HIGH SPEED HYDROSTATIC SPINDLE DESIGN

Abstract: A novel means for the design of the pockets of a hydrostatic spindle and the flow supply, collection, and pocket (15B) injection means to allow for very high speed rotation with a minimum of heat generation, and maximum load capacity; furthermore the concept of self-compensation is used to make the design insensitive to contaminants and thereby allow for the use of water based hydrostatic fluids, whereas the design of the compensator region is also advanced to facilitate high speed use. Accordingly, the present invention allows designers to increase the allowable speed and accuracy of rotary motion hydrostatic bearings. Design methodologies are presented for designing self-compensating hydrostatic bearing spindle pockets and compensators to minimize the effects variation in stiffness with angular position, cavitation in bearing pockets, and turbulent shear power generation. Although it may seem that these are unrelated effects, the design solutions required substantially overlap and are thus presented together here. The primary issue in all of the design advancements is that the flow must be kept from separating, which will induce cavitation. This will result in wear and erosion of the bearing surfaces which will lead to bearing failure. Cavitation is prevented by the use of gently changing shapes, and the introduction of pressure at locations in the bearing where left to itself, the flow would separate and create a low pressure and create a low pressure cavitating region.

Description: High Speed Hydrostatic Spindle Design The present invention relates to bearings that are supported by a thin film of pressurized fluid-liquid or gasecus-bur. preferably water or :he like, bt-iTxiZ more particularly though not exclusively directed to hydrostatic bearings useful for rotary spindles and the like suitable for very high speed rotation with a minimum of heat generation, and maximum load capacity; and exemplifying the concept of self compensation to make the design insensitive to contaminants and thereby allow for the use of such water based hydrostatic fluids. Accordingly, the present invention allows designers to increase the allowable speed and accuracy of rotary motion hydrostatic bearings. Design methodologies are o presented for designing self compensating hydrostatic bearing spindle pockets and compensators to minimize the variation in stiffness with angular position, cavitation in bearing pockets, and turbulent shear power generation, which though different effects, necessitate design solutions that require substantial overlap and are thus presented together here, the primary issue being the flow must be kept from 5 separating, which otherwise will induce cavitation, resulting in wear and erosion of the bearing surfaces and leading to bearing failure. BACKGROUND Hydrostatic bearings have been in use for a very long rime, and recent 0 improvements in compensator design, such as discussed in US patents 5.164,237 and 5,281,032 provided means to allow water (or similar water-derived or related fluids herein genetically referred to as 'water') to be used as a working fluid for the bearings. However, although water is an ideal fluid from a heat transfer perspective, and for reducing die shear forces on the spindle, its lower viscosity makes it more likely to cause turbulent flow. When a flow becomes turbulent, it causes greater viscous shear losses, and die power rate increase with speed can actually rise. In the past, when oil was used, flows rarely became turbulent and because of the high viscosity of oil, they still generated far too much heat to be used at high speeds which might give rise to turbulence. Now with the use of water made possible by die above-cited patents, high laminar speeds can be obtained reasonably, and indeed, even when the flow becomes 0 turbulent, it still only generates half the shear losses of oil. The present invention is concerned with major advancements in optimizing the shape of die bearing regions most effectively to handle high speed water flow which by its nature and the nature of wanting to ran spindles fast, creates a need to run the bearings in a turbulent state. As will also later be shown, the designs for turbulence, also make the bearing more _ 5 robust and more accurate even when they are run with oil or in the non turbulent state. The present invention allows designers to increase the allowable speed and accuracy of rotary motion hydrostatic bearings. Design methodologies are presented for designing self compensating hydrostatic bearing spindle pockets and Q compensators to minimize the effects variation in stiffness with angular position, cavitation in bearing pockets, and turbulent shear power