阅读说明：本技术 具有受控制的缸压紧部和借助于调节器来调节的控制挡板的轴向活塞机 (Axial piston machine with controlled cylinder pressure and control flap which is adjusted by means of an adjuster ) 是由 S·豪克 于 2021-03-17 设计创作，主要内容包括：本发明涉及具有能关于旋转轴线(30)旋转的缸筒(60)的轴向活塞机(10),缸筒具有沿着旋转轴线(30)的方向指向的端面(63),其中端面(63)能够沿着旋转轴线(30)的方向通过活塞-缸-单元(80)朝控制面(54)挤压,其中活塞-缸-单元(80)为此能够用控制压力来加载,其中控制面(54)能够抗扭转地来布置。按照本发明,控制压力(85)通过能调节的控制阀(90)来提供,其中控制阀(90)如此被连接到控制装置(92)上,使得其能够由控制装置(92)来调节,其中控制装置(92)实现了一种调节器(100),该调节器的调节参量(101)影响控制阀(90)的调节。(The invention relates to an axial piston machine (10) having a cylinder barrel (60) which can be rotated about a rotational axis (30), having an end face (63) which points in the direction of the rotational axis (30), wherein the end face (63) can be pressed in the direction of the rotational axis (30) by a piston-cylinder unit (80) against a control surface (54), wherein the piston-cylinder unit (80) can be acted upon for this purpose by a control pressure, wherein the control surface (54) can be arranged in a rotationally fixed manner. According to the invention, the control pressure (85) is provided by a controllable control valve (90), wherein the control valve (90) is connected to the control device (92) in such a way that it can be controlled by the control device (92), wherein the control device (92) implements a controller (100) whose control variable (101) influences the control of the control valve (90).)

1. Axial piston machine (10) having a cylinder barrel (60) which is rotatable about a rotation axis (30) and which has an end face (63) which is directed in the direction of the rotation axis (30), wherein the end face (63) can be pressed in the direction of the rotation axis (30) by a piston-cylinder unit (80) against a control surface (54), wherein the piston-cylinder unit (80) can be acted upon for this purpose by a control pressure (85), wherein the control surface (54) is arranged so as to be rotationally fixed,

characterized in that the control pressure (85) is provided by a controllable control valve (90), wherein the control valve (90) is connected to a control device (92) in such a way that it can be controlled by the control device (92), wherein the control device (92) implements a controller (100) whose control variable (101) influences the control of the control valve (90).

2. An axial piston machine according to claim 1,

wherein the control pressure (85) is connected to a first pressure sensor (111), wherein the first pressure sensor (111) is connected to the control device (92), wherein the actual variable (102) of the regulator (100) is a measured value of the first pressure sensor (111), wherein a setpoint variable (103) of the regulator (100) can be selected at least as a function of the high pressure (20) of the axial piston machine (10).

3. Axial piston machine according to one of the preceding claims,

at least one first combined characteristic curve (121) is stored in the control device (92), said first combined characteristic curve having at least the high voltage (20) as an input variable, wherein the first combined characteristic curve has a setpoint variable (103) of the regulator (100) as an output variable.

4. Axial piston machine according to one of the preceding claims,

wherein the first characteristic curve (121) has, as a further input variable, the rotational speed (n) of the cylinder barrel (60) and/or the angle of rotation (α) of the axial piston machine (10) and/or the viscosity of the pressure fluid.

5. Axial piston machine according to one of the preceding claims,

a fluid source (21) is provided, by means of which the control valve (90) is supplied with pressure fluid.

6. Axial piston machine according to one of the preceding claims,

wherein the control valve (90) comprises a single continuously adjustable control flap (91), wherein the opening cross section of the control flap (91) can be adjusted by means of the control device (92).

7. Axial piston machine according to one of the preceding claims,

wherein the control of the control valve (90) is determined from a control variable (101) and a pilot control variable (104) of the regulator (100), wherein the pilot control variable (104) is dependent at least on the high pressure (20) of the axial piston machine (10).

8. Axial piston machine according to one of claims 3 to 7,

at least one second characteristic curve (122) which is different from the first characteristic curve (121) is stored in the control device (92), wherein the second characteristic curve (122) has at least the high pressure (20) of the axial piston machine (10) as an input variable, and wherein the second characteristic curve has the pilot control variable (104) as an output variable.

