Abstract
An axial piston machine of swash plate design includes cylinder bores with respective longitudinal axes that are arranged at an acute angle with the rotational axis and that approach the rotational axis radially in the direction of their control plate-side ends. For each cylinder bore, a point of action of a resulting hydrostatic relieving force, which acts on the cylinder barrel, of the hydrostatic sliding bearing is spaced apart radially with regard to the rotational axis of the cylinder barrel further than a point of intersection of the longitudinal axis of the cylinder bore with the cylinder barrel-side bearing face of the hydrostatic sliding bearing.
Claims
1. A hydrostatic axial piston machine of swash plate design, comprising: a cylinder barrel mounted so as to rotate about a rotational axis, the cylinder barrel defining a multiplicity of cylinder bores, each of which is delimited in sections by one respective working piston and, during rotation of the cylinder barrel, is connected to high pressure or low pressure via a planar control plate having a first planar surface that bears against a first end side of the cylinder barrel via a first bearing face and a second bearing face of a hydrostatic sliding bearing; a swash plate mounted pivotably in a pivoting bearing and on which piston shoes arranged at a swash plate-side end region of each of the respective working pistons slide during the rotation of the cylinder barrel, wherein each cylinder bore has a central longitudinal axis arranged at an acute angle with respect to the rotational axis and which is radially closer to the rotational axis at the first end side of the cylinder barrel, wherein, during operation of the hydrostatic axial piston machine, the hydrostatic sliding bearing is configured so as to produce a hydrostatic relieving force acting on the cylinder barrel which, for each cylinder bore, acts at a first point of action, the first point of action being spaced apart radially from the rotational axis of the cylinder barrel by a first distance, which is greater than a second distance between the rotational axis and a first point of intersection of the central longitudinal axis of the cylinder bore with the first bearing face, which is on a cylinder-barrel side of the hydrostatic sliding bearing.
2. The axial piston machine according to claim 1, wherein the first distance is equal to or greater than a third distance between the rotational axis and a second point of action of a resulting relieving force, which, during operation of the axial piston machine, acts on the cylinder barrel, for each respective working piston, at a bearing face of the piston shoe of the respective working piston bearing on the swash plate.
3. The axial piston machine according to claim 1, wherein: the cylinder barrel defines a multiplicity of passage slots, each of which is assigned to a respective cylinder bore of the multiplicity of cylinder bores, extends from a distal end of the respective cylinder bore to the first bearing face of the hydrostatic sliding bearing, and has a cross-sectional area defined in a plane normal to a longitudinal axis of the passage slot that is less than a cross-sectional area of the respective cylinder bore defined in a plane normal to the central longitudinal axis of the respective cylinder bore; and the longitudinal axes of the passage slots are arranged substantially parallel to the rotational axis of the cylinder barrel.
4. The axial piston machine according to claim 3, wherein the longitudinal axis of the passage slot assigned to each working piston, at a control-plate side end of the passage slot, is spaced apart radially for each said working piston from the rotational axis of the cylinder barrel by a third distance, which is equal to or greater than a fourth distance between the rotational axis and a second point of action of a relieving force that acts on a swash plate end of the piston shoe of the respective working piston on a bearing face of the piston shoe in an axial direction of the respective working piston.
5. The axial piston machine according to claim 1, wherein: the hydrostatic sliding bearing is of annular configuration, and has an inner circular boundary and an outer circular boundary; the control plate is of substantially annular configuration and has at least two control openings which penetrate the control plate in an axial direction of the control plate and which have an inner and an outer circularly arcuate boundary line at an opening to the second bearing face; and passage openings open into the first bearing face so as to be aligned in the axial direction with the inner and outer boundary lines of the control openings.
6. The axial piston machine according to claim 5, wherein a radial spacing between the inner boundary line of the control opening and the inner boundary of the hydrostatic sliding bearing is substantially the same size as a radial spacing between the outer boundary line of the control opening and the outer boundary of the hydrostatic sliding bearing.
