Axial piston machine utilizing a bent-axis construction with slippers on the drive flange

Abstract

A hydrostatic axial piston machine (1) utilizing a bent-axis construction has a driveshaft (4) with a drive flange (3) rotatable around an axis of rotation (R.sub.t) inside a housing (2). A cylinder barrel (7) has pistons (10) fastened in an articulated manner to the drive flange (3). The drive flange (3) is supported on a housing-side slide face (101) by an axial bearing (100) in the form of a hydrostatically relieved sliding bearing (102) having a plurality of slippers (105). Each of the slippers (105) is mounted in an articulated manner in the drive flange (3) so that when the drive flange (3) rotates, a compensating force (F.sub.FR) acts on the slipper (105) which is in the opposite direction to the centrifugal force (F.sub.F) acting on the slipper (105). The point of application (AP) of the compensating force (F.sub.FR) on the slipper (105) is selected so that there is no tipping moment on the slipper (105) or to compensate for some or all of any tipping moment that does occur.

Claims

1. A hydrostatic axial piston machine utilizing a bent-axis construction, comprising: a driveshaft with a drive flange rotatable around an axis of rotation inside a housing; a cylinder barrel located inside the housing and rotatable around an axis of rotation, wherein the cylinder barrel includes a plurality of piston bores; a longitudinally displaceable piston located in each piston bore, wherein the pistons are fastened in an articulated manner to the drive flange, and wherein the drive flange is supported on a housing-side slide face by an axial bearing comprising a hydrostatically relieved sliding bearing having a plurality of slippers, each of which is mounted in an articulated manner in the drive flange and includes a pressure pocket on an end surface facing the slide face wherein the pressure pocket is in communication with an associated displacement chamber of the axial piston machine, wherein each of the slippers is mounted in an articulated manner in the drive flange so that when the drive flange rotates, a compensating force acts on the slipper which is in an opposite direction to the centrifugal force acting on the slipper, wherein a point of application of the compensating force on the slipper is selected so as to reduce or eliminate a tipping moment on the slipper or to compensate for some or all of the tipping moment that does occur, and a spring device that presses the slipper toward the housing-side slide face.

2. The hydrostatic axial piston machine as recited in claim 1, wherein a point of application of the compensating force in an axial direction lies at a level of a center of gravity of the slipper.

3. The hydrostatic axial piston machine as recited in claim 1, wherein the slipper is mounted in an articulated manner in a recess of the drive flange, wherein the radial support point of the slipper in the recess of the drive flange corresponds to a point of application of the compensating force.

4. The hydrostatic axial piston machine as recited in claim 1, wherein a radial support point of the slipper in the recess of the drive flange lies in a plane that is oriented perpendicular to the axis of rotation of the drive flange and is located in an axial direction in a vicinity of a center of gravity of the slipper.

5. The hydrostatic axial piston machine as recited in claim 1, wherein the slipper is mounted in an articulated manner in a recess of the drive flange, wherein a radial support point of the slipper in the recess of the drive flange is at a distance in an axial direction from a point of application of the compensating force.

6. The hydrostatic axial piston machine as recited in claim 1, wherein the slipper is in an operative connection with a compensating body that compensates in whole or in part for a tipping moment on the slipper caused by centrifugal force.

7. The hydrostatic axial piston machine as recited in claim 6, wherein the compensating body generates the compensating force that acts on the slipper and is in an opposite direction to the centrifugal force on the slipper, wherein a point of application of a compensating force generated by the compensating body and acting on the slipper lies in a vicinity of a center of gravity of the slipper.

8. The hydrostatic axial piston machine as recited in claim 6, wherein a radial support point of the slipper in the recess of the drive flange is kept at a distance in an axial direction of the center of gravity of the slipper by a first lever arm.

9. The hydrostatic axial piston machine as recited in claim 6, wherein the compensating body is mounted in an articulated manner on the drive flange by an articulated connection and is in an operative connection with the slipper in an axial direction in a vicinity of a center of gravity of the slipper, wherein the compensating force is generated by centrifugal force acting on the compensating body.

10. The hydrostatic axial piston machine as recited in claim 9, wherein the articulated connection of the compensating body on the drive flange is located in an axial direction between a center of gravity of the slipper and the center of gravity of the compensating body.

11. A hydrostatic axial piston machine as recited in claim 9, wherein the articulated connection of the compensating body with the drive flange is kept at a distance from the center of gravity of the compensating body by a second lever arm, wherein a mass of the compensating body, of the first lever arm, and of the second lever arm, are configured so that the compensating force generated by the compensating body is of a same magnitude as the centrifugal force acting on the slipper.

