EP4172491B1 - Axial piston machine having a seal ring which is spherical in sections - Google Patents

Description

The invention relates to an axial piston machine in which pistons in cylinders perform a reciprocating movement and in which the pistons have a sealing ring receptacle for a sealing ring.

What all axial piston machines have in common is that cylinders are arranged in a circle around the cylinder drum axis in a cylinder drum parallel to the cylinder drum axis. Each cylinder accommodates a piston with a piston head, with the pistons being attached to a plate with an end opposite the piston head around a plate axis or being supported on this. If the cylinder drum axis and the plate axis intersect at an angle, the pistons are forced to perform a stroke movement when the cylinder drum and/or the plate rotate.

Hydraulic displacement machines, which include axial piston machines, work according to the displacement principle. They can therefore be operated both as pumps and as motors if the pressure medium flow is controlled accordingly. Pumps and motors usually have the same structural design. In the case of a motor, a pressure medium is supplied under pressure in approximately the first half of the cylinder and the pistons in question are pressed towards the plate by the pressure in the cylinders and/or a mechanical connection to the plate. If the angle of the cylinder drum axis to the plate axis is not zero, this creates a tangential force component which, depending on the design, causes either the cylinder drum or the plate to rotate and thus generates a downforce.

In the case of a pump, the cylinder drum axis or the plate is set in rotation, depending on the pump design. If the angle of the cylinder drum axis to the swash plate axis is not zero, the continuous change in distance between the piston and the swash plate forces the piston to perform an oscillating stroke movement in which expansion phases alternate with compression phases. During a downward movement, the expansion phase, the piston allows the respective cylinder to fill with pressure medium, which is pushed out from the piston bottom in a subsequent upward movement of the piston, the compression phase, thus generating a volume flow of the pressure medium.

From the conference paper " A NOVEL AXIAL PISTON PUMP/MOTOR PRINCIPLE WITH FLOATING PISTONS DESIGN AND TESTING", Liselott Ericson and Jonas Forsell, Proceedings of the 2018 Bath/ASME Symposium on Fluid Power and Motion Control, September 12-14, 2018, Bath, United Kingdom is a prototype of an axial piston machine with a piston plate mounted on a swash plate and with floating pistons. The seal between the piston chamber and the low-pressure housing interior of the hydrostatic machine is achieved by a sealing ring that is guided between the piston and the cylinder. This sealing ring is made of a relatively soft, deformable material. The conference paper reveals a material mix of polytetrafluoroethylene, PTFE and bronze. This sealing ring has a spherical sealing surface and its outer diameter is slightly larger than the inner diameter of the cylinder in order to achieve a sealing effect even under deformation. The curvature diameter of the spherical sealing surface is significantly smaller than the piston diameter. Due to the inclined piston plate, the sealing ring is moved along the inner wall of the cylinder at the speed of the piston and also in a circular path relative to the piston bore axis. The inclined position of the sealing ring in the cylindrical piston bore would create a gap. To counteract this gap, the diameter of the sealing ring is chosen to be approximately 1% larger than the diameter of the cylinder. In the piston presented in the conference report, the sealing ring is supported on the side facing away from the cylinder drum by a support ring, whereby the support ring has a smaller outer diameter than the sealing ring and is made of polyetheretherketone, PEEK, a harder material than the sealing ring.

However, tests have shown that during operation, particularly at high pressures, at a large angle between the cylinder drum axis and the piston plate axis, and at high rotation speeds, the sealing ring tends to extrude towards the pressure-free interior of the housing, i.e. into the gap between the piston and the piston bore. The inclination of the piston axis relative to the piston bore axis results in a kinematic axial offset of the piston and sealing ring. This axial offset increases the risk of extrusion of the sealing ring on the side that protrudes further beyond the piston.

For example, at a swivel angle of 8°, the sealing ring is stretched and compressed twice by approximately 1% of its diameter during a complete rotation, which leads to material fatigue in the long term. This can lead to seal failure in the long term. However, it has also been shown that Particularly at low speeds, the preload of the sealing ring causes an increased breakaway torque and stick-slip effects, which lead to uneven running of the machine. These effects are particularly negative in speed-controlled applications, where a stable system pressure cannot be achieved by imposing a certain speed (even very low speeds). These effects therefore make controllability considerably more difficult.

From the DE 199 06 690 A1 an axial piston machine with the features of the preamble of claim 1 is known. Further prior art is in the US 2 956 845 A and JP 2020 051276 A discussed.

The object of the invention is therefore to design an axial piston machine of the type mentioned at the outset in such a way that low-friction, low-pulsation, reliable operation at the sealing point to the piston bore is ensured in the entire operating range of the machine.

This object is achieved according to the invention with an axial piston machine according to independent claim 1. A manufacturing method for the axial piston machine according to the invention is the subject of claim 14. Further advantageous embodiments are the subject of the dependent claims.

According to the invention, the sealing ring is spherical at least in an area which seals the inner walls of the cylinder during the stroke movements, i.e. it is designed with a constant radius of curvature at least in this area, the radius of curvature of the sealing ring, which is spherical in some areas, essentially corresponding to half the diameter of the cylinder. In practice, the diameter of the sealing ring is chosen to be slightly smaller than the diameter of the cylinder in order to allow sufficient play between the inner wall of the cylinder and the sealing ring. This play is, for example, around 10 µm.