7. The axial piston machine according to claim 5, wherein a radial spacing between the inner boundary line of the control opening and the inner boundary of the hydrostatic sliding bearing is smaller than a radial spacing between the outer boundary line of the control opening and the outer boundary of the hydrostatic sliding bearing.
8. The axial piston machine according to claim 7, wherein: the inner boundary line of the control opening is assigned an inner arc length which is measured along the inner boundary of the hydrostatic sliding bearing and defines substantially the same arc angle as the inner boundary line; the outer boundary line of the control opening is assigned an outer arc length which is measured along the outer boundary of the hydrostatic sliding bearing and defines substantially the same arc angle as the outer boundary line; and a ratio between a radial spacing from the inner boundary line to the inner boundary and the inner arc length is substantially identical to a ratio between the radial spacing from the outer boundary line to the outer boundary and the outer arc length.
9. The axial piston machine according to claim 5, wherein the control openings are configured as substantially kidney-shaped control openings.
10. A hydrostatic axial piston machine of swash plate design, comprising: a cylinder barrel mounted so as to rotate about a rotational axis, the cylinder barrel defining a multiplicity of cylinder bores, each of which is delimited in sections by one respective working piston and, during rotation of the cylinder barrel, is connected to high pressure or low pressure via a planar control plate that bears against a first end side of the cylinder barrel via bearing faces of a hydrostatic sliding bearing; a swash plate mounted pivotably in a pivoting bearing and on which piston shoes arranged at a swash plate-side end region of each of the respective working pistons slide during the rotation of the cylinder barrel, wherein each cylinder bore has a longitudinal axis arranged at an acute angle with respect to the rotational axis and which is radially closer to the rotational axis at the first end side of the cylinder barrel, wherein the cylinder barrel defines a multiplicity of passage slots, each of which is assigned to a respective cylinder bore of the multiplicity of cylinder bores and extends from a distal end of the respective cylinder bore to a cylinder-barrel side bearing face of the hydrostatic sliding bearing, wherein longitudinal axes of the passage slots are arranged substantially parallel to the rotational axis of the cylinder barrel, and wherein the longitudinal axes of the passage slots are spaced apart radially from the rotational axis of the cylinder barrel further than points of intersection of the longitudinal axes of the cylinder bores with a plane which is arranged perpendicularly with respect to the rotational axis and contains cylinder bore-side ends of the passage slots.
11. The axial piston machine according to claim 10, wherein the longitudinal axes of the passage slots are spaced apart radially from the rotational axis of the cylinder barrel at the cylinder bore-side ends of the passage slots further than the points of intersection of the longitudinal axes of the cylinder bores with the plane which is arranged perpendicularly with respect to the rotational axis and contains the cylinder bore-side ends of the passage slots.
12. The axial piston machine according to claim 10, wherein the hydrostatic sliding bearing is configured so as to produce a hydrostatic relieving force acting on the cylinder barrel which, for each cylinder bore, acts at a first point of action, the first point of action being spaced apart radially from the rotational axis of the cylinder barrel by a first distance, which is greater than a second distance between the rotational axis and a point of intersection of the longitudinal axis of the cylinder bore with the first bearing face, which is on a cylinder-barrel side of the hydrostatic sliding bearing.