generation. Although it may seem that these are unrelated effects, the design solutions require substantial overlap and are thus presented in combination for optimal results. The key factor is that the flow must be kept from separating, which will induce cavitation. This will result in wear and erosion of die bearing surfaces which will lead to bearing failure. Cavitation is prevented by die use of gentiy changing shapes, and the introduction of pressure at locations in the bearing where if otherwise left to itself, the flow would separate and create a low pressure cavitating region. OBJECT OF THE TNVENTTON Accordingly, it is an object of the present invention to provide new and improved bearing designs for fluid handling regions of hydrostatic bearings that are 0 used primarily for supporting high speed or high accuracy rotating spindles and the like, although the same designs can apply to linear hydrostatic bearings to allow them to be made more efficient and to miniaturize them. Other and further objects are hereafter explained and more particularly delineated in the appended claims. 5 SUMMARY In summary, there is provided a hydrostatic bearing assembly comprising a cylindrical shaft and a concentrically surrounding bearing housing that transfers radial forces, witiiout mechanical contact, between the said shaft and the housing by two pairs of opposed pressurized bearing pockets of fluid filling the gap there between o wherein the fluid flow to a cylindrical bearing pocket on one region of die shaft is regulated by a bearing cylindrical compensation section located at a region of the shaft opposite to said pocket, such that displacements of die bearing caused by applied loads cause die bearing gap to decrease in die region of said pocket, in turn causing the bearing gap to increase in the region of said fluid flow compensation section; said compensation section comprising a centrai aperture region that is connected to said pocket and is copianar with a land region that surrounds said central aperture region, said pocket being surrounded by grooves having gradually sloped side walls and which arc connected to a fluid pressure source, and in which the connection from said compensation section to said pocket, is made at an angle that directs d e flow in a direction substantially tangent to the said shaft and generally in d e direction of d e surface velocity vector of said shaft. Preferred and best mode designs are hereinafter described. DRAWINGS The invention will now be described with reference to die accompanying drawing in which: Fig. 1 is a cross section view through a spindle where the rotor is supported by self compensated hydrostatic bearings made from a sleeve with die geometry features cut in the ED and OD of the sleeve and then die sleeve is pressed into a bore of a housing; Fig. 2 is a top view of a portion of the compensation region of a self compensating bearing that illustrates three types of compensator units, one (A) of the traditional prior art type, and the others of novel through-slit high speed narrow profile type E, and a slit-wedge shaped collector type C; Fig. 3 is a side cross section view of the old and new types of compensator types, where the old type shows how fluid ha!s difficulty entering the hole at high speed while the new type channels the fluid into the slit; Fig. 4 is a cross section of a hydrostatic bearing pocket with a depressed central pocket and further depressed grooves surrounding the pocket to prevent cavitation and to provide an even flow field at the leading edge of die central pocket region; Fig. 5 is a cross section of a hydrostatic bearing pocket with a central pocket region with an elevation equal to tiiat of die surrounding bearing lands and gradually sloped shaped grooves surrounding the pocket to prevent cavitation and to provide an even flow field around the central land region: Fig. 6 is a top view of a single pocket region of what could be a many pocketed bearing, wherein the figure shows how the pocket region can be skewed to form a rhombus shape such that the pockets would dierefore overlap to provide a more uniform stiffness profile for the bearing, and the angled grooves would actively channel the water at high speed which would help to prevent turbulence and cavitation in the grooves, Fig. 7 is a top view of a single pocket region of what could be a many pocketed bearing, wherein the figure shows how the pocket region can be of conventional rectangular form also, but would take advantage of the central pocket region not being depressed but surrounded by the gradual- sloped grooves, Fig. 8 is a polar plot of the radial stiffness profile of a self compensating hydrostatic bearing that shows the lobed stiffness effect of