9. An axial piston machine according to claim 8,

wherein the second characteristic curve (122) has, as a further input variable, the rotational speed (n) of the cylinder barrel (60) and/or the angle of rotation (α) of the axial piston machine (10) and/or the viscosity of the pressure fluid.

Technical Field

The present invention relates to an axial piston machine according to the preamble of claim 1.

Background

DE 3904782 a1 discloses an axial piston machine of swash plate construction. Between the drive shaft and the cylinder barrel, a piston-cylinder unit is arranged, with which the end face of the cylinder barrel can be pressed against a control surface. The control pressure acting on the piston-cylinder unit is drawn from a control chamber, in which a high pressure is acting (anliegen).

Disclosure of Invention

The advantage of the present axial piston machine is that the leakage occurring between the end face and the control face is minimal under all operating conditions. The axial piston machine can be operated at particularly high rotational speeds and with particularly high pressures.

According to claim 1, the control pressure is provided by a controllable control valve, wherein the control valve is connected to a control device in such a way that it can be controlled by the control device, wherein the control device implements a controller, the control variable of which influences the control of the control valve. The axial piston machine is preferably embodied in a swash plate configuration, wherein the cylinder-piston units are supported on a drive shaft. The axial piston machine can also be implemented in a skew-axis configuration. The control valve is preferably electrically adjustable. The control device preferably comprises a digital computer, wherein the regulator is implemented digitally. The regulator is preferably a continuous, linear regulator, which is calculated at most preferably in a time-discrete manner. The regulator is preferably a PI regulator.

Advantageous developments and improvements of the invention are specified in the dependent claims.

It can be provided that the control pressure is connected to a first pressure sensor, wherein the first pressure sensor is connected to a control device, wherein the actual variable of the actuator is a measured value of the first pressure sensor, wherein a setpoint variable of the actuator can be selected at least as a function of the high pressure of the axial piston machine. It is also conceivable to ascertain the actual variable by means of a sensor which directly measures the thickness of the hydrostatic lubricating film between the end face and the control surface. Finally, it is precisely this thickness that should have a defined value, so that, nevertheless, an essentially complete hydraulic relief between the end face and the control surface is produced with as little leakage as possible. However, such fluid film thickness sensors are extremely expensive. With the proposed first pressure sensor, which is much less costly, a similar result can be obtained if the assigned nominal value is selected as suggested.

"completely hydraulically relieved" is to be understood as meaning a state in which there is no mechanical contact between the end faces and the control surfaces, the pressure there being transmitted hydrostatically via the pressure fluid used in the axial piston machine. The pressure fluid is preferably a liquid and at most preferably a hydraulic oil. The axial piston machine preferably has a first and a second working connection, wherein the rotation of the cylinder barrel is accompanied by a fluid flow between the first and the second working connection. The axial piston machine can have a 4-quadrant capability, i.e. it can be operated not only with two opposite directions of rotation of the cylinder barrel but also with respectively two opposite through-flow directions in each direction of rotation, irrespective of which pressures are loaded on the first and second working connections. The high pressure is the higher of the pressure on the first working connection and the pressure on the second working connection. The high pressure is preferably connected to a second pressure sensor. The second pressure sensor is preferably connected to the control device.

It can be provided that at least one first combination characteristic curve is stored in the control device, which first combination characteristic curve has at least the high voltage as an input variable, wherein the first combination characteristic curve has a setpoint variable of the regulator as an output variable. The first and/or the second characteristic combination curve to be mentioned below can be implemented in the form of a table of values. The first and/or second characteristic curve can also be realized by means of a mathematical interpolation formula. A hybrid form of these two implementations can also be considered. It is conceivable that the first and the second characteristic curve, which is still to be explained, are experimentally determined by using the fluid film thickness sensor mentioned above, in such a way that the desired ratio is produced. It is likewise conceivable for the first and/or second combined characteristic curve to be ascertained by computer simulation.