The partially spherical design of the sealing ring, in which the radius of curvature of the partially spherical sealing ring essentially corresponds to half the diameter of the cylinder, results in a sealing area that is ring-shaped, i.e. forms a closed circular line. A closed circular line causes far lower friction forces than a seal that corresponds to a surface-like seal due to unfavorable dimensions and / or geometry. When the spherical sealing ring rotates or tilts, the position of the circular sealing line on the surface of the at least partially spherical sealing ring, but since the diameter of the sealing circle line is constant due to the spherical shape and the inner diameter of the cylinder is also constant, there is always exactly the same clearance between the inner cylinder wall and the partially spherical sealing element, regardless of the position of the piston in the cylinder and regardless of the tilt angle of the piston, as long as the seal holder allows a compensating movement of the sealing ring transverse to the piston axis. This compensating movement is necessary because during the rotation of the cylinder drum the distance between the sealing ring and the cylinder drum axis varies cyclically due to the inclination of the piston axis to the cylinder drum axis.

By mounting the sealing ring in a way that allows lateral movement of the sealing ring across the longitudinal axis of the piston, the sealing ring can avoid radial and tangential forces that arise from the relative movement between the inner cylinder wall and the sealing ring, across the piston axis. This was already possible in the prior art, but because of the significantly smaller radius of curvature of the elastic sealing ring in relation to the inner cylinder diameter, the sealing line of the sealing ring would ideally only correspond to the inner cylinder diameter twice during one revolution if the inner cylinder diameter and the diameter of the sealing ring were chosen to be approximately the same. Between these two ideal positions, the sealing circle would be significantly smaller than the inner cylinder diameter and would therefore lead to leaks. For this reason, in the prior art, the diameter of the sealing ring was chosen to be slightly larger than the inner cylinder diameter. The oversized elastic sealing ring partially absorbs these force differences through reversible deformation of the elastic sealing ring, which leads to flat sealing surfaces in some places in the cylinder and at the same time to a gap between the sealing ring and the cylinder inner wall in other places. With a sealing ring diameter that is larger than the cylinder inner diameter, however, jamming of the sealing ring is unavoidable if a sealing ring made of a rigid material is selected.

With the design according to the invention, a sealing surface approximating a circular line is now created between the cylinder inner wall and the sealing ring in every position during rotation of the cylinder drum, with the clearance between the sealing ring and the cylinder wall being kept constant during the stroke movement. This makes it possible to select a non-deformable material for the sealing ring, so that extrusion of the sealing ring is avoided even under high pressures and/or high rotation speeds of the cylinder drum. At the same time or alternatively, the material of the The sealing ring should be made of a material that is particularly resistant to wear. This results in a longer service life for the sealing ring, meaning that the sealing ring needs to be replaced less often or not at all during the service life of the piston machine.

Since the diameter of a great circle in a ball is independent of the direction in which the ball is rotated, the piston element cannot jam in the cylinder during the stroke movement and simultaneous compensating movement because the diameter of the respective sealing circle remains unchanged compared to the diameter of the cylinder. The losses of the axial piston machine and the wear are thus reduced.

In one embodiment, the sealing ring is made of ceramic. Oxide ceramics, such as aluminum oxide Al2O3 or zirconium dioxide ZrO2, or alternatively non-oxide ceramics, such as silicon carbide SiC or silicon nitride Si3N4, are suitable for this purpose.

In a further embodiment, the sealing ring holder comprises a pin and the sealing ring has a central inner opening corresponding to the pin, wherein the inner opening diameter of the sealing ring is selected to be larger than the pin diameter. The difference between the pin diameter and the inner diameter of the sealing ring can thus be selected according to the required horizontal play, i.e. the play transverse to the piston's longitudinal axis.

In a further embodiment of the axial piston machine, the piston is designed in such a way that pressure equalization is possible between the piston interior and the interior of the sealing ring. Such pressure equalization can be achieved, for example, by fastening the sealing ring in the seal holder with vertical play, i.e. play in the direction of the piston's longitudinal axis, so that the pressure within the sealing ring holder can dynamically adapt to the pressure in the piston chamber via the gap of the play. In alternative embodiments, pressure equalization between the piston interior and the interior of the sealing ring is optionally achieved by one or more openings in the cover. In another embodiment, alternatively or additionally, pressure equalization bores are provided that extend from the top of the pin into the interior of the sealing ring. This makes it possible to use unequal geometries of the outer surface of the sealing ring and the inner surface of the sealing ring for targeted deformation of the sealing ring in order to increase the sealing effect of the sealing ring.

If the geometry of the outer surface of the sealing ring is different from the geometry of the inner surface of the sealing ring, the normal forces that act on the sealing ring from the pressure medium in the piston chamber are different from the normal forces in the sealing ring holder that act on the inside of the sealing ring. This can lead to deformation of the sealing ring, particularly at very high pressures of the pressure medium. This deformation, which is initially perceived as undesirable, is even increased in an embodiment according to the invention in which the central inner opening of the sealing ring has a circumferential bead-like recess.

This bead-like recess allows the sealing ring to expand further when the internal pressure in the piston chamber is high. It has been shown that even with solid cylinder drums, if a piston chamber of the cylinder drum is connected to the high-pressure side, this internal pressure force can lead to an expansion or deformation of the corresponding cylinder. Such one-sided expansion would lead to an increase in the gap between the inner cylinder wall and the sealing ring. It is therefore sensible to design the geometry of the sealing ring in such a way that it can also expand and thus the gap between the piston bore and the sealing ring remains almost constant. Since the working pressure in the piston chamber acts at the same level on the internal geometry of the sealing ring, the sealing ring will expand accordingly. The shape or wall thickness of the inner contour of the sealing ring can now be designed in such a way that the sealing ring expands exactly as much as the inner diameter of the piston bore of the cylinder drum. This keeps the gap constant. To a first approximation, this can be achieved by the bead-like recess in the sealing ring. At very high pressures, for example 350 bar and above, the cross-sectional shape of the sealing ring can be precisely determined by a geometry-optimized design of the ring geometry using appropriate deformation analyses with the finite element method.

