US3827337A

US3827337A - Hydrostatic bearings for the swash plate of a barrel-cylinder hydraulic pump or motor

Info

Publication number: US3827337A
Application number: US00239290A
Authority: US
Inventors: F Pruvot
Original assignee: Renault SAS
Current assignee: Renault SAS
Priority date: 1971-04-28
Filing date: 1972-03-29
Publication date: 1974-08-06
Anticipated expiration: 1991-08-06
Also published as: DE2220847C3; FR2135776A5; DE2220847B2; DE2220847A1

Abstract

This hydrostatic bearing for the swash plate of a hydraulic high-pressure and cylinder barrel machine is formed on a face of a bearing plate which registers with one face of a swash plate, and comprises a main bearing consisting of a set of two arcuate grooves surrounded by sealing surfaces and disposed in proper phase relationship to ports formed in the fluid distributor plate, and another set of grooves in the other face of said swash plate, identical in number to that of distributor orifices formed at the end of the barrel cylinders, said bearing further comprising at least three auxiliary hydrostatic bearings consisting of cavities formed in the other face of said bearing plate, said last-named cavities being concentric to the grooves of said main bearing.

Description

[ Aug. 6, 1974 3,040,672 6/1972 Foersten et 91/487 3,073,253 1/1963 Schollhammer........... 91/485 3,635,126 1/1972 Engel.........,...,.......................91/486 Primary Examiner-William L. .Freeh Attorney, Agent, or Firm-Stevens, Davis, Millcr & Mosher [57] ABSTRACT This hydrostatic bearing for the swash plate of a hydraulic high-pressure and cylinder barrel machine is formed on a face of a bearing plate which registers with one face of a swash plate, and comprises a main bearing consisting of a set of two arcuate grooves surrounded by sealing surfaces and disposed in proper phase relationship to ports formed in the fluid distributor plate, and another set of grooves in the other face of said swash plate, identical in number to that of distributor orifices formed at the end of the barrel cylinders, said bearing further comprising at least three auxiliary hydrostatic bearings consisting of cavities formed in the other face of said bearing plate, said United States Patent [191 Pruvot 1 HYDROSTATIC BEARINGS FOR THE SWASH PLATE OF A BARREL-CYLINDER HYDRAULIC PUMP 0R MOTOR [75] Inventor: Francois C. Pruvot, Billancourt,

France [73] Assignee: Regie Nationale Des Usines Renault, Billancourt, France [22] Filed: Mar. 29, 1972 [21] Appl. No.: 239,290

[30] Foreign Application Priority Data Apr. 28, 1971 France.............................. 71.15146 [52] 11.8. 91/489, 91/504 [51] Int. FOlb 13/04 [58] Field of Search....................... 91/485, 487489 [56] References Cited UNITED STATES PATENTS last-named cavities being concentric to the grooves of said main bearing.

4 Claims, 16 Drawing Figures Pratt 4/1939 Thoma.. 10/1942 Vickers..... 2/1961 Douglas....

3/1961 Schoellhammer HYDROSTATIC BEARINGS FOR THE'SWASH PLATE OF A BARREL-CYLINDER HYDRAULIC PUMP OR MOTOR The present invention relates in general to hydraulic machines such as motors and pumps comprising a plurality of cylinders disposed in a barrel-shape block, wherein the fluid is distributed to the various cylinders through a distributor plate of the flat or revolution type. The term pump will be used hereinafter to designate a machine according to this invention capable of operating either as a motor or as a pump.

This invention is concerned more particularly with a hydrostatic bearing adapted to balance the impeller plate of a pump of the type mentioned hereinabove and wherein the cylinder or cubic capacity of the pump is varied by changing the angle measured between the cylinder barrel axis and the pump driving shaft axis, so that a zero angle corresponds to a zero capacity.

As a consequence of the specific mode of construction of pumps of thisgeneral type, the impeller plate rigid with the driving shaft is caused to support the thrust from the axial pistons. This thrust comprises a radial component and an axial component, as those skilled in the art will readily understand.

As a rule, the axial effort is absorbed by one or a plu rality of ballor roller bearings, although pumps types are known wherein plain, hydrodynamic or hydrostatic bearings are used.

More generally, the axial thrust is absorbed by rolling contact thrust bearings whether of the ball-, needleor roller type. In other constructions this thrust component is absorbed by hydrostatic bearings similar to the bearing described and illustrated in the US. Pat. application Ser. No. 42,71 I filed by the same applicant on June 2, 1970 now US. Pat. No. 3,702,576.