12. The hydrostatic axial piston machine as recited in claim 9, wherein at least one recess is located in a vicinity of the articulated connection of the compensating body, and wherein a pressure chamber is in communication with the displacement chamber by the at least one recess.

13. The hydrostatic axial piston machine as recited in claim 6, wherein the compensating body is coaxial with the slipper and is located inside the radial dimensions of the slipper in the drive flange.

14. The hydrostatic axial piston machine as recited in claim 13, wherein the drive flange includes an additional recess, in which the compensating body is mounted in an articulated manner, wherein the additional recess is coaxial with the recess for the slipper.

15. The hydrostatic axial piston machine as recited in claim 14, wherein the additional recess is in an operative connection with the displacement chamber and the compensating body includes a connecting channel, by means of which the pressure pocket of the slipper is in communication with the displacement chamber.

16. The hydrostatic axial piston machine as recited in claim 1, wherein a pressure chamber is located between the drive flange and the slipper, wherein the pressure chamber is in communication with the displacement chamber.

17. The hydrostatic axial piston machine as recited in claim 16, wherein the slipper is sealed to the pressure chamber by a sealing device.

18. The hydrostatic axial piston machine as recited in claim 17, wherein the slipper includes a groove-shaped recess in which the sealing device is located.

19. The hydrostatic axial piston machine as recited in claim 1, wherein the drive flange is one piece with the driveshaft.

20. The hydrostatic axial piston machine as recited in claim 1, wherein the slipper is located with a rim diametric clearance in the recess of the drive flange.

21. The hydrostatic axial piston machine as recited in claim 20, wherein the slipper includes a wider-diameter portion in a vicinity of a radial support point.

22. The hydrostatic axial piston machine as recited in claim 21, wherein a radially outer area of the wider-diameter portion is a spherical surface area, the midpoint of which lies in a center of gravity of the slipper.

23. The hydrostatic axial piston machine as recited in claim 21, wherein a radially outer area of the wider-diameter portion is an annular area.

24. The hydrostatic axial piston machine as recited in claim 21, wherein a radially outer area of the wider-diameter portion is a cylindrical surface area, wherein there is a rim diametric clearance between the cylindrical surface area and the recess of the drive flange.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) Additional advantages and details of the invention are explained in greater detail below with reference to the exemplary embodiments illustrated in the accompanying schematic figures, in which like reference numbers identify like parts throughout.

(2) FIG. 1 is a longitudinal section through an axial piston machine of the invention employing a bent-axis construction;

(3) FIG. 2 is a longitudinal section through a second exemplary embodiment of an axial piston machine of the invention employing the bent-axis construction;

(4) FIG. 3 is a detail of FIGS. 1 and 2 on an enlarged scale;

(5) FIG. 4 is a detail of FIGS. 1 to 3 on an enlarged scale;

(6) FIG. 5 is a detail of FIG. 4 on an enlarged scale;

(7) FIG. 6 shows an additional exemplary embodiment of the invention in an illustration like the one in FIG. 5;

(8) FIG. 7 shows an additional exemplary embodiment of the invention in an illustration like the one in FIG. 5; and

(9) FIG. 8 shows an additional exemplary embodiment of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

(10) A hydrostatic axial piston machine 1 in the form of a band-axis machine is illustrated in FIGS. 1 and 2. The machine 1 has a housing 2 that includes a housing barrel 2a and a housing cover 2b fastened to the housing barrel 2a. A driveshaft 4 provided with a drive flange 3 is mounted in the housing 2 by bearing devices 5a, 5b so that it can rotate around an axis of rotation R.sub.t. In the illustrated exemplary embodiment, the drive flange 3 is formed in one piece with the driveshaft 4, so that the driveshaft 4 and the drive flange 3 can be manufactured as a single part.

(11) Located in the housing 2 axially next to the drive flange 3 is a cylinder barrel 7, which is installed so that it can rotate around an axis of rotation R.sub.Z and includes a plurality of piston bores 8, which in the illustrated exemplary embodiment are arranged concentrically around the axis of rotation R.sub.Z of the cylinder barrel 7. A longitudinally displaceable piston 10 is located in each piston bore 8.

(12) The axis of rotation R.sub.t of the driveshaft 4 intersects the axis of rotation R.sub.Z of the cylinder barrel 7 at the intersection point S.

(13) In the illustrated exemplary embodiment, the cylinder barrel 7 includes a central longitudinal recess 11 that is concentric to the axis of rotation R.sub.Z of the cylinder barrel 7 through which the driveshaft 4 extends. The driveshaft 4 extends longitudinally through the axial piston machine 1 and is mounted on both sides of the cylinder barrel 7 by bearing devices 5a, 5b. The driveshaft 4 is mounted with the drive flange side bearing device 5a in the housing barrel 2a and with the cylinder-barrel-side bearing device 5b in the housing cover 2b.