As an alternative to the bead-like recess, the central inner opening of the sealing ring has a stepped profile. A first step has a first inner diameter and a second step has a second inner diameter, with the second inner diameter being selected to be larger than the first inner diameter. The first inner diameter corresponds to the inner diameter of a non-stepped sealing ring. The first inner diameter can therefore be adapted to the pin diameter of the sealing ring holder, so that the first inner diameter is adapted to the transmission of torque between the cylinder drum and the piston/piston plate. can be optimized via the contact surface of the sealing ring and the journal. The second inner diameter, however, since it is not involved in the torque transmission, can then be optimized for optimal expansion in order to adapt to the expanding piston bore at increasing, high operating pressures.

In one embodiment, the sealing ring is made of metal, for example iron, a steel alloy, or another metal alloy. Hardened steel with a surface hardness greater than 48 Rockwell hardness test, HRC, is particularly suitable for this, in particular tempering steel, for example 100Cr6 with a surface hardness of approx. 62 HRC, in particular case-hardened steel, for example 16MnCr5 with a surface hardness of approx. 60 HRC. In contrast to many ceramics, sealing rings made of metal have the advantage that, if the wall is thin enough, they expand due to the internal piston pressure and thus contribute to a better seal between the sealing ring and the inner wall of the piston chamber. However, this effect can also be achieved with ceramics that have a similar modulus of elasticity to steel. For example, in the case of ceramics made of zirconium oxide ZrO _2, rings made of zirconium oxide ZrO ₂ and steel expand largely in the same way.

In a further embodiment, the surface properties of a sealing ring made of metals in terms of surface hardness, friction coefficients and wear resistance are improved by downstream processes such as nitriding, nitrocarburizing or hard material coating.

A sealing ring made from a spherical disk does not necessarily have to be symmetrical in the axial direction. A geometry with an asymmetrical spherical disk can keep the pressure-dependent play between the ball ring and the cylinder wall small in order to achieve the lowest possible leakage. The ball ring is expanded in a targeted manner by this design and the applied pump pressure.

In a further embodiment, the sealing ring in the sealing ring holder is secured with a cover against movement along the longitudinal axis of the piston. The cover forms the piston bottom and at the same time limits the migration of the sealing ring in the direction of the cover during a downward movement of the piston, i.e. in the expansion phase, apart from a planned vertical play.

In a further embodiment, the cover is attached to the piston with a screw, or by clamping or pressing. These are Fastening methods that allow the cover to be removed in the event of repairs, thus simplifying the replacement of the sealing ring in the event of wear.

Mathematically speaking, the surface of the sectioned spherical sealing ring that comes into contact with the inner walls of the cylinder is a symmetrical spherical zone. A spherical zone is the curved outside of, for example, a spherical disk or a spherical ring. A spherical disk, or also called a spherical layer, is obtained as the middle part of a solid sphere when the solid sphere is cut into three parts by two parallel planes. If the parallel planes lie on different sides of the center of the sphere and are at the same distance from the center of the sphere, then it is a symmetrical spherical disk whose outer surface results in a symmetrical spherical zone. If the two parallel cutting planes are at different distances from the center of the sphere, an asymmetrical spherical disk can also be manufactured very easily in this way. Since the technical effort required to produce a sufficiently perfect spherical shape is relatively low, such a sealing ring can be manufactured with relatively little effort from solid balls with the appropriate diameter by removing ball segments on both sides of a selected great circle of the ball, for example by milling, which creates the desired symmetrical or asymmetrical spherical disk. Such solid balls are offered as standard components with the appropriate manufacturing precision for ball joints and swivel bearings, for example, and are therefore generally available and inexpensive.

A central opening with the desired diameter can then be created in a spherical disk obtained in this way, which allows the sealing ring to be accommodated in a pin. As provided in the alternative embodiments, the inside of the sealing disk can be milled out in order, for example, to adapt the wall thickness of the sealing ring to a desired shape.

In a further embodiment, the pistons are attached to the piston plate at one end. Because changes in the position of the piston in the cylinder are completely compensated by the play of the sealing ring and the partially spherical shape of the sealing ring, the piston does not require any joints or sliding shoes at the end of the piston facing away from the piston crown. Instead, it can be firmly connected to the piston plate.

In a further embodiment, the piston diameter tapers increasingly in the area between the sealing ring holder and one end. This enables a Tilting movement of the piston within the cylinder enables this to prevent contact between the piston and the inner cylinder walls during operation.

In a further embodiment, the piston has the shape of a truncated cone in the area between the sealing ring holder and one end.

In a further embodiment, the piston bore axes of the cylinders are distributed on a first circular line, piston bore pitch circle, around a cylinder drum axis, and the piston longitudinal axes are distributed on a second circular line, piston pitch circle, around a piston plate axis, wherein the diameter (D _K ) of the second circular line is selected to be larger than the diameter (D _z ) of the first circular line. The size differences between the first circular line and the second circular line can be compensated for by the inventive design of the sealing ring and pin, thus achieving a more compact design of the axial piston machine.

In a further embodiment, this design of the piston is used in a so-called floating piston machine.

In a further embodiment, the axial piston machine is designed as a swash plate machine.