In another pump type the swash plate is not driven directly by the driving shaft and the piston rods are nearly perpendicular thereto. In this known arrangement the radial efforts are absorbed by the barrel and the very moderate radial forces act upon the swash plate, thus permitting of dispensing with the bearings usually provided for absorbing these forces.

In the above-mentioned pump types it will be seen that the radial and axial stresses applied to the swash plate vary as a function of the pump angle, i.e., the angle formed between the barrel axis and the perpendicular to the swash plate, even if the pressure remains constant.

In pumps comprising means for modifying at will the inclination of the axis of rotation of the barrel and wherein the swash plate is connected directly to the pump shaft, the axial stress received by the barrel is substantially constant under constant pressure conditions, irrespective of the barrel angle, since said axial stress is subordinate to the number and diameter of the pistons.

This made it possible to balance the axial stress received by the barrel by means of simple hydrostatic bearings. As a rule, the axial forces developing in the cylinders by the working pressure are balanced by properly shaping the cross-sectional contour of the high-pressure port in the distributor plate. A slight degree of unbalance is nevertheless preserved as a function of the magnitude of the leakages to be tolerated.

However, this solution is ill adapted to the pump swash plate, for the axial force received thereby varies substantially as the cosine of the angle of inclination of the pump barrel. Now the necessity of reducing the manufacturing cost, of keeping weight at a low level and performances at a high level, leads to a substantial, increasingly great angle of barrel inclination. In certain pumps this angle is at present as high as 45 corre sponding to an axial thrust exerted against the swash plate which varies in proportion to the cosine of an angle ranging from 0 to 45, i.e., a cosine of l to 0.707.

Under these conditions it is clear that the force necessary for hydrostatically compensating the swash plate cannot exceed 0707 times the sum of the piston thrusts (representing approximately 65 percent of this thrust, since a residual force must constantly be available for keeping the slide faces in mutual, leak-tight contact). A complementary axial bearing should be provided for absorbing an effort corresponding to 35 percent of the piston thrusts when the cylinder capacity of the pump is zero. Of course, these values apply only to a 45 angle of barrel inclination.

To compensate the variation in the axial force during the change in the barrel inclination angle it is possible for example to modify the feed pressure of the hydrostatic bearing associated with the swash plate by means of a valve, so that the variations in the bearing feed pressure correspond substantially to the variations in the axial thrust.

Although this compensating device is based on a relatively simple principle, it is attended by several inconveniences, such as:

an increment in the cross-sectional dimensions of the pump, since the pressure modulating valve or valves must be disposed along the pivot axis of the barrel support;

additional machining operations preventing an economical use of this bearing type, and

the permanent possibility for the hydrostatic bearing feed pressure to be either too low or too high, in case frictional contacts develop in anyone of these valves.

If the pressure is too high, it can be reduced by using elements adapted to produce load losses and disposed in the duct means supplying pressure fluid to the grooves of the hydrostatic bearing, which engage the bearing or sliding surface of the swash plate. These load losses cause the pressure to decrease as the output increases, thus preventing any swash plate instability. On the other hand, if the feed pressure is too low, hydrodynamic bearings similar to those equipping the cylinder barrel are used in order to absorb the non-compensated additional thrust. This hydrostatic bearing feed pressure variation effect constitutes in any case an inconvenience.

It is'therefore the primary object of the present invention to provide a hydrostatic bearing adapted to balance the axial thrust of the swash plate in an axial-type, rotary cylinder-block hydraulic pump or motor, whether of the type comprising in-line axial pistons or canting cylinder blocks, with or without connectingrods.

More particularly, this invention is directed to the provision of an economical hydrostatic bearing fed from the pressure generated by the pump or the pressure applied to the motor, irrespective of the pressure value and the angle of inclination of the swash plate or pump axis in relation to the axis of rotation of the cylinder barrel.

By way of example the pump proper may be similar (except for the pressure adjustment device) to the pump described in the US. Pat. application Ser. No. 40,541 filed on May 26, 1970 now US. Pat. No. 3,661,055. The swash plate of the pump is supported by a radial bearing of which the median plane will be merged if possible into the plane containing the centres of the ball-joints of the piston connecting-rods, which are mounted in said swash plate.