using conventional rectangular pockets; Fig. 9 is a cross section of a spindle system where radial holes bleed some of die high pressure fluid from die high pressure supply to axial holes in the rotor to provide high pressure coolant to a tool (not shown) that would be held to the spindle rotor by a standard means known to those skilled in d e art; Fig. 10 shows a plot of coolant pressure at the center of a shaft as a function of rotor radius and speed. THE INVENTION In general, a precision hydrostatic bearing supported spindle is manufactured from a housing, which bolts to the machine, a set of sleeves that contain the hydrostatic fluid flow control logic that are pressed in to die housing, and a rotating 0 shaft that fits inside die sleeves with a small clearance, typically from 5-50 microns depending on the size of the unit and the pressure and viscosity of the fluid and die speed of the spindle. A hydraulic pump supplies fluid to die bearings, and a motor rums the spindle rotor. There are five distinct, but performance coupled components to the present 5 invention: 1) In a self compensating system, such as described in before-mentioned patents wherein the collector groove of the self compensation system has, instead of a hole at one end, a longitudinal slit along the length of the collector groove to pass through the sleeve and connect witi a groove on the OD of the sleeve to o deliver die fluid to the opposed pocket: and for very high speed, the collector is wedge shaped also widi the center region machined out through the sleeve to encourage the fluid to be collected smoothly and dien in effect be rammed into the passage which would then take it to die opposed pad. 2) In a hydrostatic bearing system, wherein the bearing pad comprises an annuiar land region that encircles a central pocket region, the central region is not a uniform depressed pocket, but rather a plateau of the same height as the surrounding land region, copianar therewith, and the center plateau is surrounded by a channel that maintains die pressure of ±e plateau at a high level as would normally exist in a fully pocketed region. 3) In a hydrostatic bearing system wherein the channei diat surrounds die cena-al region is substantially deeper than the center region, so he flow into the channel, preferably by the tangential injection means described in ±e before- mentioned patents, evenly keeps the pressure in the central region high, and where ±e channel is preferably shaped to have gradually sloped sides, such as a vee channel, to minimize ±e possibility of turbulence and cavitation when ±e fluid dragged by ±e rotor breaks over ±e edge of ±e channel. 4) In a hydrostatic bearing system where ±e bearing pads are normally rectangular in shape, ±e bearing pads of ±is invention are preferably rhombus shaped so ±at portions of ±e pads overlap and ±ereby accomplish two goals: decreased radial stiffness variation with circumferential position, and increased ease of distribution of flow at high speeds because the angled pockets encourage ±e flow to spread evenly from even a single entry point or a slit ±at covers ±e leading edge, as opposed to a rectangular pocket where a single point injection tends to causes ±e flow to fan out ±us leaving some regions of ±e pocket to be at lower pressure ±an o±ers. particularly at high speed. 5) Since high pressure water is used in ±e spindles to support ±e bearings, while high pressure water is often used as a coolant for ±e cutting process, ±en ±e . same supply can be used for bo± if ±e spindle rotor has radial holes cut in it ±at intersect ±e region of ±e bearing where high pressure surrounds the compensators, said radial holes intersect axial holes which are positioned along ±e rotor to deliver ±e high pressure coolant to ±e rip of ±e tool. To describe ±e overall configuration of a typical spindle system ±at could use ±e improvements described herein. Fig. 1 shov/s a cross section view ±rough a spindle where e rotor shaft is suppoπed by self compensated hydrostatic bearings made from a sleeve 60 wi± ±e geometry features cut in ±e ID and OD of ±e sleeve and ±en ±e sleeve is pressed into a bore of a housing 90. Grooves 71a. 72a, and 73a are for collecting water as it flows out of ±e spindle, and thus ±ey are collected to drains. Grooves 80a and 81a would be connected to an external high pressure water system. High pressure fluid from these grooves flows across ±e space between ±em and into ±e collector grooves such as 4b. This groove is similar to ±e groove 4b shown in Fig. 2. but in Fig. 1, the backside of ±e groove is shown which is connected to a helical groove 69a that is machined on ±e OD of ±e sleeve 60. This groove would be connected to a hydrostatic pocket such as 15b in Fig. 