Description
BRIEF DESCRIPTION OF THE DRAWINGS
(1) In the following text, embodiments of the disclosure will be explained in greater detail using diagrammatic drawings, in which:
(2) FIG. 1 shows an axial section of an axial piston machine in the configuration, in which the bearing faces of the cylinder barrel and control plate are of planar design, and the cylinder bores and the working pistons are arranged obliquely with respect to the rotational axis of the drive shaft, in accordance with the prior art,
(3) FIG. 2 shows an axial section of a cylinder barrel which bears against a spherical control plate, with cylinder bores and working pistons which are configured obliquely with respect to the rotational axis, and the relieving forces on the hydrostatic sliding bearing and on the piston shoe in accordance with the prior art,
(4) FIG. 3 shows an axial section of a cylinder barrel which bears against a planar control plate, with cylinder bores and working pistons which are configured obliquely with respect to the rotational axis, and the relieving forces on the hydrostatic sliding bearing and on the piston shoe in an axial piston machine according to FIG. 1,
(5) FIG. 4 shows a section (analogous to FIG. 3) of a cylinder barrel according to a first embodiment of the disclosure, which cylinder barrel bears against a planar control plate, and the relieving forces on the hydrostatic sliding bearing and on the piston shoe,
(6) FIG. 5 shows an axial section of a cylinder barrel according to a second embodiment of the disclosure, which cylinder barrel bears against a planar control plate, and the relieving forces on the hydrostatic sliding bearing and on the piston shoe,
(7) FIG. 6 shows an axial section of a cylinder barrel according to a third embodiment of the disclosure, which cylinder barrel bears against a planar control plate, and the relieving forces on the hydrostatic sliding bearing and on the piston shoe,
(8) FIG. 7 shows a view in the axial direction of a detail of a planar control plate with a control opening according to a known embodiment of a control plate, and
(9) FIG. 8 shows a view in the axial direction of a detail of a planar control plate with a control opening according to a further embodiment according to the disclosure.
DETAILED DESCRIPTION
(10) The axial piston machine 2 which is shown in FIG. 1 and is designed by way of example for the configuration with bearing faces of planar design of the cylinder barrel and control plate and cylinder bores and working pistons which are arranged obliquely with respect to the rotational axis of the drive shaft comprises a housing 4, in the interior 6 of which there are, following one another axially, a swash plate 8, a cylinder barrel 10 with its rotational axis 12 and a substantially planar control plate 34 which is also called a control disk, and, furthermore, a drive shaft 62 which penetrates them with its rotational axis 70. The cylinder barrel 10 is arranged coaxially with respect to the drive shaft 62 and is coupled fixedly to the latter so as to rotate with it via a rotary driving device 68, for instance in the form of a toothed coupling. The cylinder barrel 10 bears with its end side which faces away from the swash plate 8 against the control plate 34 via a circularly annular, hydrostatic sliding bearing 42. Distributed on the circumference, cylinder bores 14 which are open toward the swash plate 8 with guide sleeves (not designated) and longitudinal axes 16 which are configured so as to be inclined at an acute angle 1 with regard to the rotational axes 12, 70 are arranged in the cylinder barrel 10. At the control plate-side end of the cylinder barrel 10, passage openings 18 which are configured coaxially with respect to the cylinder bores 14 with longitudinal axes 20 are provided. Preferably cylindrical working pistons 22 are mounted axially displaceably in the guide sleeves of the cylinder bores 14. At its control plate-side end, each working piston 22 delimits a working space 26 in the cylinder barrel 10. At the foot-side end, each working piston 22 is configured in the shape of a spherical head and is connected in an articulated manner to a piston shoe 28 which is supported on the swash plate 8 and slides on the swash plate 8 during rotation of the cylinder barrel 10. To this end, on the swash plate side, each piston shoe 28 has a planar sliding face (not designated) which bears against the swash plate 8 such that it can slide on a lubricating film which is formed by pressure medium.
(11) The housing 2 is formed from a cup-shaped housing part 4a with a housing bottom 4b and a housing shell 4c and a housing cover 4d which bears against a free edge of the housing shell 4c and is screwed thereto by means of screws (not shown). A feed line 38 and a discharge line 40 for feeding and discharging pressure medium to/from the working spaces 26 in the cylinder bores 14 are formed in the housing cover 4d. At least two control openings 36 are formed in the control plate 34 in the form of kidney-shaped through holes 36, as is also shown in FIGS. 5 and 6. The control openings 36 form sections of the feed line 38 and discharge line 40 which are formed in the housing cover 4d.