The pump may also be of a type similar to-the pump type without piston connecting-rods, described and illustrated in the U.S. Pat. application Ser. No. 151,450 filed on June 9, 1971.

The hydrostatic bearing structure for the swash plate of high-pressure cylinder barrel hydraulic machines, which comprises a fluid distributor plate formed with ports communicating with the barrel cylinders, pistons slidably mounted in said cylinders and connected through ball-joints to said swash plate, is characterized in that it is formed essentially on one face of a plate and consists of a main bearing comprising including on the one hand a first set of two arcuate shallow grooves formed in said one face registering with a homologue face of the swash plate, said grooves being surrounded by sealing surfaces and disposed in proper phase relationship to the ports formed in the pump fluid distributor plate, and on the other hand, a second set of grooves formed in said face of said swash plate, said last-named grooves being shorter than those of the first set and equal in number to distributor orifices provided at the end of the pump barrel cylinder and in proper phase relationship therewith, said hydrostatic bearing further comprising at least three elementary auxiliary hydrostatic bearings consisting of cavities formed in said face of the hydrostatic bearing plate and disposed concentrically to the grooves of the first set of said main bearing and on a greater radius than the radius of said grooves of the main bearing.

Other features characterizing this invention will appear as the following description proceeds with reference to the attached drawing illustrating diagrammatically by way of example a balancing or compensating device according to this invention. In the drawing:

FIG. 1 is a longitudinal view showing a pump equipped with the hydrostatic bearing according to this invention;

FIG. 2 is a plane elevational view showing the hydrostatic bearing distributor plate, as seen in the direction of the arrows II-I1 of FIG. 1;

FIG. 2b is a plane view showing diagrammatically the fluid supply to the hydrostatic bearings of the distributor plate of FIGS. 2 and 3;

FIG. 3 is a plane elevational view of the hydrostatic bearing distributor plate as seen in the direction of the arrows III-III of FIG. 1;

FIG. 4 is a plane elevational view of the swash plate as seen in the direction of the arrow IV-IV of FIG. 1;

FIG. 5 is a plane view showing a modified form of embodiment of the hydrostatic bearing distributor plate;

FIG. 6 is a plane view showing diagrammatically the fluid supply to the hydrostatic bearings of the distributor plate of FIGS. 2 and 3 in the case of a pump adopted to operate as a motor;

FIG. 7 is a plane view showing diagrammatically another modified form of embodiment of the method of supplying fluid to the hydrostatic bearings of the distributor plate of FIG. 6;

FIG. 8 is a simplified and modified form of embodiment of the plate shown in FIG. 7, in which a duct selector is used;

FIG. 9 is a typical exemplary form of embodiment of the non-return valves of FIG. 7;

FIG. 10 is a section taken along the line X-X of FIG. 9; and

FIG. 11 is another modified form of embodiment of the distributor plate of the hydrostatic bearing shown in FIG. 2; and

FIGS. 12 to 15 are sections taken respectively along lines l212, 13-13, l414 and 15-15 of FIG. 2.

Referring first to FIG. 1, the swash plate 1 of a hydraulic barrel pump is guided radially by a needle bearing 2 and is rotatably assembled through splines with a driving shaft 3. A universal joint 4 connects this shaft 3 to the cylinder barrel 5 rotatably mounted in antifriction bearings 6 and inclined to the shaft axis. Pistons 7 are slidably mounted in the cylinders or bores 8 of barrel 5 and operatively connected through their rods 9 to the swash plate 1, the end ball-joints or heads 10 of these connecting-rods being fitted in part-spherical recesses 11 formed in one face of swash plate 1.

The opposite face 12 of swash plate 1 bears against a slide face 13 of a plate 14 of a hydrostatic bearing structure according to the present invention.

This hydrostatic bearing structure is adapted to withstand the axial stress transmitted to swash plate 1 and adjusts the axial position of swash plate 1 in order to eliminate frictional contacts between the face 13 of bearing plate 14 and the corresponding face 12 of swash plate 1, and also for maintaining a sufficient relative spacing between these faces in order to avoid an excessive exposure thereof to the fluid pollution and limit pump leakages to reasonable values. Finally, the bearing must preserve all these properties during variations in the pressure and angle of cant of the pump barrel.