(14) The driveshaft 4 is equipped on the drive flange side end with torque transmission means 12, such as splines, for the introduction of a drive torque or for the tapping of an output torque. The opposite, cylinder-barrel-side end of the driveshaft 4 that extends through the axial piston machine 1 ends in the vicinity of the housing cover 2b. In the housing cover 2b, to hold the driveshaft 4 and the bearing device 5b, there is a boring 14 that is concentric to the axis of rotation R.sub.t of the driveshaft 4 and, in the illustrated exemplary embodiment, is a through hole.

(15) For control of the feed and discharge of hydraulic fluid in the displacement chambers V formed by the piston bores 8 and the pistons 10, the cylinder barrel 7 is in contact with a control surface 15, which is provided with kidney-shaped control bores that form an inlet port 16 and an outlet port of the axial piston machine 1. For connection of the displacement chambers V formed by the piston bores 8 and the pistons 10 with the control bores, the cylinder barrel 7 is provided with a control opening 18 at each piston bore 8.

(16) The axial piston machine 1 illustrated in FIGS. 1 and 2 is in the form of a constant displacement machine with a fixed displacement volume. On a constant displacement machine, the angle of inclination , and thus the pivoting angle of the axis of rotation R.sub.Z of the cylinder barrel 7, is fixed and constant with respect to the axis of rotation R.sub.t of the drive flange 3 and/or the driveshaft 4. The control surface 15 with which the cylinder barrel 7 is in contact is formed on the housing 2, in the illustrated exemplary embodiment on the housing cover 2b, or on a control disc located non-rotationally in the housing 2.

(17) The pistons 10 are each fastened to the drive flange 3 in an articulated manner. Between each piston 10 and the drive flange 3, there is a joint 20 in the form of a spherical joint. In the illustrated embodiment, the articulated connection is a ball joint, which is formed by a ball head 10a of the piston 10 and a spherical cap-shaped recess 3a formed in the drive flange 3 in which the piston 10 is fastened by the ball head 10a.

(18) The pistons 10 each have a collar section 10b with which the piston 10 is positioned in the piston bore 8. A piston rod 10c of the piston 10 connects the collar segment 10b with the ball head 10a.

(19) To make possible a compensating movement of the pistons 10 during rotation of the cylinder barrel 7, the collar segment 10b of the piston 10 is located in the piston bore 8 with at least some rim clearance. The collar segment 10b of the piston 10 can be spherical. To create a seal between the pistons 10 and the piston bores 8, sealing means 21, such as a piston ring, are located on the collar segment 10b of the piston 10.

(20) For mounting and centering of the cylinder barrel 7, a spherical guide 25 is located between the cylinder barrel 7 and the driveshaft 4 respectively. The spherical guide 25 includes a spherical segment 26 of the driveshaft 4 on which the cylinder barrel 7 is located with a hollow spherical segment 27 located in the vicinity of the central longitudinal bore 11. The midpoint of segments 26, 27 lies at the intersection point S of the axis of rotation R.sub.t of the driveshaft 4 and the axis of rotation R.sub.Z of the cylinder barrel 7.

(21) To achieve the drive of the cylinder barrel 7 during operation of the axial piston machine 1, a drive joint 30 is located between the driveshaft 4 and cylinder barrel 7 that couples the driveshaft 4 and the cylinder barrel 7 in the direction of rotation. The driver device is not illustrated in detail in FIG. 1 and can be any conventional device.

(22) In FIG. 2, where identical components are identified by the same reference numbers, a drive joint 30 as the drive device is located between the driveshaft 4 and the cylinder barrel 7. In the illustrated exemplary embodiment, the drive joint is a constant velocity joint utilizing a cone-beam construction and makes possible a rotationally synchronous drive of the cylinder barrel 7 with the driveshaft 4, so that the result is a smooth, synchronous rotation of the cylinder barrel 7 with the driveshaft 4.

(23) In the illustrated exemplary embodiment, the drive joint 30 is a constant velocity joint, such as a cone-beam half-roller joint 31.

(24) The cone-beam half-roller joint 31 is formed by a plurality of roller pairs 50, 51 which are located between the driveshaft 4 and a sleeve-shaped driver element 40 non-rotationally connected with the cylinder barrel 7. In this case, the driveshaft 4 also extends through the drive joint 30.

(25) Each of the plurality of roller pairs 50, 51 of the cone-beam half-roller joint 31 includes two (a pair) of semi-cylindrical half-rollers 50a, 50b, 51a, 51b. The semi-cylindrical half-rollers 50a, 50b, 51 a, 51b, are each formed by a cylindrical body flattened essentially to an axis of rotation RR.sub.t, RR.sub.Z. On the flattened sides, the half-rollers arranged in pairs 50a, 50b, 51a, 51b each have plane slide faces GF at which the two half-rollers 50a, 50b, 51a, 51b of a roller pair 50, 51 are in contact with each other forming a planar contact.