It can be provided that the first characteristic curve has, as further input variables, the rotational speed of the cylinder barrel and/or the angle of rotation of the axial piston machine and/or the viscosity of the pressure fluid. The axial piston machine preferably has an adjustable displacement volume. In the case of a swash plate machine, the pivot angle describes the necessary deflection of a pivoting cradle (Schwenkwiege). In the case of a skew-axis machine, the swivel angle describes the angle between the axis of rotation of the drive shaft and the axis of rotation of the cylinder barrel. The viscosity of the pressure fluid is preferably ascertained using a temperature sensor, the viscosity of which is ascertained from the corresponding measured values and the temperature-viscosity characteristic curve of the pressure fluid.

A fluid source can be provided through which the control valve is supplied with pressurized fluid. The fluid source is preferably formed by the first and/or second working connection. The first and second working connections can be connected on the input side to a directional valve, wherein the control valve can be connected on the output side to the directional valve. The fluid source can also be formed by a separate control oil pump.

It can be provided that the control valve comprises a single continuously adjustable control flap, wherein the opening cross section of the control flap can be adjusted by means of the control device. Preferably, there is a time-dependent correlation between the respective actuating signal at the control valve and the opening cross section of the control flap. In particular, the association mentioned is not defined by the control loop of the further subordinate. For example, the regulating current on the control valve is more or less proportional to the opening cross section of the control flap. The control valve is preferably an 2/2 proportional directional valve, which is preferably constructed at most as a directly controlled structure. It is also conceivable to use 3/2 proportional directional valves which are additionally connected to a fluid absorption source, such as a substantially pressure-free interior space of a housing or a tank. It is also conceivable to use a pressure reduction valve as the control valve to hydraulically adjust the control pressure. With the previously proposed very simple control valve, however, the improvements of the regulating method explained below can be realized, which significantly improve the stability of the overall system without increasing the hardware complexity as a result.

Provision can be made for the control of the control valve to be ascertained from a manipulated variable of the regulator and a pilot control variable, the pilot control variable being dependent at least on the high pressure of the axial piston machine. This can minimize the duration until the adjustment state is reached. The stability of the regulation is further improved. The adjustment of the control valve is preferably determined by the sum of the manipulated variable of the regulator and the pilot control variable.

It can be provided that at least one second characteristic curve, which is different from the first characteristic curve, is stored in the control device, wherein the second characteristic curve as an input variable at least has the high pressure of the axial piston machine, and wherein the second characteristic curve as an output variable has the pilot control variable. The first and second characteristic curves can be summarized as a three-dimensional characteristic curve, i.e. as a characteristic curve having a plurality of output variables. In contrast, the first and/or second characteristic curve preferably has only one single output variable.

It can be provided that the second characteristic map has, as further input variables, the rotational speed of the cylinder barrel and/or the angle of rotation of the axial piston machine and/or the viscosity of the pressure fluid. In this way, the pilot control variable can be flexibly adapted to different operating conditions of the axial piston machine.

It goes without saying that the features mentioned above and those yet to be explained below can be used not only in the respectively specified combination but also in other combinations or alone without leaving the scope of the invention.

Drawings

The invention is explained in detail below with the aid of the figures. Wherein:

fig. 1 shows a longitudinal section through an axial piston machine according to the invention;

fig. 2 shows a hydraulic circuit diagram of the axial piston machine according to fig. 1;

fig. 3 shows a section of fig. 1 in the region of a piston-cylinder unit;

figure 4 shows a longitudinal section of the cylinder;

fig. 5 shows a side view of the cylinder drum according to fig. 4;

FIG. 6 shows a perspective view of the control panel; and is

Fig. 7 shows a schematic diagram of the adjustment of the axial piston machine according to fig. 1.

Detailed Description

Fig. 1 shows a longitudinal section through an axial piston machine 10 according to the invention. The axial piston machine 10 comprises a housing 40, which is formed here by a first and a second housing part 41, 42. The first housing part 41 is embodied in the form of a pot, the open side of which is covered by a plate-like second housing part 42, so that a closed interior is created. In the housing 40, the drive shaft 33 is rotatably supported about the axis of rotation 30 by means of first and second rotary bearings 31, 32. The first and second rotary bearings 31, 32 are designed here as tapered roller bearings. The drive shaft 33 projects from the housing 40 with a journal 34, which is provided with splined shaft toothing (Keilwellenverzahnung) here.