Fig. 1

eine Schemazeichnung einer Axialkolbenmaschine mit den erfindungsgemäß ausgestalteten Kolben in einer neutralen Position

Fig. 2

eine Schemazeichnung einer Axialkolbenmaschine mit den erfindungsgemäß ausgestalteten Kolben in einer geschwenkten Position

Fig. 3

kegelstumpfförmiger Aufbau eines Kolbens

Fig. 4

zylinderförmiger Aufbau eines Kolbens

Fig. 5

einen kegelstumpfförmigen Kolben mit montiertem Dichtring

Fig. 6

Beispiel eines symmetrischen Dichtringes, das nicht um einen Teil der Erfindung, sondern um einen Stand der Technik handelt, der das Verständnis der Erfindung erleichtert.

Fig. 7

Beispiel eines asymmetrischen Dichtringes, das nicht um einen Teil der Erfindung, sondern um einen Stand der Technik handelt, der das Verständnis der Erfindung erleichtert.

Fig. 8

Ausführungsbeispiel eines symmetrischen Dichtringes mit innenseitigem Wulst

Fig. 9

Ausführungsbeispiel eines Dichtringes mit stufenförmiger Innenseite

Fig. 10

Beispiel eines Dichtrings mit kontinuierlicher Aufweitung seines Innendurchmessers in seinem oberen Bereich, das nicht um einen Teil der Erfindung, sondern um einen Stand der Technik handelt, der das Verständnis der Erfindung erleichtert.

Fig. 11

einen Kolben mit einem Dichtring mit wulstartiger Innenausnehmung und Druckausgleichsbohrung

Fig. 12

einen Kolben mit einem Dichtring mit gestuftem Innenprofil und Druckausgleichsbohrung

The invention will now be described and explained in more detail using embodiments shown in the drawings. They show:

Fig.1

a schematic drawing of an axial piston machine with the pistons designed according to the invention in a neutral position

Fig.2

a schematic drawing of an axial piston machine with the pistons designed according to the invention in a pivoted position

Fig.3

truncated cone-shaped structure of a piston

Fig.4

cylindrical structure of a piston

Fig.5

a truncated cone-shaped piston with mounted sealing ring

Fig.6

Example of a symmetrical sealing ring which is not part of the invention but is prior art which facilitates the understanding of the invention.

Fig.7

Example of an asymmetric sealing ring which is not part of the invention but is prior art which facilitates the understanding of the invention.

Fig.8

Example of a symmetrical sealing ring with inner bead

Fig.9

Example of a sealing ring with stepped inner side

Fig.10

Example of a sealing ring with a continuous widening of its inner diameter in its upper region, which is not part of the invention but is prior art which facilitates the understanding of the invention.

Fig. 11

a piston with a sealing ring with a bead-like inner recess and pressure compensation bore

Fig. 12

a piston with a sealing ring with stepped inner profile and pressure compensation bore

Figure 1 and Figure 2 show the schematic structure of a so-called floating piston machine, representative of the structure and function of axial piston machines. Figure 1 and Figure 2 show the same floating piston machine in different working conditions. The structure and function of a floating piston machine are well known to the expert, so that in Figure 1 and Figure 2 only the interaction of a piston 2 with a cylinder drum 7, a piston plate 8 and a swash plate 9 is described. The piston plate 8 is supported on the swash plate 9 and is rotatably mounted on it. Figure 1 shows the floating piston machine 1 in a neutral state, in which the swivel plate 9 and cylinder drum 7 are aligned parallel to each other, while Figure 2 the floating piston machine 1 in a state in which the swash plate 9 and the cylinder drum are not aligned parallel to each other.

In the exemplary embodiment, a plurality of cylinders 3 are circular and evenly distributed around a cylinder drum axis 70 of a cylinder drum 7. In the exemplary embodiment, the cylinders 3 are designed as piston bores 3 and will be referred to as such from now on. However, it is clear to the person skilled in the art that a cylinder 3 can also be manufactured in a manner other than through a piston bore. In order to avoid harmonic vibrations, an odd number of piston bores 3 is usually selected. On the top side 71 of the cylinder drum 7, each piston bore 3 has a connecting bore 33, via which a pressure medium can be supplied to or discharged from the so-called high-pressure side of the floating piston machine 1 to the piston bores 3.

The cylinder drum 7 is mounted in such a way that rotation around the cylinder drum axis 70 is permitted. For the transmission of torques, a shaft 72 is arranged on the cylinder drum 7, which in the case of the floating piston machine operating mode acts as a pump drive shaft and in the case of the floating piston machine operating mode acts as a The engine provides an output shaft. The distance R from a piston bore axis 30 to the cylinder drum axis 70 is 45 mm in the described embodiment, while the piston bores 3 each have an inner diameter D of 15 mm. In order to better illustrate the invention, the figures are not true to scale and show details in some cases greatly enlarged.

The pistons 2 are rotationally symmetrical. The axis of symmetry of the pistons 2 is also referred to below as the longitudinal axis 20 of the piston 2. Figure 3 shows the basic structure of a piston 2 with a piston head 21 at its upper end and a piston foot 22 at its lower end. The direction indication "upwards" in connection with a piston 2 refers to a movement of the piston 2 within the piston chamber 31 in the direction of the piston head 21, while the direction indication "downwards" refers to a movement of the piston 2 within the piston chamber 31 in the direction of the piston foot 22. The piston head 21 typically has a larger diameter than the piston foot 22. The piston 2 can therefore be designed according to the Figure 2 in its central region 24 have the shape of a truncated cone. It is important that the diameter of the piston head 21 is selected such that the piston head 21 does not come into contact with an inner wall 32 of the piston bore 3 at any time during operation of the piston machine. Insofar as this is taken into account, the piston 2 can also be designed in the shape of a cylinder in its central region 24, as is the case in Fig.4 is shown.