(12) In FIGS. 1 to 5, the passage openings 18 have a smaller diameter than the cylinder bores 14 and extend from the distal ends of the cylinder bores 14 as far as their openings into the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 42. In FIGS. 1 to 3, the passage openings 18 are formed coaxially with the cylinder bores 14.
(13) With regard to the connection of the working spaces or the cylinder bores in the cylinder barrel to the control openings in the control plate in axial piston machines with cylinder bores which run obliquely with respect to the rotational axis of the cylinder barrel and control plate of planar configuration, it has been known up to now that the cylinder bores open directly into the cylinder barrel-side bearing face of the hydrostatic sliding bearing, such as in DE 40 35 748 A1 (FIG. 2) which was cited at the outset, or via passage openings which are configured coaxially with the cylinder bores, as shown, for instance, in GB 1 073 216 which was cited at the outset or in the appended FIG. 1. In both cases, the longitudinal axis 16 of the cylinder bore 14 extends through a point of intersection 72 with the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 42; said point of intersection 72 is to be considered to be a point of action 74 of the relieving force F.sub.E10, which acts on the cylinder barrel 10, of the hydrostatic sliding bearing 42, as is shown in FIG. 3.
(14) For improved comprehension and as a distinction with respect to the disclosure, FIG. 2 shows an axial section of a cylinder barrel 10 which bears against a spherical control plate 34s with cylinder bores 14 and working pistons 16 which are configured obliquely with respect to the rotational axis 12 of the cylinder barrel 12. Furthermore, FIG. 2 illustrates the relieving force F.sub.E10 which is generated at the hydrostatic sliding bearing between the cylinder barrel 10 and the control plate 34s and acts on the cylinder barrel 10 and the relieving force F.sub.E28, which is generated by way of the contact of the piston shoe 28 with the swash plate 8, for a swash plate (not shown in FIG. 2) which is arranged perpendicularly with respect to the rotational axis 12 in its basic position.
(15) The relieving force F.sub.E28 which is transmitted by the piston shoe 28 has a direction of action which is directed perpendicularly with respect to the swash plate 8, and is axial in the case of FIG. 2. According to a vector decomposition of the relieving force F.sub.E28, a relieving force F.sub.E22 acts on the working piston 22 in the axial direction of the working piston 22, that is to say in the longitudinal direction 16 of the cylinder bore 14. The acute angle 1, with which the longitudinal axis 16 of each cylinder bore 14 runs obliquely with respect to the rotational axis 12, is then at most approximately 5 or less in the known configurations. For small angles 1 of this type, the magnitude of the relieving force F.sub.E28 which is absorbed by the piston shoe 28 is substantially equal to the relieving force F.sub.E22 which acts axially on the working piston 22 (cos 1cos 5=0.996).
(16) The relieving force F.sub.E10 which acts on the cylinder barrel 10 from the spherical control plate 34s acts at the point of intersection 72 of the longitudinal axis 16 of the cylinder bore 14 in the direction perpendicularly with respect to the bearing faces of the spherical sliding bearing 34s and in the direction of the longitudinal axis 16 of the cylinder bores 14, that is to say in the opposite direction to the relieving force F.sub.E22 which is transmitted from the piston shoe 28 to the working piston 22.
(17) In order to comprehend the disclosure, FIG. 3 then shows an axial section of a cylinder barrel 10 which bears against a planar control plate 34 with cylinder bores 14 and working pistons 16 which are configured obliquely with respect to the rotational axis 12 of the cylinder barrel 12. Furthermore, FIG. 3 illustrates the relieving force F.sub.E10 which is generated at the hydrostatic sliding bearing 42 between the cylinder barrel 10 and the planar control plate 34 and acts on the cylinder barrel 10 and the relieving force F.sub.E28 which is generated by the contact of the piston shoe 28 with the swash plate 8 for a swash plate (not shown in FIG. 3) which is arranged perpendicularly with respect to the rotational axis 12 in its basic position.