Inside the housing 40, the drive shaft 33 is surrounded by a cylinder 60, which is connected in a rotationally fixed manner to the drive shaft 33 about the axis of rotation 30 by means of a splined shaft toothing (66 in fig. 4). The cylinder 60 has an end face 63 pointing in the direction of the axis of rotation 30, with which it bears against the control surface 54. The end face 63 is flat here and arranged perpendicular to the axis of rotation 30. However, it is also possible to use an end face 63 which is rotationally symmetrical about the axis of rotation 30 and which is, for example, concavely curved. The latter solution is common, for example, for axial piston machines of oblique-axis construction. Received in the cylinder 60 are a plurality of linearly movable working pistons 13, which are preferably arranged in a uniformly distributed manner around the axis of rotation 30. The axis of movement of the working piston 13 is arranged here at a slight inclination with respect to the axis of rotation 30.

The control surface 54 is arranged here on the separate control panel 50, wherein the control surface can also be arranged directly on the housing 40. The control plate 50 is connected to the housing 40 in a rotationally fixed manner about the axis of rotation 30, wherein it is supported on the housing 40 in the direction of the axis of rotation 30. During operation of the axial piston machine 10, a relative rotation thus occurs between the end face 63 and the control surface 54, a hydrostatic lubrication film being formed there. The advantage of the invention is that the leakage caused by such a lubricating film is minimal under all operating conditions, wherein at the same time minimal wear can be observed. The axial piston machine 10 can in particular be operated at high pressure and/or high rotational speed.

The present axial piston machine 10 is constructed in a swash plate configuration, wherein the present invention can also be applied to an axial piston machine of a swash plate configuration. The pivoting cradle (Schwenkwiege) 15 is capable of pivoting movement about a pivot axis arranged perpendicular to the axis of rotation 30. The swivel axis intersects the axis of rotation 30, wherein the swivel axis can also be arranged at a distance from the axis of rotation 30. The working pistons 13 are each supported on a flat surface of a pivoting cradle 15 by means of a slide 14, wherein the corresponding contact surfaces are preferably relieved hydraulically. The pivot angle of the pivoting cradle can be adjusted by means of a first and a second adjusting cylinder 16, 17, which are each designed as single-acting cylinders. The axial piston machine 10 can be adjusted to zero passage, i.e. the flow direction can be reversed only by pivoting the pivoting cradle 15, without the direction of rotation of the cylinder barrel changing. However, the invention can also be used in axial piston machines whose displacement volume (Verdraengungsvolumen) is adjustable in one direction only, starting from zero, or in axial piston machines whose displacement volume is constant.

It is also pointed out that the piston-cylinder unit 80 with which the cylinder tube 60 can be pressed in a defined manner hydraulically against the control surface 54, wherein reference is made to the explanation in relation to fig. 3 for further details.

Fig. 2 shows a hydraulic circuit diagram of the axial piston machine 10 according to fig. 1. The axial piston machine 10 has a first and a second working connection 11, 12, which are arranged here on a second housing part (42 in fig. 1). The axial piston machine 10 has a 4-quadrant capability in this case, i.e. it can be operated with two opposite directions of rotation of the drive shaft, wherein in both cases two opposite flow directions are possible. The pressure on the first and second working connections 11, 12 can be largely arbitrary here. The axial piston machine 10 can therefore alternately be operated as a pump and as a motor. The drive shaft of the axial piston machine 10 is in rotary drive connection with a motor 12, which can be embodied as an electric motor or as a combustion motor. The rotation of the drive shaft is accompanied by a fluid flow between the first and second working connections 11, 12, provided that the displacement volume adjusted is not equal to zero.

The adjusted pivot angle and, indirectly, the displacement volume can be measured here with a pivot angle sensor 113. The temperature and indirectly the viscosity of the pressure fluid can be measured here by a temperature sensor 115, which can be connected to the first and second working connections 11, 12. The rotational speed of the drive shaft or the cylinder can be measured here with a rotational speed sensor 114.