The piston plate 8 is designed as a circular disk through whose circular disk center a piston plate axis 80 extends perpendicular to the piston plate 8. The piston plate 8 is rotatably mounted so that the piston plate 8 can rotate about the piston plate axis 80. The swash plate 9 is also designed as a circular disk and through whose circular disk center a swash plate axis 90 extends perpendicular to the swash plate 9. In the neutral state of the floating piston machine 1, the piston plate axis 80 and the swash plate axis 90 are in line with the cylinder drum axis 70.

In the following, a plane that extends perpendicularly around the cylinder drum axis 70 is referred to as the cylinder drum plane 75 and a plane that extends perpendicularly to the piston plate axis is referred to as the piston plate plane 85. In the neutral state, the cylinder drum plane 75 and the piston plate plane 85 are aligned parallel to each other. When the cylinder drum 7 rotates, the distance between the bottom side 72 of the cylinder drum 7 and the top side 81 of the Piston plate 8 is constant. Due to the constant distance, the pistons 2 do not perform any lifting movement. This distance between the bottom 72 of the cylinder drum and the top 81 of the piston plate 8 is therefore referred to below as the neutral distance S0.

In this embodiment, the piston plate 8 is designed to be pivotable relative to the cylinder drum plane 85. When the swash plate 9 is pivoted, care must be taken to ensure that the cylinder drum axis 70 and the swash plate axis 90 intersect at a pivot point X at an angle α. Since the piston plate 8 slides on the swash plate 9 and thus the piston plate 8 and swash plate 9 always remain parallel to one another, a geometric law dictates that the angle α at which the cylinder drum plane 75 and the piston plate plane 85 intersect corresponds to the pivot angle α. The pivot angle α also corresponds to the angle at which the piston axes 20 are tilted relative to the cylinder bore axes 30. At a pivot angle α = 0°, the neutral position, the piston axes 20 are aligned parallel to the piston bore axes 30.

At a swivel angle α not equal to 0°, one half of the piston plate 8 is tilted away from the cylinder drum 7 and the other half of the piston plate is inclined towards the cylinder drum 7, so that during rotation the distance between the cylinder drum underside 72 and the piston plate top side 81 changes continuously. During one rotation, the piston plate 8 passes through a maximum distance S _max starting from the middle distance after a quarter full circle rotation; after another quarter full circle rotation, the top side 81 of the piston plate 8 returns to the middle distance; after another quarter full circle rotation, the top side 81 of the piston plate 8 passes through a minimum distance S _min to the underside of the cylinder drum 7 and after another quarter full circle rotation, the piston plate 8 returns to its starting point. In order to achieve these positions in the Figure 2 To clarify, these distances and the two pistons or piston chambers are shown for an even number n of piston bores.

Since the pistons 2 are firmly connected to the piston plate 8 with their piston base 22, the pistons 2 are forced to perform these up and down movements when the cylinder drum 7 and piston plate 8 rotate. During the upward movement, the piston chamber 31, which is sealed to the inside of the housing by the sealing ring 5, shrinks until the piston 2 reaches a top dead center TDC, where it changes its direction of stroke. The top dead center TDC of the piston 2 is identical to the position at which the Piston plate 8 has reached the minimum distance S _min . In the subsequent downward movement, the piston chamber enlarges until the piston 2 reaches a bottom dead center UT, where the downward stroke movement changes back into an upward stroke movement. The bottom dead center UT is identical to the position at which the top 81 of the piston plate 8 has reached a maximum distance S _max from the bottom 72 of the cylinder drum 7.

The piston base 22 is advantageously shaped as a cylinder because this allows the piston base 22 to be received in a through hole in the piston plate 8. Since the piston either widens out to form a truncated cone after the piston base 22 or forms a step to the larger cylindrical middle part 24, the piston 2 is supported on the piston plate top 81 in order to divert the forces acting on the piston head 21 in the piston chamber 31 into the piston plate 8.

If the middle part 24 is not widened compared to the piston foot 22, this support can alternatively be achieved by the receptacles for the piston foot 22 being designed as blind holes and the respective piston foot 22 being supported in the respective blind hole. The piston feet 22 are fixed against any kind of movement, for example by pressing them into the through hole or the blind hole. Alternatively, a connection can also be made in another form-fitting or force-fitting manner, for example by pressing in, shrinking, threading or welding.

Fig.5 shows a piston 4 with a sealing ring 5 mounted in a sealing ring holder 4. The sealing ring holder 4 in this case has a pin 23 centered on the piston head 21, which receives a central opening 51 of the sealing ring 5. The inner diameter d _i of the central opening 51 is selected to be significantly larger than the diameter dz of the pin 23. A movement of the sealing ring 5 in the direction of the longitudinal axis 20 of the piston 2 is limited by a cover 6, which is mounted on the pin 23.

Fig.6 shows a sealing ring 5 in its simplest manufacturing embodiment. The sealing ring 5 of the Figure 6 is a spherical disk in which the spherical disk has the same height h/2 upwards and downwards from an equatorial plane 58 of the sealing ring. The equatorial plane 58 contains the great circle on the outer surface 52 of the sealing ring, which is perpendicular to the sealing ring axis 50. With the same height h/2 of the sealing ring on both sides of the equatorial plane, this is a symmetrically designed sealing ring 5. From the radius of curvature r This results in the diameter d _a of the sealing ring, which is ideally slightly smaller than the piston diameter d.