(26) The half-rollers 50a, 50b, 51a, 51b are located in the radial direction inside the reference circle of the pistons 10 and at a distance from the axes of rotation R.sub.t, R.sub.Z. Therefore, the drive joint 30 can be located in a space-saving manner inside the reference circle of the pistons 10 and the driveshaft 4 can be located radially inside the half-rollers of the cone-beam half-roller joint 31.

(27) Each roller pair 50, 51 has a cylinder-barrel-side half-roller 50a, 51a that corresponds to the cylinder barrel 7 and a driveshaft side half-roller 50b, 51b that corresponds to the driveshaft 4, and are in contact with each other on the flat slide faces GF.

(28) The cylinder-barrel-side half-rollers 50a, 51a of the corresponding roller pair 50, 51 are each held in a cylindrical, or at least partly cylindrical, cylinder-barrel-side receptacle 55a, and the driveshaft side half-rollers 50b, 51b of a roller pair 50, 51 are held in a respective cylindrical, or at least partly cylindrical, driveshaft side receptacle 55b, and are secured in the respective cylindrical receptacle 55a, 55b in the longitudinal direction of the corresponding axis of rotation.

(29) Each half-roller 50a, 51a, 50b, 51b is provided in the cylindrical segment with a collar 60 which is engaged in a groove 61 of the corresponding receptacle 55a, 55b.

(30) In FIG. 2, the driveshaft side half-roller 50b of the roller pair 50 is represented by darker lines and the cylinder-barrel-side half-roller 50a in contact with the half-roller 50b is represented in fine lines. The cylinder-barrel-side half-roller 51a of the roller pair 51 is represented in darker lines and the driveshaft side half-roller 51b in contact with the half-roller 51a is represented in fine lines. Of the half-rollers 50b and 51a, the flattened, plane slide surfaces GF that lie in the sectional plane of FIG. 2 are shown.

(31) On the cone-beam half-roller joint 31 as illustrated in FIG. 2, the axes of rotation RR.sub.t of the driveshaft side half-rollers 50b, 51b are inclined with respect to the axis of rotation R.sub.t of the driveshaft 4 by an angle of rotation . The axes of rotation RR.sub.t of the driveshaft side half-rollers 50b, 51b intersect the axis of rotation R.sub.t of the driveshaft 4 at the intersection point S.sub.t. The individual axes of rotation RR.sub.t of the plurality of driveshaft side half-rollers 50b, 51b therefore form a cone beam around the axis of rotation R.sub.t of the driveshaft 4 with the tip at the intersection point S.sub.t.

(32) Accordingly, the axes of rotation RR.sub.z of the cylinder-barrel-side half-rollers 50a, 51a are inclined by an angle of inclination with respect to the axis of rotation R.sub.z of the cylinder barrel 7. The axes of rotation RR.sub.z of the cylinder-barrel-side half-rollers 50a, 51a intersect the axis of rotation R.sub.z of the cylinder barrel 7 at the intersection point S. The individual axes of rotation of the plurality of cylinder-barrel-side half-rollers 50a, 51a therefore form a cone beam around the axis of rotation R.sub.z of the cylinder barrel 7 with the tip at the point of intersection S.

(33) The angles of inclination of the axes of rotation RR.sub.z of the cylinder-barrel-side half-rollers 50a, 51a with respect to the axis of rotation R.sub.z of the cylinder barrel 7 and the axes of rotation RR.sub.t of the driveshaft side half-rollers 50b, 51b with respect to the axis of rotation R.sub.t of the driveshaft 4 are numerically identical. The angles of inclination of the axes of rotation RR.sub.z, RR.sub.t of the half-rollers of the driveshaft 4 and cylinder barrel 7 to be coupled with each other are therefore identical. Consequently, on the corresponding roller pairs 50, 51, each of the axes of rotation RR.sub.t corresponding to the driveshaft 4 and the axes of rotation RR.sub.z corresponding to the cylinder barrel 7 of the two half-rollers that form a roller pair intersect in pairs in a plane E that corresponds to the line bisecting the angle between the axis of rotation R.sub.t of the driveshaft 4 and the axis of rotation R.sub.z of the cylinder barrel 7. The points of intersection SP lying in the plane E at which the axes of rotation RR.sub.t corresponding to the driveshaft 4 intersect in pairs with the axes of rotation RR.sub.z corresponding to the cylinder barrel 7 of the two half-rollers that form a roller pair are illustrated in FIG. 2. The plane E is inclined at one-half the angle of inclination of the pivoting angle /2 with reference to a plane E1 that is perpendicular to the axis of rotation R.sub.t of the driveshaft 4 and a plane E2 that is perpendicular to the axis of rotation R.sub.z of the cylinder barrel 7. The plane E runs through the point of intersection S of the axes of rotation R.sub.t, R.sub.z.