6 via a connecting hole 21b. In Fig. 1, a similar connecting from a helical groove on ±e backside of the spindle would be hole 69c ±at connects to ±e leading edge of ±e gendy sloped groove 21c to supply fluid to ±e pocket 21c. Fluid flows across ±e pocket and out ±e lands 13c and 18c to flow out ±e drain grooves 71a and 72a. Fluid ±at flows all ±e way across ±e pocket flows into groove 17c. To proceed wi± ±e discussion of features ±at make ±e subject invention amenable to very high speed rotation, consider Fig. 2 which is a top view of a portion of ±e compensation region of a self compensating bearing ±at illustrates for purposes of comparing side-by-side ±ree types of compensator units, one labeled A of ±e traditional prior-art type, and ±e unit B showing a novel ±rough-slit high speed narrow profile type, and ±e unit C a novel slit-wedge shaped collector type, ±ese showing just a portion of ±e hydrostatic bearing unit more completely detailed in Fig. 6, actually residing next to a hydrostatic pocket, supplying fluid to ±e pocket opposite to its neighboring pocket. Thus this forms ±e feedback loop characteristic of self compensating hydrostatic bearings described in ±e previously cited patents. Returning to Fig. 2, lands la and lb contain ώe high pressure fluid ±at is input to ±e channel 2 from slits 6b and 6a. Slits or angled entrance holes help to reduce turbulence. A traditional self compensator has a land region 3 a and a central collector groove 4a. The fluid rushes to one end of ±e groove where it goes down a hole 5 and to a helical groove on ±e backside of ±e sleeve, such as 60 in Fig. 1, and to ±e opposing bearing pocket. The issue is that at high speed, large recirculating flows occur in ±e groove 4a which mhibit flow into ±e hole 5. To solve ±is problem, in accordance wi± ±e invention, ±e lands 3b surround a groove 4b which is actually a slit cut all the way through ±e sleeve. This allows ±e fluid to fall into ±e helical groove on ±e back of the sleeve, as soon as it enters ±e pocket. Thus at high speeds, ±e flow is far more effectively collected. Also, tiiis helps to allow the compensator wid± to be minimized, because a 1 mm wide slit 20 mm long can handle ±e same flow as a 5 mm diameter hole. Ano±er novel design option shown in Fig. 2C is ±at of a wedge shaped compensator wherein ±e lands 3c are shaped like a wedge, and ±e entire central region 4c is hollowed out to connect wi± ±e helical groove on the backside of ±e sleeve. This type of compensator can run at very high speed and act to collect ±e fluid most efficiendy, but is generally used for spindles ±at have direction in one direction, such as a grinder spindle. In order more fully better to illustrate why ±e cut-through slit is a better design ±en a simple hole, consider Fig. 3 which is a side cross section view of ±e old and new types of compensator types, where ±e old type (A) shows how fluid has difficulty entering ±e hole at high speed while ±e new type (B) channels ±e fluid into ±e slit. In Fig. 3A, ±e fluid is being dragged in ±e gap 8a between the rotor surface 7a and ±e compensator, where it enters ±e groove region 4a. At ±e end of ±e compensator, some of ±e fluid goes down the hole 5 and into ±e helical channel 0 9a. but much of it gets rammed out ±rough ±e gap 8a. The through-slit compensator B, on ±e o±er hand, has ±e spindle rotor 7b dragging ±e fluid through ±e gap 8b and into ±e slit 4b where it has a long space to be directed toward ±e helical groove 9b which takes the fluid off to an opposed pocket. All of the through-slit compensators would also be very useful for linear bearings as well, where the size of 5 ±e compensator needs to be minimized, for example in a narrow profile bearing ±at is intended to be compatible wi± modular linear motion ball bearing units and ±us have extreme space restrictions. Once ±e fluid is collected from ±e compensators and taken to ±e pockets, ±e manner in which it enters ±e pockets is also very important Fig. 4 shows a type of o pocket 19a ±at would be used for non-turbulent flows in ±e pocket but care must be taken from inciting turbulence at ±e leading edge of ±e pocket Thus as ±e rotor 10a drags ±e fluid across ±e gap 11a and it encounters drain groove 12a which has gendy sloped sides, ±e flow is less likely to separate, as it would if ±e groove 12a had a rectangular profile, and ±en it channels into the pocket region. The fluid ±en is dragged across the land 13a and ±e fluid encounters the gendy sloped fluid distribution groove 15a which is deeper ±an the central pocket region 16a. The groove being sloped and deeper encourages ±e high pressure fluid from the compensator via groove 14a ±at is entering ±e pocket via hole 21a to be evenly distributed across ±e