The first and second working connections 11, 12 are connected on the input side to a switching valve 22. The higher of the pressure at the first working port 11 and the pressure at the second working port 12 is thus applied to the output side of the directional control valve. This pressure is referred to within the scope of the present application as the high pressure 20. This high pressure is used here as a fluid source 21 for supplying pressure fluid to the control valve 90. It is to be noted here that a separate control oil pump can also be used as a fluid source. The high pressure 20 is measured here with a second pressure sensor 112.

The high pressure 20 is conducted via a control valve 90 to two transfer gaps 55 in the control plate 50, which are connected in parallel to the control valve 90. The control valve 90 comprises a continuously adjustable control flap 91, the opening cross section of which can be adjusted from zero up to a predetermined maximum value. The adjustment is preferably carried out electrically, at most preferably by means of an electromagnet 94, the adjusting force of which acts directly on the valve slide or valve cone of the adjusting valve 90. The control valve 90 is biased into the locked position by means of a spring 93. The control valve is designed here as an 2/2 directional proportional valve. The pressure loaded downstream of the control damper 91 is referred to as the control pressure 85. The control pressure 85 is measured here by means of a first pressure sensor 111. All the sensors 111, 112, 113, 114, 115 shown in fig. 2 are connected to the control device 92, wherein the electromagnet 94 of the control valve 90 is also connected thereto, so that the control valve 90 can be adjusted by the control device 92. The control device 92 effects the regulation described below with reference to fig. 7. The control means preferably comprises a programmable digital computer.

Fig. 3 shows a section of fig. 1 in the region of the piston-cylinder unit 80. The piston-cylinder unit 80 comprises a first and a second annular piston 81, 82, respectively, which are received between the drive shaft 33 and the cylinder 60. The corresponding contact gaps, which are preferably cylindrical with respect to the axis of rotation 30, are each substantially fluid-tightly sealed by a sealing ring 86, in particular an O-ring. The first annular piston 81 is supported on the drive shaft 33 in a form-fitting manner on a shoulder in the direction of the axis of rotation 30 facing away from the control surface (54 in fig. 1). The second annular piston 82 is supported on the cylinder tube 60 in the direction of the axis of rotation 30 toward a control surface (54 in fig. 1) by a safety ring 84. The cylinder 60, the drive shaft 33 and the first and second annular pistons 81, 82 together define a fluid chamber 84 which is substantially fluid-tightly closed except for the fluid passage 65. The cylinder 60 is minimally movable in the direction of the axis of rotation 30 relative to the drive shaft 33, so that the control pressure 85 present in the fluid chamber 84 presses the cylinder 60 against a control surface (54 in fig. 1). The corresponding reaction force is supported on a first rotary bearing (31 in fig. 1) by the drive shaft 33. The mounting positions of the respective tapered roller bearings are selected accordingly so that the two tapered roller bearings form an O-shaped arrangement.

Furthermore, the projections 83 on the first and second annular pistons 81, 82 are pointed out. These projections are first arranged so that the first and second annular pistons 81, 82 cannot cover the orifice 68 of the fluid passage. Furthermore, the respective surfaces, where the first and second annular pistons 81, 82 can come into contact, are implemented small relative to the total end face of the first or second annular pistons 81, 82. In each position of the annular pistons 81, 82, it is thereby ensured that the control pressure 85 causes a sufficiently large contact pressure of the cylinder barrel 60 on the control surface (54 in fig. 1).

Fig. 4 shows a longitudinal section of the cylinder 60. Cylinder bores 64 can be seen, wherein all cylinder bores 64 are embodied identically. The cylinder bore 64 has a cylindrical section 67 in which the associated working piston (13 in fig. 1) is received in a linearly movable and substantially fluid-tight manner. The cylindrical section can be formed by a separate sleeve made of a plain bearing material, such as bronze. The cylinder bore 64 opens with the first opening 61 into the end face 63. The cross-sectional area of the first orifice 61 can be the same as or slightly smaller than the cross-sectional area of the cylindrical section 67. In the latter case, the result of pressing the cylinder tube 60 against the control surface (54 in fig. 1) is brought about solely by the hydraulic pressure in the cylinder bore 64. Within the scope of the invention, this force is not too great as compared to the force that can be achieved with the piston-cylinder unit (80 in fig. 3) in order to be able to set the corresponding pressing force in a meaningful manner. The invention allows a particularly large first port 61, so that the axial piston machine does not form air pockets in the intake region even at high rotational speeds of the cylinder barrel 60.