(34) The half-rollers 50a, 50b, 51a, 51b of the respective roller pairs 50, 51 are located in the vicinity of the points of intersection SP of the axes of rotation RR.sub.t, RR.sub.z, as a result of which, at the points of intersection SP of the two half-rollers of the respective roller pairs 50, 51, the transmission of force between the plane slide faces GF takes place to drive the cylinder barrel 7.

(35) As a result of the position of the points of intersection SP of the two half-rollers of the respective roller pairs 50, 51 in the plane E, the perpendicular and radial distances from the points of intersection SP to the axis of rotation R.sub.t of the driveshaft 4 and to the axis of rotation R.sub.z of the cylinder barrel 7 are numerically equal. On account of the equal lever arms formed by the radial distances of the points of intersection SP, the angular velocities of the driveshaft 4 and of the cylinder barrel 7 are equal, as a result of which the cone-beam half-roller joint 31 forms a constant velocity joint that makes possible a rotationally synchronous and uniform drive and rotation of the cylinder barrel 7.

(36) In the axial piston machine 1 illustrated in FIGS. 1 and 2, for the axial mounting of the drive flange 3 on a housing-side slide face 101 of the housing 2, an axial bearing 100 is provided that is in the form of a hydrostatically relieved (balanced) sliding bearing 102. The hydrostatically relieved sliding bearing 102 comprises a plurality of slippers 105, each of which is mounted in an articulated manner so that it can move longitudinally in the drive flange 3, and is provided on an end surface facing the slide face 101 with a pressure pocket 106, which is in communication with an associated displacement chamber V of the axial piston machine 1 for the supply of hydraulic fluid. A slipper 105 is preferably associated with each piston 10.

(37) The pressure pockets 106 in the slippers 105 are each in communication via a communication channel 107 in the drive flange 3 and a communicating channel 108 in the piston 10 with the respective displacement chamber V which is formed by the piston bore 8 and the piston 10 located in it. The housing-side slide face 101 can be created directly in the housing 2 oras in the illustrated exemplary embodiment, on a circular bearing washer 109 which is non-rotationally fastened to the housing 2.

(38) The function of the axial bearing 100 is to hydrostatically relieve (balance) the axial forces on the drive flange 3 that occur during operation of the axial piston machine 1. As illustrated in FIG. 3, the piston force F.sub.K present on the pressurized pistons 10, which acts in the longitudinal direction of the pistons 10, is decomposed at the center point M of the articulated connection 20 into an axial force F.sub.A, which is directed parallel to the axis of rotation R.sub.t of the driveshaft 4 and of the drive flange 103, and a transverse force F.sub.Q, which is oriented perpendicular to it and generates the torque. The axial force F.sub.A (and, thus, the axial force component of the piston force F.sub.K) is relieved by a hydrostatic relief force F.sub.E generated by the slipper 105. As a result of this hydrostatic relief of the axial force F.sub.A, the bearing devices 5a, 5b of the driveshaft 4 can be made smaller than in prior machines, so that lower mass inertia occurs in the bearing devices 5a, 5b and compact dimensions of the axial piston machine 1 can be achieved.

(39) The slippers 105 are each pressed by a spring device 110, such as a compression spring, toward the housing-side slide face 101 and are thus pressed against the housing-side slide face 101.

(40) The slippers 105 are each located so that they can move longitudinally in a recess 111 of the drive flange 103. In the illustrated exemplary embodiment, the recesses 111 are each formed by a receptacle boring oriented concentric to the axis of rotation R.sub.t of the driveshaft 4 and of the drive flange 103. Between the drive flange 3 and each slipper 105 there is a pressure chamber D, which is in communication via the connecting channels 107 and 108 with the displacement chamber V. Located in each slipper 105 is a respective connecting channel 112 that connects the pressure pocket 106 with the pressure chamber D and, therefore, with the associated displacement chamber V. The pressure chamber D and the pressure pocket 106 are designed so that an additional hydrostatic application force is active that presses the slipper 105 against the slide face 101.

(41) Each slipper 105 is sealed by a sealing device 115 from the pressure chamber D. The slipper 105 is provided with a groove-shaped recess 116 in which the sealing device 115, such as an O-ring, is located.

(42) At high speeds of rotation of the axial piston machine 1, as illustrated in FIG. 4 the mass m of the slipper 105 results in a centrifugal force F.sub.F directed radially outwardly that is applied to the center of gravity SP of the slipper 105.