(18) As has already been explained with regard to FIG. 2, the relieving force F.sub.E28 which is transmitted by the piston shoe 28 to the cylinder barrel 10 has a direction of action which is directed perpendicularly with respect to the swash plate 8 and is axial in the case of FIG. 3, with the component F.sub.E22 which acts on the working piston 22 in its axial direction. The relieving force F.sub.E10 which acts on the cylinder barrel 10 from the planar control plate 34 acts at the point of intersection 72 of the longitudinal axis 16 of the cylinder bore 14 in the direction perpendicularly with respect to the planar bearing faces 44 and 46 of the sliding bearing 42, that is to say in the direction parallel to the rotational axis 12 of the cylinder barrel 10, and in the opposite direction to the relieving force F.sub.E28 which is transmitted from the piston shoe 28 to the cylinder barrel 10. On account of the oblique position of the cylinder bore 14 and depending on the axial spacing along the rotational axis 12 between the point of action 78 of the relieving force F.sub.E28 and the point of action 74 of the relieving force F.sub.E10, there is a radial offset 76 with regard to the rotational axis 12 between said points of action 74, 78 which are shown in FIG. 3, which radial offset 76 corresponds to the spacing of the lines of action of the relieving force F.sub.E28 and the relieving force F.sub.E10. On account of said radial offset 76 of the relieving force F.sub.E28 from the piston shoe 28 and of the relieving force F.sub.E10 of the hydrostatic bearing 42, a tilting moment is produced which acts on the cylinder barrel 10. As has already been said, the object of the disclosure then consists in making said tilting moment smaller or avoiding it. According to the disclosure, this is achieved by virtue of the fact that, for the cylinder bore 14, the point of action 74 or the line of action of the resulting hydrostatic relieving force F.sub.E10, which acts on the cylinder barrel 10, of the hydrostatic sliding bearing 42 is arranged spaced apart radially with regard to the rotational axis 12 of the cylinder barrel 10 further than the point of intersection 72 of the longitudinal axis 16 of the cylinder bore 14 with the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 42. The further the point of action 74 is spaced apart radially from the rotational axis 12 further, the smaller the axial offset 76 and the tilting moment which is caused as a result on the cylinder barrel 10.
(19) According to a first embodiment which is shown in FIG. 4, an approach of displacing the point of action 74 (and the line of action) of the relieving force F.sub.E10 which acts on the cylinder barrel 10 radially further to the outside or so as to be spaced apart further from the rotational axis 12 consists in providing, at the distal end of each cylinder bore 14, a passage slot 18 which is of narrowed cross section in comparison with said cylinder bore 14 and has a longitudinal axis 20 which is arranged substantially parallel to the rotational axis 12 of the cylinder barrel 10. In the axial direction, the passage slot 18 bridges the spacing between the distal end of the cylinder bore 14 and the plane, in which the passage slot 18 opens at the point of action 74 of the relieving force F.sub.E10 into the cylinder barrel-side bearing face 46 of the hydrostatic sliding bearing 42, the radial spacing R36 between its longitudinal axis 20 and the rotational axis 12 of the cylinder barrel 10 remaining substantially the same. In contrast, the longitudinal axis 16 of the cylinder bore 14 approaches the rotational axis 12 in the radial direction in the region of the axial extent of the passage slot 18 and intersects the cylinder barrel-side bearing face 46 of the hydrostatic sliding bearing 42 at the point of intersection 72. The radial offset between said point of intersection 72 and the point of action 74 of the relieving force F.sub.E10 brings about the reduction in the axial offset 76 which is shown in FIG. 3 and the reduction which is desired according to the disclosure of the tilting moment on the cylinder barrel 10.
(20) In the exemplary embodiment according to FIG. 5, in which the pistons lie less obliquely with respect to the rotational axis 12 than in the exemplary embodiments according to FIGS. 2 to 4, it is ensured that the point of action 74 of the relieving force F.sub.E10 is spaced apart radially from the rotational axis 12 to the same extent as the point of action of the relieving force F.sub.E28 on the piston shoe 28. As a result, the tilting moment on the cylinder is approximately zero.