To this end, the swash plate 1 of which one face 12 is illustrated in FIG. 4 comprises a set of arcuate shallow grooves 15 to be subsequently referred to as the second set. The projections of the centres of balljoints 10 on this face 12 are substantially coincident with the centres of the corresponding grooves 15 and these grooves are exposed to the same pressure as the fluid in the corresponding cylinder 8. To this end, a hole 16 formed in the piston, the passage 17 in connecting-rod 9 and another orifice 18 in swash plate 1 are used, this last-named orifice l8 interconnecting the cavity 11 of ball-joint 10 to the arcuate groove 15.

As shown in FIG. 2, a first set of two arcuate shallow grooves or

cavities

19, 20 are formed in the face 13 of bearing plate 14 of this hydrostatic bearing; these grooves or

cavities

19, 20 are bound radially by arcs having radii R, and R respectively, their lengths being equal to those of the two radii limiting radially the arcuate grooves 15 of swash plate 1 in FIG. 4.

The arcuate grooves or

cavities

19, 20 of the first set have the same angular arrangement or disposition as the ports 21, 22 of distributor plate 23 contacting the barrel 5 (FIG. 1). However, these grooves are slightly shorter so that the fluid is distributed only through the orifices 24 formed at the end of barrel 5 and registering with ports 21, 22 of distributor plate 23. The

grooves

19, 20 of bearing plate 14 of the hydrostatic bearing are surrounded by sealing surfaces 5,, S The same applies to the distributing ports 21, 22 of plate 23.

It is known that as a rule the force pressing the cylinder barrel of a piston pump against its distributor plate is balanced to the extent of 90 to 95 percent (although lower and higher values have already been recorded) by the reaction forces due to the fluid pressure, said reaction forces acting upon the barrel in distributing ports and under the sealing surfaces of said plate.

In the case of the hydrostatic bearing of a swash plate a considerably lower balancing rate is selected. In the present example it is contemplated to balance 70 per cent of the piston thrust when the barrel axis has its maximum cant value.

Since the axial force applied to the swash plate varies substantially as the cosine of the canting angle of the barrel, the force necessary for hydrostatically balancing the swash plate cannot exceed 0.866 times the sum of the piston thrusts, the maximum canting angle of the barrel being assumed to be equal to 30 (cos 0.866).

The reaction force of grooves 19, and sealing surfaces S and S mentioned hereinabove, with a maximum canting angle of 30, is therefore T 70 percent X 0.866, i.e., 60.62 percent of the piston thrust.

In other words, the pressure acting upon the swash plate will balance 70 percent of the apparent piston thrust at the maximum canting angle (30). and about 60 percent of the actual piston thrust if the barrel canting is zero. Under these conditions, auxiliary bearings adapted to absorb the remaining percentage of the piston thrust must be provided, in addition to the abovedescribed hydrostatic bearing.

These auxiliary bearings are also hydrostatic bearings and due to the relatively moderate variation in the force to be compensated the thickness of the fluid film between the swash plate and these bearing varies but very moderately during the variation in the barrel canting angle.

As shown in FIG. 2, the residual effort of the piston thrust is absorbed by four elementary auxiliary

hydrostatic bearings

25, 26, 27 and 28.

Basically, three elementary auxiliary hydrostatic bearings would be sufficient, notably if the pump orifices connected to the high-pressure and low-pressure sides are constantly the same. In this case, the three auxiliary

hydrostatic bearings

29, 30 and 31 of a bearing plate 14' are disposed as shown in FIG. 5. The first and second elementary

hydrostatic bearings

29 and 30 are supplied in this case from the high-pressure groove 20 through an an element (not shown) producing a laminar loss of pressure. The surface of these two

bearings

29, 30 is such that they absorb the residual effort under a pressure considerably lower than the rated pressure of the pump.

A third auxiliary hydrostatic bearing 31 is also supplied with pressure fluid through the high-pressure groove 20 and its surface is reduced for its load is considerably lower than that applied to the other auxiliary

hydrostatic bearings

29 and 30. It will be noted that the third bearing 31 may also be supplied through the lowpressure groove 19, provided however that the cramming" pressure of the pump is high enough. In this case the surface of the third bearing 31 should be definitely greater. Actually, the arrangement of these three bearings must be such that the point of application of the force to be balanced lies within the polygon containing the centres of gravity of the surface areas of the auxiliary hydrostatic bearings.

This arrangement is also applicable to pumps having two directions of rotation, even if they are operated as motors, provided only that the high pressure groove be always the same.