face of ±e central pocket region 16a. This high pressure line of fluid also minimizes ±e potential for cavitation. After ±e fluid flows across ±e central pocket region 16a, it encounters ±e other side of the groove system, ±e gendy sloped depressed groove 17a. This groove may also have an entrance hole to it from ±e compensator to allow for bi-directional rotation of ±e spindle. The fluid ±en is gendy directed by the gende slope to flow over ±e land 18a and out ±e bearing, where it would leave the bearing via ±e gendy sloped drainage groove 20a. For a very high speed design where ±e fluid will be turbulent no matter what ±e pocket dep± is, it is actually advantageous to use a design 19b shown in Fig. 5. Here once ±e fluid is collected from ±e compensators and taken to ±e pockets via groove 14b. ±e manner in which it enters ±e pockets is also very important. Fig. 5 shows a type of pocket 19b ±at would be used for turbulent flows in ±e pocket. As ±e rotor 10b drags ±e fluid across ±e gap 11a and it encounters drain groove 12b which has gently sloped sides, ±e flow is less likely to separate, as it would if ±e groove 12b had a rectangular profile, and then it channels into ±e pocket region. Here ±e fluid encounters ±e gendy sloped fluid distribution groove 15b which is deeper ±an ±e central pocket region 16b which is actually at ±e same elevation as ±e lands 13b and 18b. The groove being sloped and deeper encourages the high pressure fluid from ±e compensator via groove 14b ±at is entering ±e pocket via hole 21b to be evenly distributed across ±e face of ±e central pocket region 16b. This high pressure line of fluid also minimizes ±e potential for cavitation. The fluid in ±e region 16b is turbulent, but unlike viscous flow where ±e shear stress is proportional to ±e speed and ±e pocket dep±. which is why non turbulent bearings use deep pockets, a turbulent flow shear stress is dependent on ±e gap- based Reynolds number, and hence ±e pocket is typically greater ±an 3 mm deep. It is better to have ±e pocket have no dep±. This also gready increases ±e squeeze film damping effect and also gives a large hydrodynamic lift at higher speeds. Thus ±is design yields an incredibly high dynamic stiffness. After ±e fluid flows across ±e central pocket region lόb, it encounters e o±er side of ±e groove system, ±e 0 gendy sloped depressed groove 17b. This groove may also have an entrance hole to it from ±e compensator to allow for bi-directional rotation of die spindle. The fluid ±en is gendy directed by ±e gende slope to flow over ±e land 18b and out ±e bearing, where it would leave ±e bearing via ±e gendy sloped drainage groove 20b. In ±is new invention, ±e center region of ±e pocket is no longer depressed 5 ±e way conventional pockets are, but rather it has ±e same elevation as the surrounding land region. This gives ±e spindle very high squeeze film damping and shock overload resistance by virtue of ±e existence of a small gap between ±e raised central portion and ±e rotating rotoπ ±us when a load is suddenly applied, ±e fluid cannot be instantaneously squeezed out and the fluid film prevents ±e rotor from o coming in physical contact wi± ±e bearing housing. This concept has been utilized in linear bearings, but was always ±ought to be bad for spindles because ±e small gap would generate increased viscous shear forces. Indeed, while ±is may be true for oil, for low viscosity water in spindles operated at speeds ±at compete wi± ball bearings, turbulence in ±e pocket would generate more heat than if ±e pocket did not have any increased dep±. The concept here is to keep ±e central portion of ±e pocket at ±e same copianar elevation as ±e bearing lands and to surround it wi± a vee shaped groove ±at prevents turbulence and cavitation in ±e spindle, which keeps ±e central region being held at a high pressure to support a load applied to ±e shaft Now ±at it has been established how ±e fluid is best collected by ±e compensators and ±en distributed to ±e pockets. Fig. 6 illustrates how ±e pockets themselves can be shaped to enhance ±e uniform flow of fluid and discourage cavitation and turbulence ±at would cause erosive wear. Fig. 6 is a top view of a single pocket region of what could be a many pocketed bearing, wherein ±e figure 0 shows how ±e pocket region 16b is skewed to form a rhombus shape such ±at ±e pockets in a multi-pocket bearing would ±erefore overlap to provide a more uniform stiffness profile for ±e bearing, and ±e angied grooves 15b and 17b and side grooves 22 and 21 would actively channel ±e water ±at enters ±em from hole 21b at high speed, which would help to prevent turbulence and cavitation in ±e grooves. 