Between two adjacent cylinder bores 64, in each case a fluid channel 65 is arranged, which runs obliquely to the rotational axis 30 in such a way that it opens on the one hand into the end face 63 with the second opening 62, wherein it opens on the opposite end into a fluid chamber (84 in fig. 3) with the third opening 68. All fluid channels 65 are identically constructed with respect to one another. The flow channel is embodied straight, wherein it is embodied here as a cylindrical stepped bore. The section forming the second aperture 62 has a smaller diameter than the other section. In this way, it should be possible to minimize the surface of the associated transfer gap (55 in fig. 6) so that the hydraulic forces acting there are smaller than the forces of the piston-cylinder unit.

The cylinder 60 is designed here as a one-piece construction. In particular in the region of the cylindrical section 67 and the end face 63, the cylinder barrel is, for example, case hardened, for example, carbonitrided. The cylinder 60 is made of steel or cast iron, for example.

Fig. 5 shows a side view of the cylinder drum 60 according to fig. 4. It can be seen that the first apertures 61 are of a shape other than circular. Its width in the radial direction is in particular implemented smaller than its length in the peripheral direction. The first opening 61 is arranged as close as possible to the spline shaft toothing 66, in order to exert a particularly low centrifugal force on the pressure fluid there. By this measure, the maximum permissible rotational speed of the cylinder 60 can be increased. The first aperture 61 defines a reference circle 70, the center of which is the axis of rotation 30. The radially outermost point of each first aperture 61 lies on this reference circle 70. The second apertures 62 are all arranged radially outside the reference circle 70. As a result, no direct fluid connection exists between the cylinder bore 64 and the control pressure (85 in fig. 2) in any position of the cylinder tube 60. In other words, the fluid connection is always made via an adjustable control flap (91 in fig. 2).

In this case, a total of nine cylinder bores 64 are provided, which are arranged uniformly distributed around the axis of rotation 30. Between all cylinder bores 64, in each case one fluid channel with a corresponding second port 62 is arranged. The spacing (71 in fig. 4) of all the second openings relative to the axis of rotation 30 is implemented identically. In principle, it is conceivable to provide fluid channels only between a part of the cylinder bores 64. The transfer gap (55 in fig. 6) must then extend over a larger section of the circumference, so that a larger hydraulic force is generated there, which is undesirable.

Fig. 6 shows a perspective view of the control panel 50, specifically from the control surface 54. The control plate 50 is basically implemented as a flat plate having a constant thickness. The rear side, which is not visible in fig. 6, is embodied completely flat. In contrast, the control surface 54 is provided with an inner annular recess 57, an annular recess 56 and an outer annular recess 58. The inner annular recess 57 and the annular groove 56 define a surface with which the control surface 54 rests against the end face of the cylinder barrel in the region of the first and second control recesses 51, 52. This surface defines the leakage occurring there. The surface is designed in such a way that, under all operating conditions, a hydrostatic lubricating film of such a thickness is formed that the end faces are essentially completely separated from the control surfaces 54. However, the mentioned lubricating film should not be thicker.

The first and second control recesses 51, 52 are each formed as a circularly curved elongated hole, wherein the respective center of the circle is defined by the axis of rotation 30. The control recess is arranged in line with the first bore (61 in fig. 6) of the cylinder. They can be provided with notches 53 at both ends in the peripheral direction, with which pressure peaks during operation of the axial piston machine are minimized. The first and second control clearances 51, 52 pass through the control plate 50 in the direction of the axis of rotation 30. The control recess is constructed as a mirror-symmetrical construction, since the present axial piston machine has a 4-quadrant capacity.

Two transfer recesses 55 are arranged between the annular groove 56 and the outer annular recess 58. These clearance areas pass through the control plate 50 in the direction of the axis of rotation 30. The care-of recesses are each formed as a circularly curved slot, the respective center of the circle being defined by the axis of rotation 30. The width of the passage gap in the radial direction is equal to the corresponding width of the second openings (72 in fig. 4), the passage gap being arranged in alignment with these second openings. The length of the transfer gap in the circumferential direction is selected such that in each rotational position of the cylinder, at least one second opening is located above the transfer gap 55. The length of the transfer gap is slightly greater than half the separation distance of the second opening, which is visible in fig. 5. The two care-of recesses 55 are arranged mirror-symmetrically to one another. It goes without saying that any arbitrary number of transfer gaps 55 can be used, provided their length in the circumferential direction is correspondingly designed.