(43) Support for the centrifugal force F.sub.F is provided by an opposite compensating force F.sub.FR directed radially inwardly on the drive flange 3 which, in the exemplary embodiment illustrated in FIGS. 1 to 4, lies in the vicinity of the recess 111.

(44) To prevent a tipping of the slippers 105 away from the housing-side slide face 101 as a result of a tipping moment generated by the centrifugal force F.sub.F, in the axial piston machine 1 these slippers 105 are each mounted in an articulated manner in the drive flange 103 so that the point of application AP of the compensating force F.sub.FR is located on the slipper 105 so that no tipping moment occurs on the slipper 105. The position of the force pair that is formed by the centrifugal force F.sub.F and the compensating force F.sub.FR acting in the opposite direction to each other is therefore selected according to the invention so that no tipping moment caused by centrifugal force occurs on the slipper 105.

(45) The radial support point A of the slipper 105 in the recess 111 of the drive flange 3 on which the compensating force F.sub.FR is applied is located in a plane EE that is oriented perpendicularly to the axis of rotation R.sub.t of the drive flange 3 and is located in the axial direction in the vicinity of the center of gravity SP of the slipper 105. The radial support point A therefore forms the point of application AP of the compensating force F.sub.FR. Consequently, the centrifugal force F.sub.F and the compensating force F.sub.FR in the opposite direction have lines of action that are aligned with each other.

(46) The force pair formed by the centrifugal force F.sub.F and the opposite compensating force F.sub.FR therefore consists of forces that are directly opposite to each other, so that the centrifugal force F.sub.F and the opposite compensating force F.sub.FR have no lever anus on the support point A of the slipper 105 in the recess 111 and, therefore, no tipping moment caused by centrifugal force occurs on the slippers 105.

(47) To achieve the articulated mounting of the longitudinally displaceable slipper 105 in the recess 111 of the drive flange 103, the slipper 105, as illustrated in FIG. 5, is located with a rim diametric clearance DS1 in the recess 111 of the drive flange 103 and a diametric widening in the area in which the support point A is located.

(48) FIGS. 5 to 7 illustrate on a larger scale the areas in FIGS. 1 to 4 in which the support point A and the plane EE are located. In the exemplary embodiment illustrated in FIGS. 1 to 5, the radial outer area of the wider-diameter portion is in the form of a spherical surface SF on the slipper 105 that is located inside the recess 111. The midpoint MP of the spherical surface SF lies in the center of gravity SP of the slipper 105. The spherical surface SF guarantees an articulated mounting of the slipper 105 in the recess 111 that guarantees an effective compensation for tipping forces exerted on the slipper 105.

(49) FIGS. 6 and 7 illustrate alternative embodiments that can be used with the axial piston machine 1.

(50) As illustrated in FIG. 6, the radially outer area of the wider-diameter portion of the slipper 105 in the vicinity of the plane EE (and thus in the vicinity of the support point A) is in the form of a cylindrical surface ZF, the generated surface of which is concentric with the longitudinal axis of the slipper 105. To prevent tipping of the slipper 105 in the recess 111, a rim diametric clearance DS2 is provided between the cylindrical surface ZF and the recess 111 of the drive flange 3. The rim diametric clearance DS2 is less than the rim diametric clearance DS1 in the other areas of the slipper 105.

(51) As illustrated in FIG. 7, the radial outer area of the wider-diameter portion is in the form of an annular area RF of the slipper 105 in the vicinity of the plane EE (and thus in the vicinity of the support point A). The annular area in the form of an annular area RF has a radius R, the foot of which is located on the plane EE and at a radial distance from the center of gravity SP of the slipper 105.

(52) FIG. 8 illustrates an additional embodiment of an axial piston machine 1 utilizing the bent-axis construction, in which identical components are identified by the same reference numbers.

(53) In the exemplary embodiment illustrated in FIG. 8, the slippers 105 are each mounted in the drive flange 103 in an articulated manner and can move longitudinally so that when the drive flange 103 is in rotation, a compensating force F.sub.FR acts on the slipper 105 which is directed opposite to the centrifugal force F.sub.F acting on the slipper 105. The point of application AP of the compensating force F.sub.FR on the slipper 105 is selected to provide total or partial compensation for a tipping moment on the slipper 105 caused by centrifugal force.

(54) Each slipper 105 is in an operative connection with an additional compensating body 200 that fully or partly compensates for a tipping moment on the slipper 105 caused by the centrifugal force F.sub.F.

(55) The compensating body 200 generates the compensating force F.sub.FR that acts on the slipper 105, and is in the opposite direction to the centrifugal force F.sub.F acting on the slipper 105. The point of application AP of the compensating force F.sub.FR generated by the compensating body 200 and acting on the slipper 105 lies in the center of gravity SP of the slipper 105.