FIG. 6 illustrates diagrammatically a bearing plate 114 of hydrostatic bearing of a swash plate, in the case of a pump adapted to operate as a motor, in either direction of rotation, but wherein the barrel canting angle may vary from 0 to a maximum value on one side only of the pump shaft axis.

The auxiliary bearing 125 is supplied with fluid from groove via an element 132 producing a loss of pressure; this bearing is also connected to the groove 119 of the main bearing through another element 133 also producing a loss of load. The auxiliary bearing 126 is supplied with fluid from the same groove 119 through an element 134 producing a loss of load; this bearing also communicates with groove 120 through an element 13S producing a loss of load. Bearing 127 is supplied from groove 119 via an element 136 producing a loss of load and bearing 128 is supplied from groove 120 via an element 137 also producing a loss of load.

It is clear that whatever the

groove

119 or 120 under pressure, three bearings are constantly supplied directly with high-pressure fluid.

Experience teaches that in the case of a pump comprising seven pistons it is advantageous that the surface area of each auxiliary

hydrostatic bearing

25, 26, 27, 28 lies within the range of 1 to 1.5 times the crosssectional area of a piston.

The pressure in the various hydrostatic bearings in operation will then lie in the range of 12 to 21 40 percent of the operating pressure of the pump. The leakage outputs of each hydrostatic bearing are then relatively low with respect to the pump output and without any appreciable influence on its volumetric efficiency. The elements producing the aforesaid losses of load are easy to manufacture, as will be explained presently, without allowing them to become abnormally sensitive to the fluid pollution.

During the variation in the angle of cant of the cylinder barrel 5 the corresponding variations in the distance between the hydrostatic bearings and the swash plate will be of the order of 0.00l millimeter.

As shown in FIG. 7 illustrating another form of embodiment of four auxiliary hydrostatic bearings the bearing plate 214 of the hydrostatic bearings of the swash plate comprise two

auxiliary bearings

227 and 228 fed from

grooves

219 and 220, respectively, through

elements

236, 237 producing the losses of pressure. The other two

bearings

225, 226 are fed through the high-

pressure groove

219 or 220 through the medium of

corresponding elements

233, 234 producing losses of pressure, and also of

check valves

238, 239 adapted to isolate the low-pressure groove.

Referring to FIG. 8, it will be seen that the

non-return valves

238, 239 of FIG. 7 can be replaced with a fluidway selector 240.

In all the cases contemplated hereinabove the arrangement may be such that all the elements producing losses of load be substantially the sam'es in a same pump. This requirement, which is not compulsory, as will be explained presently, may lead to a substantial simplification in the manufacturing process. By way of example, the elements producing the losses of pressure may be constructed by using cylindrical inserts comprising a helical groove.

In the case of mass production it is known that it is advantageous to use sintered metals for making the bearing plates 14 or distributor plates 23, whether on the barrel side or on the swash plate side. In fact, the front face 13 of distributor plate 14 of the hydrostatic bearing may be used as obtained from the sintering pro cess, without any machining operation, for the balancing rate due to the bearings is relatively moderate so that the diameters of the auxiliary hydrostatic bearings 25 to 28 or of the sealing surfaces 8,, S need not be machined within extremely accurate tolerances.

As shown on FIGS. 2 and 3 and on the sectional view of FIGS. 12 to 15,

orifices

41 and 42 interconnect the grooves 19, of front face 13 of FIG. 2 of bearing plate 14 of the hydrostatic bearing to the rear face 43 of FIG. 3. These orifices open into

supply ducts

44 and 45 respectively and the fluid is directed to

elements

46, 47, 48, 49, respectively, consisting of a pair of circular grooves concentric to the axis of the swash plate and formed in the rear face 43 of plate 14 of said hydrostatic bearing, these elements producing losses pressure along the fluid path as shown in the diagram in FIG. 2b.

Orifices

50, 51 connect the

elements

48, 46 aforesaid to the cavities of the auxiliary hydrostatic bearings 26 and of front face 13. The flow of fluid from port 19 to the auxiliary fluid bearing 25 passes through orifice 41, duct 44, groove 49 and orifice 51 and to auxiliary bearing 26 through orifice 41, duct 44, groove 48 and orifice 50. Similarly, the flow of fluid from port 20 passes through orifice 42, duct 45, groove 46, orifice 51 to auxiliary bearing 25, and through orifice 42, duct 45, groove 47 and orifice 50 to auxiliary bearing 26.

