers la poche opposée à la gorge du collecteur (4), ce qui a pour effet d'assurer une force de rappel proportionnelle au déplacement radial de l'arbre (1).
PCT/IB1997/000081 1997-02-05 1997-02-05 Paliers hydrostatiques auto-compensateurs Ceased WO1998035166A1 (fr)

Priority Applications (2)

Application Number Priority Date Filing Date Title
PCT/IB1997/000081 WO1998035166A1 (fr) 1997-02-05 1997-02-05 Paliers hydrostatiques auto-compensateurs
AU16143/97A AU1614397A (en) 1997-02-05 1997-02-05 Self-compensating hydrostatic bearings

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
PCT/IB1997/000081 WO1998035166A1 (fr) 1997-02-05 1997-02-05 Paliers hydrostatiques auto-compensateurs

Publications (1)

Publication Number Publication Date
WO1998035166A1 true WO1998035166A1 (fr) 1998-08-13

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PCT/IB1997/000081 Ceased WO1998035166A1 (fr) 1997-02-05 1997-02-05 Paliers hydrostatiques auto-compensateurs

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WO (1) WO1998035166A1 (fr)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1207315A1 (fr) * 2000-11-15 2002-05-22 Westwind Air Bearings Limited Palier aérostatique
EP3176450A1 (fr) * 2015-12-03 2017-06-07 Flender-Graffenstaden S.A.S. Palier hydrostatique à fonction hydrodynamique

Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2449297A (en) * 1941-03-26 1948-09-14 James M Degnan Automatic fluid pressure balancing system
US2459825A (en) * 1941-05-24 1949-01-25 Cincinnati Milling Machine Co Bearing
US3508799A (en) * 1968-07-22 1970-04-28 Atomic Energy Commission Gas bearings
FR2214062A1 (fr) * 1973-01-17 1974-08-09 Centraine B K
EP0105050A1 (fr) * 1982-09-30 1984-04-11 Fürstlich Hohenzollernsche Hüttenverwaltung Laucherthal Palier hydrostatique
US4685813A (en) * 1986-07-28 1987-08-11 Moog Inc. Hydrostatic bearing
US5281032A (en) * 1990-11-08 1994-01-25 Alexander Slocum Self-compensating hydrostatic bearings for supporting shafts and spindles and the like for rotary and translational motion and methods therefor
WO1995024569A1 (fr) * 1994-03-10 1995-09-14 Advanced Engineering Systems, Operations & Produc Ts Inc Conception d'une broche hydrostatique a haute vitesse

Patent Citations (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2449297A (en) * 1941-03-26 1948-09-14 James M Degnan Automatic fluid pressure balancing system
US2459825A (en) * 1941-05-24 1949-01-25 Cincinnati Milling Machine Co Bearing
US3508799A (en) * 1968-07-22 1970-04-28 Atomic Energy Commission Gas bearings
FR2214062A1 (fr) * 1973-01-17 1974-08-09 Centraine B K
EP0105050A1 (fr) * 1982-09-30 1984-04-11 Fürstlich Hohenzollernsche Hüttenverwaltung Laucherthal Palier hydrostatique
US4685813A (en) * 1986-07-28 1987-08-11 Moog Inc. Hydrostatic bearing
US5281032A (en) * 1990-11-08 1994-01-25 Alexander Slocum Self-compensating hydrostatic bearings for supporting shafts and spindles and the like for rotary and translational motion and methods therefor
WO1995024569A1 (fr) * 1994-03-10 1995-09-14 Advanced Engineering Systems, Operations & Produc Ts Inc Conception d'une broche hydrostatique a haute vitesse

Cited By (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1207315A1 (fr) * 2000-11-15 2002-05-22 Westwind Air Bearings Limited Palier aérostatique
EP3176450A1 (fr) * 2015-12-03 2017-06-07 Flender-Graffenstaden S.A.S. Palier hydrostatique à fonction hydrodynamique
WO2017092903A1 (fr) * 2015-12-03 2017-06-08 Flender-Graffenstaden S.A.S. Palier hydrodynamique
CN108474413A (zh) * 2015-12-03 2018-08-31 浮林德-格拉芬斯塔登有限公司 流体动力轴承
US10443651B2 (en) 2015-12-03 2019-10-15 Flender-Graffenstaden S.A.S. Hydrodynamic bearing
CN108474413B (zh) * 2015-12-03 2019-11-08 弗兰德-格拉芬斯达登有限公司 流体动力轴承

Also Published As

Publication number Publication date
AU1614397A (en) 1998-08-26

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