(21) This is achieved by virtue of the fact that, in a longitudinal axis 20 of each passage slot 18, which longitudinal axis 20 is still parallel to the rotational axis 12, the radial spacing of the passage slot 18 and the control opening 36 from the rotational axis 12 is enlarged in comparison with a position centrally with respect to the control plate-side end region of the cylinder bores 14. The longitudinal axes 20 of the passage slot 18 and the control opening 36 are at a smaller spacing from the rotational axis 12 than the line of action of the forces F.sub.E10 and F.sub.E28 d. Moreover, the length L.sub.i of the gap section which extends to the inside is smaller than the length L.sub.a of the gap section which extends to the outside of the hydrostatic sliding bearing 42. Both measures lead to the point of action of the force F.sub.E10 migrating to the outside further away from the rotational axis 12.
(22) In the exemplary embodiment according to FIG. 6, the longitudinal axis 20 of the passage slot 18 and the control opening 36 is situated so far away from the rotational axis 12, as in the exemplary embodiment according to FIG. 5, that the line of action of the force F.sub.E28 lies on the longitudinal axis 20. In addition, the length L.sub.i of the gap section which extends to the inside is smaller and the length L.sub.a of the gap section which extends to the outside of the hydrostatic sliding bearing 42 is larger than in the exemplary embodiment according to FIG. 5. This leads to the line of action of the relieving force F.sub.E10 being radially further away from the rotational axis than the line of action of the force F.sub.E28. The special advantage of this exemplary embodiment in comparison with the exemplary embodiment according to FIG. 5 lies in the fact that the rotational speed is increased, at which the cylinder barrel 10 tends to rise up from the control plate. The special advantage of the exemplary embodiment according to FIG. 5 in comparison with the exemplary embodiment according to FIG. 6 lies in the higher rotational speed, at which self-suction still occurs. This is as a result of the smaller spacing of the passage slots 18 and the control openings in the control plate 34 from the rotational axis 12.
(23) Here, the acute angle 1 of the obliquely running longitudinal axis 16 of the cylinder bore 14 with respect to the rotational axis 12 is approximately 2. In general, it can be up to 5 in size. The established range lies between 1 and 4.
(24) FIGS. 7 and 8 show views in the axial direction of a detail of a planar control plate 34 with its surface which faces the cylinder barrel 10 in the state in which it is installed into an axial piston machine. Said surface forms the control plate-side bearing face 46 of the hydrostatic sliding bearing 42 which is otherwise of annular configuration and, in addition to the control plate-side bearing face 46, comprises the cylinder barrel-side bearing face 44 (shown in FIGS. 3 to 6) and has an inner circular boundary 48 and an outer circular boundary 50. The inner and outer boundary 48 and 50 of the hydrostatic sliding bearing 42 are either configured by virtue of the fact that the control plate-side bearing face 46 is configured as an annular elevation 52 which projects from that side of the control plate 34 which faces the cylinder barrel 10, as shown in FIGS. 1, 3 and 4, or by virtue of the fact that the cylinder barrel-side bearing face 44 is configured as an annular elevation 54 which projects from that side of the cylinder barrel 10 which faces the control plate 34, as shown in FIGS. 5 and 6. In both cases, a step is formed on the inner circumference and on the outer circumference of the annular elevation, as shown in FIGS. 3 to 6. The control plate 34 is of substantially annular configuration and has at least two substantially kidney-shaped control openings 36, of which in each case only one is shown in FIGS. 7 and 8. Each control opening 36 penetrates the control plate 34 in the axial direction and has an inner and an outer circularly arcuate boundary line 36i and 36a at its opening into the control plate-side bearing face 46. The opening of each passage opening 18 of the cylinder barrel 10 into the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 42 is arranged so as to be aligned axially with the inner and outer boundary lines 36i and 36a of each control opening 36. That is to say, in other words, the radii, that is to say the spacings from the rotational axis 12 of the cylinder barrel 10, of the inner boundary line 36i and the outer boundary line 36a correspond for each control opening 36 to the radial spacings of the inner edges and the radial spacings of the outer edges of the openings of the passage openings 18 into the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 48, and the radial spacing between the inner boundary line 36i and the outer boundary line 36a correspond to the diameters of the passage openings 18 at their openings into the cylinder barrel-side bearing face 44 of the hydrostatic sliding bearing 48.