We first consider the case where the piston plate plane 85 is aligned parallel to the cylinder drum plane 75 and the cylinder drum axis 70 coincides with the piston plate axis 80 and the swash plate axis 90, i.e. the neutral position. When the cylinder drum 7 and piston plate 8 rotate in the neutral position, the pistons 2 do not perform a stroke movement because no relative movements occur in the direction of the piston bore axes 30. This means that no vertical forces act on the sealing ring 5, i.e. forces parallel to the cylinder drum axis 70.

Let us now look at the Figure 2 the conditions when the swash plate 9 is inclined relative to the cylinder drum 7 by a swivel angle α <> 0°. A rigid piston head describes an elliptical path when the cylinder drum 7 rotates within the piston bore 3, whereby the vertices of the main axis of this elliptical path are passed through at the top dead center OT and bottom dead center UT. In the situation as shown in Figure 2 As shown, the piston 2, when it reaches the top dead center OT, would protrude beyond the part of the inner wall 32 of the piston bore 3 that is the smallest distance from the cylinder drum axis 70, i.e. is closer to the cylinder drum axis 70. In contrast, when the piston 2 reaches its bottom dead center UT, the piston 2 would protrude beyond the part of the inner wall 32 of the piston bore 3 that is the greatest distance from the cylinder drum axis 70. In the illustration of the Figure 2 Both pistons 2 would thus press against the respective right cylinder wall 31. In the case of a rigid piston head 21 and a rigid cylinder drum 7, this would inevitably lead to the piston head 21 jamming in the piston bores 3.

This jamming is counteracted in two ways in the floating piston machine 1 according to the invention. Firstly, the piston plate 8 is mounted so that it can move on the swash plate 9. The pressures from the piston chambers 31 are transmitted via the rigid pistons 2 to the piston plate 8 and move the piston plate 8 on the swash plate 9. This is shown in the Figure 2 This can be seen in that the piston plate axis 80 is now located to the left of the swash plate axis 90. On the other hand, the sealing ring 5, because it is slidably accommodated within the sealing receptacle 4, can avoid the forces acting on the sealing ring 5 from the inner walls 32 of the piston bore 3 transversely to the piston longitudinal axis. The inner diameter d _i of the sealing ring 5 and the diameter dz of the pin are ideally coordinated so that the resulting clearance δ _Q is large enough that the sealing ring 5 can follow the elliptical path in conjunction with the displacement of the piston plate 8 on the swash plate 9 without jamming. If this clearance is set correctly, a torque can be transmitted from the cylinder drum 7 via the sealing rings 5 to the piston plate 8, so that the piston plate is carried along by the cylinder drum 7. Alternatively, the piston plate 8 can be synchronized with the cylinder drum 7, for example via a gear, which opens up greater freedom with regard to the inner sealing ring geometry and the pin 23.

Due to the partially spherical outer surface 52 of the sealing ring 5 with a radius of curvature r that essentially corresponds to half the piston bore diameter D/2, the piston bore inner wall 32 and the sealing ring 5 touch in a circular line, the sealing circle 59, regardless of how much the piston longitudinal axis 20 is tilted relative to the piston bore axis 30 and thus how deeply the piston 2 penetrates into the piston bore 3 during its stroke movement. As a result, the plane in which the sealing circle 59 lies is always perpendicular to the piston bore axis 30. This reduces wear in the contact between the sealing ring and the piston bore and in turn makes the axial piston machine more efficient and robust. The service life of the metal sealing ring 5 is therefore significantly longer than an elastic sealing ring according to the state of the art.

In the following, the circular line on which the piston bore axes 30 are distributed around the cylinder drum axis is referred to as the piston bore pitch circle and the diameter of the piston bore pitch circle is referred to as the piston bore pitch circle diameter D _z . The piston feet 22 and in particular the piston longitudinal axes 20 of the individual pistons 2 intersect the piston plate 8 perpendicularly and are evenly distributed on a circular line, which is referred to below as the piston pitch circle, around the piston plate axis 80. The diameter of the piston pitch circle is referred to below as the piston pitch circle diameter D _K.

In one embodiment, the pistons 2 are arranged on the piston plate 8 such that in the neutral position the longitudinal axes 20 of the pistons 2 and the longitudinal axes 30 of a respective piston bore 3 coincide. Thus, the piston pitch circle diameter D _K and the piston bore pitch circle diameter D _z are identical. If the distance R of the piston bore axes 20 to the cylinder drum axis 70 is 45 mm as mentioned above, the piston bore pitch circle diameter D _Z is D _Z = 2R = 90 mm and the piston pitch circle diameter D _K is also 90 mm.

However, it has been shown that the piston pitch circle diameter D _K can also be chosen to be larger than the piston bore pitch circle diameter D _z . In a second design variant, the piston pitch circle diameter D _K is chosen to be 90.4 mm. A piston pitch circle diameter D _K which is larger than the piston bore pitch circle diameter D _z has the advantage that the floating piston machine can be built more compactly because a larger swivel angle α can be achieved with the same clearance δ _Q. A piston pitch circle diameter D _K which is larger than the piston bore pitch circle diameter D _z is made possible by the sealing rings 5 which are mounted so as to be displaceable transversely to the piston axis 20 and which compensate for the larger piston axis distance D _K by moving the sealing rings 5 in the sealing ring holder 4.