5 These grooves surround ±e central non-depressed pocket region 16b and ±us act to hold it at ±e high pressure as indicated by ±e opposed compensator to ±e pocket In ±is design, ±e central pocket region 16b is also at ±e same elevation as ±e bearing lands 13b and 18b as in Fig. 5. Fig. 6 also shows ±e o±er attributes of ±e system ±at were indicated in Fig. 1. Here ±ere are drain grooves 91 and 93, and leakage o prevention lands 97 and 98 and a pressurized groove 2 to supply ±e compensators. The pressurized fluid flows across ±e compensator land 3b and into ±e through-slit collector 4b to be taken by an unshown helical groove on ±e back of ±e bearing sleeve to a pocket ±at is geometrically opposed to ±e pocket 16b. The rhombus shaped pockets could be easily made using ±e ElectroDischarge Me±od (EDM), which is a implemented on common EDM machines. However, in some cases, for simplicity and lower speeds, it may be desirable to use a standard rectangular pocket as shown in Fig. 7. Here ±ere are similar features as found in Fig. 6. There is an external leakage land 99, and drain grooves 73a, 72a, and 71a. There is a pressure groove 80a and leakage prevention lands 97 and 98. The high pressure fluid in ±e fluid supply groove 80a flows across ±e compensator lands 3b and into ±e ±rough-slit collector 4b to be channeled into a helical groove on ±e OD of ±e bearing sleeve ±at would be taken to a pocket opposite to pocket 16a. Fluid 0 from ±e pockets leaves ±e bearings via drain grooves 71a, 12a, 20a, and 72a, which help to maximize ±e pocket pressure differential in ±e bearing. This results in much greater flow dirough ±is type of bearing, ±an ±rough bearings ±at do not use drains grooves between ±e pockets, but it results in much greater differential pocket pressures which increases ±e load capacity of ±e bearings. The high pressure fluid 5 from ±e compensator enters ±e bearing region via hole 21 ±e fluid flows across ±e centrally depressed pocket region 16a. which note does not need side grooves 21 and 22 as does ±e non depressed pocket region in Fig. 6. The fluid comes to ±e end of ±e pocket region 16a and may ±en encounter a channel 17a which is also gently sloped if ±e bearing is to be bi-directional, or if ±e bearing is unidirectional, as for a o grinder, ±en ±e groove 17a can be eliminated to reduce cost If this is ±e case, ±en ±e end of ±e pocket region 61a must be gendy sloped, and which may actually feed a fan-shaped slit at ±e leading edge of ±e deep gendy sloping edge groove 15a. To illustrate ±e effect of a non-rhombus pocket, Fig. 8 shows a polar plot of ±e radial stiffness profile of a self compensating hydrostatic bearing ±at shows ±e lobed stiffness effect of using conventional rectangular pockets. If rhombus pockets were used, ±en ±is effect would be reduced by 50% or more. This would yield greater spindle accuracies. Thus in summary, ±e invention concerns ±e shaping of ±e pockets in a manner quite unlike any previous design for hydrostatic bearing pockets, where in ±e new invention ±e center region of ±e pocket is no longer depressed ±e way conventional pockets are, but ra±er it has the same elevation as ±e surrounding land region, ±ereby granting the spindle very high squeeze film damping and shock overload resistance by virtue of ±e existence of a small gap between ±e raised central portion and ±e rotating rotor. Fur±ermore, to maintain high hydrostatic pressure over ±is central region of ±e pocket, it is surrounded by a groove ±at has sloped sides ±at minimize ±e chance of flow separation and cavitation ±at could o±erwise erode ±e pockets, and ±us since ±e central region is at a small gap and it is surrounded by a smoo± gendy sloped groove, turbulent flow, which will happen at high speeds particularly when water is used, will cause far less drag forces ±an if a deep pocketed highly turbulent flow design were used. In addition, ±e pocket may be shaped like a rhombus, so ±at adjacent pockets actually overlap, which increases accuracy and encourages the flow in ±e groove surrounding the pocket to flow around ±e central land region wi±out causing recirculating flow patterns which lead to increased turbulence and erosive wear. Fur±ermore, ±e said surrounding groove is fed high pressure by self compensation means to support a loaded shaft, where ±e high pressure is delivered in proportion to ±e load on ±e shaft by means of a self compensation system where a zone of ±e bearing adjacent to ±e pockets surrounds compensation units by high constant pressure from a pump which flows across a land region that encloses a collector groove, or slit, ±at is cut all ±e way through ±e bearing sleeve ±at ±e pockets and compensators are manufactured into. The sleeve is shrink fit into a large housing to provide a good seal and ±erefore enclosing helical grooves on ±e back