Similarly,

other orifices

52 and 53 connect the

grooves

19, 20 of front face 13 to the rear face 43 of plate 14. The fluid is fed through

ducts

54, 55 formed in said face 43 and lead into

orifices

56, 57 opening in turn into the front face 13.

Duct means 58, 59 formed in the front face 13 and

orifices

60, 61 return the fluid to the rear face 43; these last-mentioned

orifices

60, 61 open at a radius greater than that of

groove elements

46, 47, 48 and 49. The outlet orifices of the holes formed in the rear face 43 continue as

semi-circular grooves

62, 63 concentric to grooves 46 to 49, thus producing losses of pressure and supplying fluid to

orifices

64, 65 opening in turn into the cavities of the auxiliary

hydrostatic bearings

28, 27 formed in the front face 13 of bearing plate 14.

To simplify the drawing all the surfaces of FIGS. 2 and 3 which are flush with the front face 13 or rear face 43 are shown with slight hatches.

A similar procedure could be used for obtaining the elements corresponding to FIG. 7 which produce the losses of pressure. As shown in FIGS. 9 and 10, the

non-return valves

238, 239 may consist for instance of balls 66 of a diameter slightly inferior to the thickness of bearing plate 214 and opening on the rear face 243. The cavities 67 and ball-valve seats 68 are formed directly in the sintered pieces.

The losses of load or pressure are produced in most instances by grooves having width of 0.3 to 1 mm, according to the pump size. These grooves may be either formed directly by sintering of machined on the rear face 43 of plate 14 (or 114, 214). In this last case, the angular extent of the grooves producing losses of load are preferably equal or whole multiples of one another, so that they can be formed simultaneously by using several tools.

In this case a primary requirement is that the rear face 43 of plate 14 be in conformity with the surface to be engaged by this face 43, so that the radial leakages are kept within negligible limits. This requirement is easily met by resorting to conventional manufacturing or production methods.

The proper operation of the hydrostatic bearing depends mostly on the stability of its bearing plate 14. In fact, if the front face 13 of this plate receives the total thrust frompistons 7, the various areas of its rear face 43 will be under pressure. To ensure the stability of plate 14, the sum of the forces applied to its rear face must be inferior to the piston thrust and the moment of the forces exerted on this face must be lower than the resultant moment of the piston thrust, so that the plate cannot tilt. The equilibrium of plate 14 is obtained by laterally limiting the sealing surfaces surrounding the elements producing the losses of pressure and the fluid inlet ducts.

The hydrostatic bearing according to this invention is advantageously applied to axial pumps of the type described for example in the U.S. Pat. application Ser. No. 151,450 filed on June 9, 1971, already mentioned hereinabove. In this pump, the piston shoes engage a rotary plate of which the lower face bears against a hydrostatic bearing of a swash plate. Since the maximum canting angle of the swash plate of these pumps is somewhat smaller (about 20) an improved balancing of the grooves may be expected.

Of course, the specific application described and contemplated herein should not be construed as limiting the present invention. Thus, for instance, the

grooves

19, 20 of plate 14 may be supplied either from an external source of fluid under pressure, or from the pressure supplied by the pump itself, but without causing the fluid to flow through the hollow connectingrods. Many other variations, which cannot be described in detail herein, will readily occur to those skilled in the art without departing from the scope of the invention.

Other shpaes and arrangements of elements adapted to produce losses of load may be contemplated and provide substantial advantageous in certain cases.

Thus, for example, if one of the three auxiliary hydrostatic bearings simultaneously under load is considerably less loaded than the other two, it is possible as illustrated in FIG. 11 to dispose between the two

grooves

319, 320 of the main bearings of plate 314 an element 335 producing a loss of pressure preventing the less loaded auxiliary bearing (in this case bearing 326 if groove 320 is the high-pressure one) from lifting abnormally the swash plate 1 of the pump. This element 335 is connected to the

auxiliary bearings

325 and 326 through

similar elements

333 and 334, and also through

valves

338, 339 to

grooves

319, 320.

Finally, it will be seen that if groove 320 is the highpressure one the pressure prevailing in

auxiliary bearings

325, 326 and 328 respectively may be varied by causing a variation, for the same compensation rate, in the diameters of the edges of groove 320 and in the diameters of the sealing surface.