In another in Figure 8 In the embodiment shown, the inner wall of the sealing ring 5 is provided with an inner bead 54, so that the sealing ring 5 has, for example, a constant material thickness over its height h in the vertical direction. The background to a geometry of the sealing ring that deviates from the pure ring shape is as follows:
If a piston chamber 31 of the cylinder drum 7 is connected to the high-pressure side via the connecting bores 33, this high pressure (up to 350 bar or more) acts on the inner wall 32 of the bore of the cylinder drum 7, which forms the piston chamber 31. It has been shown that this internal pressure force, despite the solid design of the cylinder drum 7, can lead to an expansion or deformation of the corresponding piston bore 3. Such a one-sided expansion would lead to an increase in the gap 34 between the piston bore 3 and the sealing ring 5. In order to compensate for this disadvantage, the invention proposes that the geometry of the sealing ring 5 be designed in such a way that when the inside of the sealing ring 5 is subjected to radial pressure forces, the sealing ring can expand accordingly and thus the gap 34 between the piston bore 3 and the sealing ring 5 ideally remains constant over the entire range of the operating pressure. Through the clearance δ _Q and δ _H, the pressure finds its way into the area behind the sealing ring or into the space between the pin 23 and the inner diameter of the sealing ring 5. Since the working pressure in the piston chamber 31 acts at the same level on the inner geometry of the sealing ring 5, the sealing ring 5 will expand accordingly if the wall thickness or cross-sectional profile is adjusted accordingly.

In a first variant, this can be achieved by providing the sealing ring 5 with a bead-like recess 54 on its inner side 53. This bead-like Recess 54 can, for example, be designed such that the sealing ring 5 has an approximately uniform horizontal thickness z over its vertical course h. Due to this uniform horizontal thickness z, the sealing ring can be deliberately weakened in order to give way to a pressure acting on the inside of the sealing ring by widening, i.e. increasing its outer diameter d _a .

In an example of the sealing ring, which is not part of the invention but is prior art that facilitates the understanding of the invention as described in Figure 7 As shown, a reduction in the sealing ring wall thickness is achieved by designing the sealing ring 5 asymmetrically. This means that the height h ₂ of the sealing ring measured upwards from its equatorial plane 58 is greater than the height h ₁ of the sealing ring measured downwards from its equatorial surface 58. In this way, the smaller wall thickness z ₂ of the sealing ring 5 at its upper end compared to the wall thickness z ₁ of the sealing ring at its lower end is deliberately accepted in order to accommodate the high pressure of the pressure medium in the piston interior. In this way, the desired expansion of the sealing ring can be set via the upper height h ₂ .

In another in Figure 9 In the embodiment shown, the inner diameter of the sealing ring is stepped. In its upper part, i.e. the part facing the cover of the piston 2, the inner diameter d ₂ is chosen to be larger than the inner diameter d _i in its lower part. This provides an alternative to an approximately constant sealing ring cross-sectional thickness z according to the Figure 6 The embodiment shown achieves that the sealing ring 5 yields to a higher operating pressure in its upper region due to the lower material thickness z ₂ , while the sealing ring 5 largely retains its shape in its lower region due to the higher material thickness z ₁ and thus the adjustment between the sealing ring inner diameter d _i and the pin diameter d _z is not changed. The desired widening of the sealing ring in its upper region can be adjusted in particular by the upper diameter d ₂ and the height at which the gradation between the upper and lower regions is arranged.

In an example" which is not part of the invention but a prior art which facilitates the understanding of the invention, which is in Fig.10 As shown, the inner diameter of the sealing ring expands continuously over its height upwards, whereby the wall thickness of the sealing ring 5 decreases even more with the height and can thus more easily yield to the pressure of the sealing ring in the interior 57. In its lower area, the sealing ring 5 extends over a first height h ₁ from the equatorial plane downwards, and in its upper area over a second height h ₂ upwards. Depending on how much expansion is required, the expansion of the interior 57 of the sealing ring 5 can begin at the equatorial plane 58 as shown, but can also begin above or alternatively below the equatorial plane 58. For this purpose, both a symmetrically designed sealing ring 5, in which the first height h ₁ is chosen to be equal to the second height h ₂ , as well as in Figure 10 As shown, an asymmetrically designed sealing ring 5 can be used, in which the first height h ₁ is selected to be different from the second height h _2. If required, a geometry-optimized design of the ring geometry as a function z(h) over the height of the sealing ring 5 can also be determined with sufficient accuracy, for example, via corresponding deformation analyses using the finite element method.

Since the expansion of the piston inner wall 32 depends on many factors, such as the material used for the cylinder drum 7, the piston bore diameter d, the wall thicknesses between two adjacent piston bores 3, to name the most important, no general formula can be given here. However, laboratory tests have shown that at operating pressures of 350 bar, the expansion of the piston bore 3 with the dimensions selected in the exemplary embodiment can be between 10 µm and 30 µm, and in special individual cases even more or less. One method of determining the cross-sectional thickness z of the sealing ring therefore consists in first determining the deformation of the piston bore 3 at the highest intended operating pressure in a first step. In a series of tests, sealing rings 5 with different cross-sectional thicknesses z are exposed to the highest intended operating pressure and the resulting increase in diameter Δd of the sealing ring 5 is determined. The sealing ring geometry is then selected, that is, in this case the sealing ring 5 with the cross-sectional thickness z in which the difference Δd between the measured piston inner wall diameter d + Δd under load with the highest operating pressure and the sealing ring diameter d _i + Δ d _i under load with the highest operating pressure corresponds to the selected clearance between the piston inner wall 32 and the sealing ring 5.