(25) In the embodiment (shown in FIG. 7) of the hydrostatic sliding bearing 42 which has been customary up to now, each control opening 36 is arranged with regard to its radial dimension in ratio to the radial dimension of the control plate-side bearing face 46 in such a way that a radial spacing L.sub.1 between the inner boundary line 36i of the control opening 36 and the inner boundary 48 of the hydrostatic sliding bearing 42 is substantially the same size as a radial spacing L.sub.1 between the outer boundary line 36a of the control opening 36 and the outer boundary 50 of the hydrostatic sliding bearing 42.
(26) For the embodiment which is shown in FIG. 7, the point of action 74 of the relieving force E.sub.E10 which acts on the cylinder barrel-side bearing face 44 of the cylinder barrel 10 is arranged radially approximately in the middle between the inner and the outer boundary line 36i and 36a of the control opening 36. That is to say, the point of action 74 is arranged on the pitch circle of the control openings 36 with the pitch circle radius R.sub.36 and radially approximately in the middle between the inner and outer circular boundary 48 and 50 of the hydrostatic sliding bearing 42 which are assigned the radii R.sub.48 and R.sub.50.
(27) If, in this configuration, a control opening 34 on the high pressure side of the axial piston machine 2 lies opposite an opening of one of the passage slots 18 in the cylinder barrel 10 during operation of said axial piston machine 2, a gap is formed between the cylinder barrel-side and the control plate-side bearing face 44 and 46 around the passage opening 36 on account of the high pressure which is generated by way of loading with an operating pressure medium. Said gap comprises a part section which extends radially to the inside from the inner boundary line 36i of the control opening 34 and a part section which extends radially to the outside from the outer boundary line 36a of the control opening 36. Said part sections have the same length in the radial direction and substantially the same height in the axial direction in the case of corresponding planarity of the bearing faces 44 and 46. However, said part sections have different widths in the circumferential direction of the circularly annular hydrostatic sliding bearing 42. As a consequence, during operation, the leakage flows Q.sub.i for the operating pressure medium through the gap section which extends radially to the inside is smaller than the leakage flows Q.sub.a through the gap section which is directed radially to the outside.
(28) As shown in FIG. 8, this embodiment of the hydrostatic sliding bearing 42 and, in particular, of the control plate 34 can be modified in such a way that, for each control opening 36, the radial spacing L.sub.i between the inner boundary line 36i and the inner boundary 48 of the hydrostatic sliding bearing 42 is smaller than a radial spacing L.sub.a between the outer boundary line 36a and the outer boundary 50 of the hydrostatic sliding bearing 42. In other words, this modification consists in that the radial position of each control opening 36 is arranged offset to the inside in the radial direction from the radial middle, as shown in FIG. 7, with regard to the inner and outer boundary 48 and 50 of the hydrostatic sliding bearing 42, as is shown in FIG. 8, for example. One effect of this modification is that the gap section which extends radially to the inside from the inner boundary line 36i between the bearing faces 44, 46 is given as smaller length L.sub.i, with the result that the leakage flows Q.sub.i through said gap section for the operating pressure medium becomes greater, and that the gap section which extends radially to the outside from the outer boundary line 36a is given a greater length L.sub.a, with the result that the leakage flow Q.sub.a through said gap section becomes smaller. As a result, firstly the leakage flows Q.sub.i and Q.sub.a through the two gap sections are equalized at least partially. Secondly, in this configuration, a point of action or a line of action of the relieving force F.sub.E10 which acts on the cylinder barrel-side bearing face 44 of the cylinder barrel 10 is arranged radially further to the outside than the middle in the radial direction between the inner and outer circular boundary 48 and 50 of the hydrostatic sliding bearing 42. This modification of the design of the hydrostatic sliding bearing 42 therefore has an influence on the tilting moment which acts on the cylinder barrel 10. This influence or this modification can be realized in addition or as an alternative to the influences of the designs of the passage slots 18 in the cylinder barrel 10 which are described with regard to FIGS. 4 to 6.