(56) The radial support point A of the slipper 105 in the recess 111 of the drive flange 3 is kept at a distance in the axial direction from the center of gravity SP of the slipper 105 by a first lever arm c.

(57) The compensating body 200 is mounted on the drive flange 103 by the articulated joint 210 in an articulated manner and is in an operative connection with the slipper 105 in the center of gravity SP. The compensating force F.sub.FR is generated by the centrifugal force F.sub.F2 acting on the compensating body 200.

(58) In the illustrated exemplary embodiment, the compensating body 200 is coaxial with the slipper 105, is mounted in an articulated manner, and is longitudinally movable within the radial dimensions of the slipper 105 in the drive flange 3.

(59) The drive flange 3 is provided with an additional recess 211 in which the compensating body 200 is mounted in an articulated manner and so that it can move longitudinally. The additional recess 211 is coaxial with the recess 111 for the slipper 105 and has a smaller diameter than that of the recess 111.

(60) The additional recess 211 is in communication via the connecting channel 107 in the drive flange 3 and the connecting channel 108 in the piston 10 with the displacement chamber V. The compensating body 200 is provided with a connecting channel 212, by means of which the pressure pocket 106 of the slipper 105 is in communication with the displacement chamber V.

(61) In the illustrated exemplary embodiment, the compensating body 200 is connected with the slipper 105 by a ball joint 220, the midpoint MMP of which is located in the center of gravity SP of the slipper 105. The ball joint 220 in the illustrated exemplary embodiment is formed by a ball head on a journal-shaped segment of the compensating body 200 and a recess in the form of a spherical cap in the slipper 105.

(62) For articulated installation of the compensating body 200 in the recess 211, which can move longitudinally in the recess 211, the compensating body 200 is located with a rim diametric clearance DS3 in the recess 211, and the articulated joint 210 is formed by a wider-diameter portion of the compensating body 200. In the illustrated exemplary embodiment, the radially outer area of the compensating body 200 in the vicinity of the wider-diameter portion is an annular area analogous to FIG. 7. The radial outer surface of the compensating body 200 in the vicinity of the expanded diameter can alternatively be designed analogous to FIGS. 5 and 6. The articulated joint 210 forms a radial support point B, with which the compensating body 200 is supported in the recess 211. The articulated joint 210, and thus, the support point B of the compensating body 200 on the drive flange 3, is located in the axial direction between the center of gravity SP of the slipper 105 and the center of gravity SK of the compensating body 200. The center of gravity SK of the compensating body 200 is kept at a distance from the articulated connection 210 and, thus, from the support point B, by the lever arm a.

(63) In the exemplary embodiment illustrated in FIG. 8, the spring device 110 is located in the recess 211 and applies pressure to the compensating body 200, which is in an operative connection with the slipper 105. Alternatively the spring device 110 can be located in the recess 111 and can apply pressure to the slipper 105 directly.

(64) The pressure chamber D that applies pressure to the slipper 105 is located between the slipper 105, the recess 111, and the compensating body 200. To achieve communication of the compression chamber D with the displacement chamber V, in the vicinity of the articulated connection 210 of the compensating body 200 there is at least one recess 215. The pressure chamber D is therefore in communication via the recess 215 and the rim diametric clearance DS3 of the compensating body 200 with the connecting channel 107.

(65) To achieve articulated mounting of the slipper 105 in the recess 111, and thus, to make it possible to control the tipping of the slipper 105 in the recess 111, the slipper 105 in FIG. 8 is provided analogous to FIG. 7 with a cylindrical outer area, whereby tipping is controlled by a corresponding rim diametric clearance. Between the cylindrical outer area of the slipper 105 and the recess, there is a relatively short guide length, so that in connection with an appropriately dimensioned rim diametric clearance, the control of the tipping of the slipper 105 becomes possible. Alternatively, the slipper 105 can be mounted in the recess 111 of the drive flange analogous to FIGS. 5 and 6.

(66) In FIG. 8, without compensation measures, the centrifugal force F.sub.F would be supported at the support point A and with the lever arm c between the center of gravity SP of the slipper 105 on which the centrifugal force F.sub.F is applied and the support point A of the slipper 105 in the recess 111, a tipping moment of the slipper 105 caused by centrifugal force would occur, which would cause the slipper 105 to tip away from the housing-side slide face 101. Compensation for some or all of this tipping moment caused by centrifugal force can be provided by the additional compensating bodies 200. The additional compensating body 200 applies the compensating force F.sub.FR in the direction opposite to the centrifugal force F.sub.F in the center of gravity SP of the slipper 105.