of ±e sleeve. Said slit ±at has been cut through connects to a helical groove ±at wraps around ±e back of the sleeve and ±en connects wi± a hole or a slit to ±e groove on ±e opposite side of ±e bearing ±at is surrounded by a drain groove region and an elevated land region. Tne groove itself ±en surrounds a central elevated region, which has ±e same elevation as ±e land region, so ±e central region is ±en at ±e same pressure as ±e pocket and hence supports ±e load on ±e shaft Therefore when ±e rotor is turning and a load is applied, diroush cutting forces for example, ±e gap between ±e compensator and ±e rotor increases on one side of ±e rotor and decreases on ±e o±er side of the rotor, so more fluid flows into one compensator ±an ±e o±er, and ±e fluid flow to one pocket is greater ±an to ±e o±er and ±e differential pressure across e pockets acts to balance ±e load on ±e rotor. In ±e case when ±e rotor is spinning at a high rate, ±e fluid ±at flows into ±e compensator immediately falls into ±e slit and through ±e sleeve into ±e groove ±at takes it to ±e opposed pocket region. Thereby, wi± ±e use of ±e through-slit, ±e high speed rotor does not drag ±e fluid to ±e end of ±e collector groove at ±e center of ±e compensator where it would o±erwise form an energy robbing pressure decreas g recirculating flow. Fur±ermore. and because ±e efficiency of ±e system is increased, it becomes plausible to increase ±e power of ±e pressure supply unit and allow for some of ±e high pressure fluid to bleed off through radial holes in ±e rotor and ±en for ±em to meet wi± axial holes ±at supply high pressure fluid to ±e tool held in ±e spindle, ±ereby obtaining integral through-±e-tool cooling. Depending on ±e degree of speed and load capacity desired and the cost that is to be spent, any or all of these design advances can be combined as desired. Because water is being used as the hydrostatic fluid, and because water is also used as a cutting fluid, with appropriate additives, in many applications, it may be desirable to use the same fluid for the bearings and for the tools. This would minimize the cost of the spindle, because currently, ball bearing spindles use a very expensive coupling to provide ±e spindle with high pressure oil to lubricate the bearings under their races, and high pressure water based coolants to cool ±e tool and wash away chips. Fig. 9 shows a schematic cross section of a spindle system where rad al holes 95 and 98 bleed some of the high pressure fluid from ±e high pressure supply groove 81a and 80a to axial holes 96 and 97 in ±e rotor 50 to provide high pressure coolant to a tool (not shown) that would be held to the spindle rotor by a standard means known to those skilled in ±e art. The spindle has ±e same construction otherwise as shown in Fig. 1, wi± a housing 90, sleeve 60 for ±e bearing logic to be machined in before being shrunk into ±e housing, and drain grooves 71a, 72a and 73 a. The pressure (in metric units) at any inner and outer radius combination can be calculated from ±e following, where p is ±e density of ±e fluid, ω is the speed, and τ\ and X2 are ±e spindle OD and axial hole location radius respectively: p = - p available s centrifugal ^centrifugal Fig. 10 shows a plot of coolant pressure at ±e center cf a shaft as a function of. rotor radius and speed. If ±e coolant supply was 1500 psi, as is ±e case for a very high performance spmdle where ±e bearing shear area is to be minimized to minimize shear power losses, and if ±e supply pressure groove was at a radius of 100 mm, then at 20.000 rpm, 950 psi would De used up by ±e centrifugal pressure generated. This wcuid still leave a respectable 650 psi for the through ±e tool coolant system. Note ±at when e spindle is not turning, but ±e hydrostatics are on, coolant would be streaming out of ±e spmdle. Thus it may be necessary in some instances to provide a centrifugal force activated valve 101 and 102 in Fig. 9 ±at only allow fluid o into ±e tool at fairly high predetermined speeds. Fuπher modifications of ±e invention will also occur to persons skilled in ±e art, and all such are deemed to fall within ±e spirit and scope of ±e invention as defined by ±e appended claims. 5