The annular groove 56 is provided with a second through hole 132 at the bottom of the groove, which passes through the control plate 50 in the direction of the rotation axis 30. There, the leakage quantity which accumulates in the annular groove 56 is discharged into the interior of the housing for further introduction there into the tank via a leakage connection on the housing. A first through hole 131 in the center of the control plate 50 is penetrated by a drive shaft (reference numeral 33 in fig. 1).

Fig. 7 shows a schematic diagram of the adjustment of the axial piston machine according to fig. 1. The respective regulator 100 is preferably a continuous linear regulator, which is embodied, for example, as a PI regulator. The control deviation 105 is formed by the difference between the setpoint variable 103 and the actual variable 102. The actual variable is the control pressure measured by the first pressure sensor 111. The setpoint variable 102 is ascertained by means of a first combined characteristic 121. The input variable of the first characteristic map is the high pressure measured by the second pressure sensor 112, wherein the measured values of the rotational speed sensor 114 and/or the pivot angle sensor 113 and/or the temperature sensor 115 can be used as further input variables. The first characteristic curve 121 is designed such that the adjustment results in a substantially constant thickness of the hydrostatic lubricating film between the end face of the cylinder barrel and the control surface.

The control variable 101 of the controller acts on the control valve 90. The regulating valve 90 can comprise a single adjustable control flap 91, since unavoidable leakage would cause a continuous reduction of the control pressure (85 in fig. 2), if this effect is not overcome by a corresponding opening of the control flap 91.

It has been shown that a particularly stable system performance can be achieved by this particularly simple regulation.

The regulator 100 can be supplemented by a pre-control. A corresponding pilot control variable 104 is preferably added to the manipulated variable 101 of the regulator 100. The manipulated variable 104 can be ascertained by means of the second combined characteristic 122. The input variable of the second characteristic map 122 can be a measured value of the second pressure sensor 112 and/or of the rotational speed sensor 114 and/or of the pivot angle sensor 113 and/or of the temperature sensor 115. The second combined characteristic curve (Kennfeld) 112 can be ascertained, for example, by: the system is operated with constant measured values at the sensors mentioned without any preliminary control until the constant manipulated variable 101 reaches an average value. This manipulated variable is then used as a pilot control variable. By this measure, the time duration until the adjusted state is reached can be shortened.

List of reference numerals:

n number of revolutions of cylinder

Temperature of T pressure fluid

Angle of rotation alpha

10 axial piston machine

11 first working joint

12 second working joint

13 working piston

14 sliding seat

15-rotation cradle

16 first adjusting cylinder

17 second adjusting cylinder

20 high pressure

21 fluid source

22 change valve

23 Motor

30 axis of rotation

31 first rotary bearing

32 second swivel bearing

33 drive shaft

34 axle journal

40 casing

41 first housing part

42 second housing piece

50 control panel

51 first control gap

52 second control gap

53 gap

54 control surface

55 hand over the gap

56 annular groove

57 annular space inside

58 annular space outside

60 cylinder barrel

61 first orifice

62 second orifice

63 end face

64 Cylinder hole

65 fluid passage

66 spline shaft tooth part

67 cylindrical section of a cylinder bore

68 third orifice

70 reference circle

71 radial spacing of the second aperture relative to the axis of rotation

72 diameter of the second orifice measured in radial direction

80 piston-cylinder unit

81 first ring piston

82 second annular piston

83 projecting part

84 fluid chamber

85 control pressure

86 sealing ring

87 safety ring

90 control valve

91 control baffle

92 control device

93 spring

94 electromagnet

100 regulator

101 regulating variable

102 actual parameter

103 rated parameter

104 pre-control parameter

105 deviation of regulation

111 first pressure sensor

112 second pressure sensor

113 gyration angle sensor

114 revolution speed transducer

115 temperature sensor

121 first combined characteristic curve

122 second combined characteristic curve

131 first through hole

132 second perforation.