Alternatively or in addition to a pressure equalization via the vertical and horizontal play of the sealing ring 5 in the sealing ring holder 4, a pressure equalization between the piston interior 31 and the interior 57 of the sealing ring 5 can also be achieved by one or more openings in the cover 6. Figure 11 shows an embodiment of a piston 2 with a sealing ring 5 with a bead-like recess on the inner wall 54 of the sealing ring 5. In this embodiment, a pressure equalization between the piston interior 31 and the interior 57 of the sealing ring 5 is provided by one or more pressure equalization bores 9, which extend from the top of the cover 6 through the pin 23 downwards and then in the radial direction of the pin 23. Such a pressure equalization is suitable both for sealing rings 5 with a continuous course of the sealing ring thickness z, and, as in Figure 12 shown for Sealing rings with a stepped inner profile. In this embodiment, pressure equalization between the piston interior 31 and the interior 57 of the sealing ring 5 is also provided by one or more pressure equalization bores 9, which extend from the top of the cover 6 through the pin 23 downwards and then in the radial direction of the pin 23.

Claims (15)

1. An axial piston machine (1), in which pistons (2) in cylinders (3) execute a stroke movement, and in which the pistons (2) have a sealing ring seat (4) for a sealing ring (5), the sealing ring seat (4) being designed such that it permits a movement of the sealing ring (5) transversely to a longitudinal axis (20) of the piston (2),

   the sealing ring (5) being spherical at least in a region which effects a seal during the stroke movements on inner walls (32) of the cylinder (3), the radius of curvature of the sealing ring (5), which is spherical in certain regions, corresponding substantially to half the diameter (d) of the cylinder (3),

   the cross section of the sealing ring (5) being designed such that, at a high operating pressure, the deformation of the sealing ring (5) by the operating pressure largely compensates for a widening of the inner cylinder wall (32) by the operating pressure,

   characterized in that

   the central inner opening (51) of the sealing ring (5) has a circumferential bead-like recess (54), or

   the central inner opening of the sealing ring (5) has a stepped profile (55, 56).
2. The axial piston machine (1) according to claim 1,
   characterized in that
   the sealing ring (5) consists of a rigid material which is, in particular, resistant to wear.
3. The axial piston machine (1) according to either of the preceding claims,
   characterized in that
   the sealing ring seat (4) comprises a pin (23) and the sealing ring (5) has a central inner opening (51) corresponding to the pin (23), the inner diameter (d_i) of the sealing ring (5) is selected to be greater than the pin diameter (d₂).
4. The axial piston machine (1) according to any of the preceding claims, characterized in that the piston (2) is designed such that pressure compensation between the piston interior (31) and the interior (57) of the sealing ring is made possible.
5. The axial piston machine (1) according to claim 4, characterized in that a horizontal clearance (δ_Q) between the inner diameter of the sealing ring (5) and the pin (23) and a vertical clearance (δ_H) of the sealing ring (5) within the sealing ring seat (4) are selected to be at least great enough that they make the pressure compensation between the piston interior (31) and the interior (57) of the sealing ring (5) possible.
6. The axial piston machine (1) according to claim 4, characterized in that the pressure compensation between the piston interior (31) and the interior (57) of the sealing ring (5) is optionally made possible by one or more openings in the cover (6) and/or one or more pressure compensation bores which extend from the upper side of the pin (23) into the interior of the sealing ring (5).
7. The axial piston machine (1) according to any of claims 1 to 6,
   characterized in that
   the sealing ring (5) is secured in the sealing ring seat (4) against a movement along the longitudinal axis (20) of the piston (2) by a cover (6).
8. The axial piston machine (1) according to claim 7,
   characterized in that
   the cover (6) is attached to the piston (2) by means of a screw or by clamping or by pressing.
9. The axial piston machine (1) according to any of claims 1 to 8,
   characterized in that
   the piston (2) is fastened to a piston plate (8) by one end (22).
10. The axial piston machine (1) according to any of claims 1 to 9,
    characterized in that
    the piston diameter in the region between the sealing ring seat (4) and the one end (22) tapers increasingly.
11. The axial piston machine (1) according to claim 10,
    characterized in that
    the piston (2) has the shape of a truncated cone in the region between the sealing ring seat (4) and the one end (22).
12. The axial piston machine (1) according to any of claims 1 to 11,

    in which the cylinders (3) are distributed over a cylinder barrel (7) about a cylinder barrel axis (70), and in which the pistons (20) are distributed over a piston plate (8) about a piston plate axis (80),

    characterized in that

    a rotation of the cylinder barrel (7) about the cylinder axis (70) and a rotation of the piston plate (8) about the piston plate axis (80) are synchronized with each other by synchronization means, the synchronization not taking place by a torque transmission via the pistons (2).
13. The axial piston machine (1) according to any of claims 1 to 12, in which the piston bore axes (30) of the cylinders (3) are distributed on a first circular line about a cylinder barrel axis (70), and in which the piston longitudinal axes (20) are distributed on a second circular line about a piston plate axis 80,
    characterized in that
    the diameter (D_K) of the second circular line is selected to be greater than the diameter (Dz) of the first circular line.
14. A method for producing a sealing ring according to any of claims 1 to 13,
    characterized in that
    a solid sphere is selected as the starting product, and in that two spherical segments are removed parallel to a great circle of this solid sphere, as a result of which a spherical disk is obtained.
15. The method for producing a sealing ring according to claim 14,
    characterized in that
    a central bore is made through the axis of rotation of the spherical disk.

Images