Claims: CLAIMS 1. A hydrostatic bearing assembly comprising a cylindri¬ cal shaft and a concentrically surrounding bearing housing that transfers radial forces, without mechanical contact, between the said shaft and the housing by two pairs of opposed pressurized bearing pockets of fluid filling the gap therebetween where¬ in the fluid flow to a cylindrical bearing pocket on one region of the shaft is regulated by a bearing cylindrical compensation section located at a region of the shaft opposite to said pocket, such that dis- placemencs of the bearing caused by applied loads cause the bearing gap to decrease in the region of « said pocket, in turn causing the bearing gar> to increase in the region of said fluid flow co pensa- tion section; said compensation section comprising a central aperture region that is connected to said pocket and is copianar with a land region that surrounds said central aperture region, said pocket being surrounded by grooves having gradually sloped side walls and which are connected to a fluid pressure source, and in which the connection from said compensation section to said pocket, is made at an angle that directs the flow in a direction sub¬ stantially tangent to the said shaft and generally in the direction of the surface velocity vector of said shaft. 2. A hydrostatic bearing assembly as claimed in claim 1 and in which the shaft comprises a rotary spindle. 3. A hydrostatic bearing assembly as claimed in claim 2 and in which the said gap is sufficiently small to provide the spindle with a high squeeze film damping and shock overload resistance. 4. A hydrostatic bearing assembly as claimed in claim 3 and in which the said gradually sloped side walls are adjusted to minimize the chance of fluid separa- tion and cavitation and produce less drag forces i'n response to flow turbulence. 5. A hydrostatic bearing assembly as claimed in claim 2 and in which said grooves are longitudinal slit openings . 6. A hydrostatic bearing assembly as claimed in claim 3T and in which the slit openings are of wedge shape. 7. A hydrostatic bearing assembly as claimed in claim 5 and in which helical grooves are provided around the back of a bearing surface sleeve with which the collector groove slits communicate through the bear¬ ing sleeve and which connects with a groove on the opposite side of the bearing. 8. A hydrostatic bearing assembly as claimed in claim 3 and in which the bearing padsj each comprising the central pocket region and an annular encircling land region, are of substantially rhombus shape so that portions of the pads may overlap to decrease radial stiffness variation with circumferential position and to cause more even fluid flow spread. 9. A hydrostatic bearing assembly as claimed in claim 2 and in which a common fluid supply is used for the spindle shaft cooling and bearing operation, with ; apertures formed in the shaft that intersect the region of the bearing where the high fluid pressure surrounds the compensation section. 10. A hydrostatic bearing assembly as claimed in claim 2 and in which the same is clustered in a head with one or more further similar hydrostatic bearing spindle assemblies. 11. A hydrostatic bearing assembly comprising a pair of opposing bearing surfaces , one bearing surf ce hav¬ ing opposed pressurized bearing pockets of fluid filling the gap between the surfaces wherein the fluid flow to a bearing pocket on one region of said one surface is regulated by a bearing compensation section located at a region opposite to said pocket, such that displaceme ts of the said one bearing sur¬ face caused by the applied loads cause the bearing gap to decrease in the region of said pocket, in turn causing the bearing gap to increase in the region of said fluid flow compensation section; said compensation section comprising a central aperture region that is connected to said pocket, surrounded with a land region, said pocket being surrounded by grooves having gradually sloped side walls and which are connected to a fluid pressure source, and in which the connection from said compensation section to said pocket, is made at an angle that direccs the flow in a direction substantially tangent to the said one bearing surface. 12. A hydrostatic bearing assembly as claimed in claim 11 and in which the bearing portions comprising the central pocket region and the surrounding land region are of substantially rhombus shape. 13. A hydrostatic bearing assembly as claimed in claim 11 and in which the said grooves are one of longitudinal 5 slit and slit-wedge ooenings. 14. A hydrostatic bearing assembly as claimed in claim 11 and in which a common fluid supply is used for the bearing surface cooling and the bearing operation. 15. A hydrostatic bearing assembly as claimed in claim 11 10 and in which the bearing surfaces comprise a cylindrical shaft and a concentrically surrounding t bearing housing. 16. A hydrostatic bearing assembly as claimed in claim 11 and in which the bearing surfaces comprise a pair of j_5 opposed linearly extending surfaces. 20

Inventors: SLOCUM ALEXANDER, SLOCUM, ALEXANDER
Applicants: SLOCUM ALEXANDER, TERADYNE INC, ADVANCED ENGINNEERING SYSTEMS, OPERATIONS & PRODUC, SLOCUM, ALEXANDER
Application year: 1995
Main classification: F16C32
IPC codes: F16C29, F16C32, F16C37