(29) As can be gathered from FIG. 8, the inner boundary line 36i of the control opening 36 is assigned an inner arc length 56 which is measured along the inner boundary 48 of the hydrostatic sliding bearing 42 and defines substantially the same arc angle as the inner boundary line 36i. Furthermore, the outer boundary line 36a of the control opening 36 is assigned an outer arc length 58 which is measured along the outer boundary 50 of the hydrostatic sliding bearing 42 and defines substantially the same arc angle as the outer boundary line 36a. The radial arrangement of the control opening 36 or the boundary lines 48 and 50 is advantageously selected in such a way that the leakage flows Q.sub.i and Q.sub.a of operating pressure medium through the gap section which extends radially to the inside from the inner boundary line 36i and the gap section which extends radially to the outside from the outer boundary line 36i are substantially of the same size, because the overall leakage flow is then at a minimum. The following condition arises from the requirement that Q.sub.i is equal to Q.sub.a and under consideration of the fact that each of said leakage flows is proportional to the variable (gap width measured in the circumferential directiongap height measured in the axial direction/gap length measured in the radial direction): 56gap height/L.sub.i=58gap height/L.sub.a; and the following condition follows as a result of cancelling out the gap height: (56/L.sub.i)=(58/L.sub.a). In order to minimize the overall leakage flow, the ratio L.sub.i/56 between the radial spacing L.sub.i from the inner boundary line 36i into the inner boundary 48 and the inner arc length 56 is therefore to be selected to be substantially the same as the ratio L.sub.a/58 between the radial spacing L.sub.a from the outer boundary line 36a to the outer boundary 50 and the outer arc length 58. According to this condition, the radius R.sub.36 of the pitch circle of the control opening 36 is smaller in FIG. 8 than in FIG. 7 for otherwise identical (radial) dimensions R.sub.48 and R.sub.50 of the control plate-side bearing faces 46.
(30) The tilting moment on the cylinder barrel 10 is substantially avoided or eliminated if the point of action 74 of the resulting hydrostatic relieving force F.sub.E10, which acts on the cylinder barrel 10, of the hydrostatic sliding bearing 42, which corresponds to the radius R.sub.36 of the pitch circle of the control opening 36 is at the same spacing from the rotational axis 12 of the cylinder barrel as the point of action 78 of the relieving force F.sub.E28 which acts on the cylinder barrel 10 at the piston shoe 28, or, in other words, if the line of action of the relieving force F.sub.E10 of the hydrostatic sliding bearing 42 on the line of action of the relieving force F.sub.E28 coincide at the piston shoe 28.
(31) For an axial piston machine of swash plate design, in which the longitudinal axes of the cylinder bores are arranged at an acute angle with the rotational axis and the cylinder bores approach the rotational axis radially in the direction of their control plate-side ends, it is disclosed that, for each cylinder bore, a point of action of a resulting hydrostatic relieving force, which acts on the cylinder barrel, of the hydrostatic sliding bearing is spaced apart radially with regard to the rotational axis of the cylinder barrel further than a point of intersection of the longitudinal axis of the cylinder bore with the cylinder barrel-side bearing face of the hydrostatic sliding bearing.
(32) A hydrostatic axial piston machine according to the disclosure of swash plate design can particularly advantageously be used as a component of a hydraulic hybrid drive in a motor vehicle, in particular in a passenger motor vehicle. In vehicles, in particular, it comes down to a satisfactory degree of efficiency with low costs, a compact design and high rotational speeds which can be achieved.