In fact, the centre of gravity of the forces applied to the swash plate by the pistons is subordinate only to the diameter of the imaginary cylinder containing the axes of the multiple cylinders of the barrel, and also of the cylinder containing the centres of the ball-joints of the swash plate; on the other hand, the position of the groove reaction depends on the groove diameters and also on the diameters of the sealing surfaces. It will be seen that by varying these diameters it is possible to shift the point of application of the reaction of the hydrostatic bearings and thus produce a variation in the pressure prevailing in the various bearings (the sum of the pressures remaining on the other hand constant).

What is claimed as new is:

l. A hydrostatic bearing structure for the swash plate of a high-pressure cylinder-barrel hydraulic machine, which comprises a fluid distributor plate formed with ports communicating with the barrel cylinders through distribution orifices at the end of said cylinders, and a bearing plate, the bearing structure being formed on only one front face of the bearing plate and consisting of a main bearing including on the one hand a first set of two arcuate shallow grooves formed in said one face registering directly with a homologue rear face of the swash plate, said grooves being surrounded by sealing surfaces and on the other hand a second set of arcuated grooves formed in said face of said swash plate, said last-named grooves being shorter than those of the first set and equal in number to the distribution orifices provided at the end of the pump barrel cylinders, said hydrostatic bearing further comprising at least three elementary auxiliary hydrostatic bearings consisting of cavities formed in said front face of the hydrostatic bearing plate and disposed concentrically to the grooves of the first set of said main bearing and on a greater radius than the radius of said grooves of the main bearing, the rear face of said bearing plate comprising circular recesses concentric to the axis of said plate, said bearing plate having further ducts in its rear face which are connected to said recesses of the rear face, a first set of orifices opening through the plate respectively into the grooves of the front face and into said ducts, and a second set of connecting orifices interconnecting the recesses on the rear face to said cavities of elementary auxiliary hydrostatic bearings in the front face.

2. Hydrostatic bearing according to claim 1, wherein at least one of said auxiliary hydrostatic bearings is adapted to be supplied with fluid through two separate elements producing losses of pressure and connected to the grooves of said main hydrostatic bearing, respectively.

3. Hydrostatic bearing according to claim 2, wherein said bearing plate comprises checkvalves interposed between the elements producing a loss of pressure and connected to said auxiliary hydrostatic bearings, and the grooves of the first set of said main hydrostatic bearing, said valves being so disposed as to isolate the low-pressure groove.

4. Hydrostatic bearing according to claim 1, wherein each groove of said first set formed in the front face of said bearing plate delivers fluid under pressure to a set of at least three auxiliary elementary hydrostatic bearings through elements producing losses of pressure whichever the groove of said first set is under pressure, the arrangement of said three auxiliary elementary hydrostatic bearings being such that two bearings are common in said sets of auxiliary bearings and that the point of application of the axial force to be balanced, which is applied to said swash plate, be constantly located within the limits of the polygon containing the centers of gravity of the surfaces of aforesaid elementary auxiliary hydrostatic bearings.

Claims

1. A hydrostatic bearing structure for the swash plate of a high-pressure cylinder-barrel hydraulic machine, which comprises a fluid distributor plate formed with ports communicating with the barrel cylinders through distribution orifices at the end of said cylinders, and a bearing plate, the bearing structure being formed on only one front face of the bearing plate and consisting of a main bearing including on the one hand a first set of two arcuate shallow grooves formed in said one face registering directly with a homologue rear face of the swash plate, said grooves being surrounded by sealing surfaces and on the other hand a second set of arcuated grooves formed in said face of said swash plate, said last-named grooves being shorter than those of the first set and equal in number to the distribution orifices provided at the end of the pump barrel cylinders, said hydrostatic bearing further comprising at least three elementary auxiliary hydrostatic bearings consisting of cavities formed in said front face of the hydrostatic bearing plate and disposed concentrically to the grooves of the first set of said main bearing and on a greater radius than the radius of said grooves of the main bearing, the rear face of said bearing plate comprising circular recesses concentric to the axis of said plate, said bearing plate having further ducts in its rear face which are connected to said recesses of the rear face, a first set of orifices opening through the plate respectively into the grooves of the front face and into said ducts, and a second set of connecting orifices interconnecting the recesses on the rear face to said cavities of elementary auxiliary hydrostatic bearings in the front face.

3. Hydrostatic bearing according to claim 2, wherein said bearing plate comprises check-valves interposed between the elements producing a loss of pressure and connected to said auxiliary hydrostatic bearings, and the grooves of the first set of said main hydrostatic bearing, said valves being so disposed as to isolate the low-pressure groove.