(67) The compensating force F.sub.FR results from the centrifugal force F.sub.F2 directed radially outwardly of the compensating body 200, which originates from the mass m.sub.2 of the compensating body 200, and is applied at the center of gravity SK of the compensating body 200, in connection with the reversal of the direction of force radially inwardly by, the selection of the support point B.

(68) In the exemplary embodiment illustrated in FIG. 8, the mass m.sub.2 of the compensating body 200, of the first lever arm c, and of the second lever arm a, are designed so that the compensating force F.sub.FR generated by the compensating body 200 is essentially of the same magnitude as the centrifugal force F.sub.F acting on the slipper 105. Consequently, compensation for the tipping moment of the slipper 105 can be provided by means of the additional compensating body 200 and a tipping of the slipper 105 away from the housing-side slide face 101 at high rotational speeds can be prevented.

(69) The invention is not limited to the exemplary embodiments illustrated and/or described above.

(70) In FIGS. 1 to 7, as a result of the position of the support point A in the plane EE that runs through the center of gravity SP, there is no tipping moment on the slipper 105. It goes without saying that the plane EE in which the support point A is located can be at a slight distance in the axial direction from the center of gravity SP, so that there is only a partial compensation of the tipping moment. As a result of this position of the plane EE, a short lever arm in the axial direction occurs between the force pair formed by the centrifugal force F.sub.F and the compensating force F.sub.FR, which can be tolerated with a corresponding sizing of the force applied by the spring 110 and the hydrostatic relief.

(71) The selection of the hydrostatic relief by the slipper 105 can be made so that the hydrostatic relief force F.sub.E equals the axial force F.sub.A, so that exact compensation can be provided for the axial force F.sub.A. This design can be incorporated in an axial piston machine in the form of a constant displacement machine with a constant displacement volume.

(72) Alternatively, the hydrostatic relief force F.sub.E can be less than the axial force F.sub.A, so that the remaining differential of the axial force from these two forces is absorbed by the drive-flange-side bearing device 5a.

(73) Alternatively, the hydrostatic relief force F.sub.E can be greater than the axial force F.sub.A, so that the remaining differential of the axial force from these two forces is absorbed by the cylinder-barrel-side bearing device 5b.

(74) Instead of in the form of a constant displacement machine, the axial piston machine 1 can be constructed as a variable displacement machine with a variable displacement volume. In a variable displacement machine, the angle of inclination (and thus the pivoting angle of the axis of rotation R.sub.Z of the cylinder barrel 7) is variable with respect to the axis of rotation R.sub.t of the driveshaft 4 for variation of the displacement volume. The control surface 15 with which the cylinder barrel 7 is in contact is for this purpose located on a cradle body, which is located in the housing 2 so that it can pivot around a pivoting axis that lies in the point of intersection S of the axis of rotation R.sub.t of the driveshaft 4 and the axis of rotation R.sub.Z of the cylinder barrel 7 and is oriented perpendicular to the axes of rotation R.sub.t and R.sub.Z. Depending on the position of the cradle body, the angle of inclination (and thus the pivoting angle of the axis of rotation R.sub.Z of the cylinder barrel 7) varies with respect to the axis of rotation R.sub.t of the driveshaft 4. The cylinder barrel 7 can be pivoted into a null position in which the axis of rotation R.sub.Z of the cylinder barrel 7 is coaxial with the axis of rotation R.sub.t of the driveshaft 4. Starting from this null position, the cylinder barrel can be pivoted to one or both sides, so that the axial piston machine can be constructed in the form of a unilaterally pivotable or as a bilaterally pivotable variable displacement machine.

(75) In a variable displacement machine in which the displacement volume is varied by varying the pivoting angle , the axial force F.sub.A varies as a result of the splitting of the force in the articulated joint 20. In the event of a reduction of the displacement volume by a reduction of the pivoting angle , the axial force F.sub.A increases. The selection among the above-mentioned three cases for the design of the hydrostatic relief force F.sub.E can therefore be made as a function of the selection of the hydrostatic relief force F.sub.E in the range of the pivoting angle of a variable displacement machine.

(76) It goes without saying that the driver element 40 can be constructed in one piece with the cylinder barrel 7.

(77) Instead of the driveshaft 4 that extends through the cylinder barrel 7 and is supported on bearings on both sides in the housing, the driveshaft 4 provided with the drive flange 3 can be supported by two bearing devices and cantilevered in the housing 2.

(78) It will be readily appreciated by those skilled in the art that modifications may be made to the invention without departing from the concepts disclosed in the foregoing description. Accordingly, the particular embodiments described in detail here are illustrative only and are not limiting to the scope of the invention, which is to be given the full breadth of the appended claims and any and all equivalents thereof.