CN113994091A

CN113994091A - Radial piston machine with spherical pistons

Info

Publication number: CN113994091A
Application number: CN202080043421.6A
Authority: CN
Inventors: 德克·比彻; 丹尼尔·弗拉赫
Original assignee: Moog GmbH
Current assignee: Moog GmbH
Priority date: 2019-06-19
Filing date: 2020-06-18
Publication date: 2022-01-28
Also published as: EP3987183A1; WO2020254501A1; US20220349394A1; DE102019116680A1

Abstract

The invention relates to a radial piston machine (1) having cylinders (5) arranged in a cylinder carrier (16) and a piston element (21) arranged in each cylinder (5) and connected to a guide element (22), wherein the guide element (22) runs on a sliding surface (14) and thereby forces the piston element (21) to a stroke movement. The piston element (21) is designed in the region of the piston element (21) in a spherical manner, which during the stroke movement leads to a seal at the inner wall (51) of the cylinder (5), whereby a linear seal is formed, which makes possible a more compact construction compared to radial pumps with cylindrical piston elements.

Description

Radial piston machine with spherical pistons

The invention relates to a radial piston machine having a piston which executes a stroke movement in a cylinder. Such radial piston machines can be used as work machines, for example as pumps, but also as motors. In general, for all radial piston machines, cylinders are arranged in the rotor and a piston element is arranged in each cylinder, which is connected to a guide element, wherein the guide element runs on a sliding surface and thereby forces the piston element to a stroke motion.

In radial piston machines, the cylinders are arranged radially with their longitudinal axis in the rotor. The radial piston machine belongs to a hydraulic displacer, which works according to the displacement principle. Thus, these hydraulic displacers can be operated both as pumps and as motors when the pressure medium flow is controlled correspondingly. The pump and motor are generally of the same design.

The radial piston machine can be further divided into an internally pressurized radial piston machine and an externally pressurized radial piston machine. In the case of an internally pressurized radial piston machine, the working chamber of the cylinder is filled and emptied from the inside, i.e. for example via a radial hollow shaft, with a pressure medium. In this case, the cylinder block rotates about the radial hollow shaft. In this case, the pistons arranged in the cylinders are supported at the outer ring, so the internally pressurized radial piston machine is referred to as an externally supported radial piston machine. An outer ring is located eccentrically to the hollow shaft, at which outer ring the working piston is supported.

In contrast, in the case of an externally pressurized radial piston machine, the pressure medium is fed radially from the outside into the cylinder block, wherein the piston arranged in the cylinder block is supported on a centrally arranged eccentric shaft. An externally pressurized radial piston machine is therefore also referred to as an internally supported radial piston machine.

In commercial radial piston pumps, drive torque is transmitted from a drive shaft to a cylinder star, wherein a plurality of radially oriented cylinders are arranged in a star. The cylinder star is rotatably provided on the control shaft diameter. The pistons arranged radially in the cylinders of the cylinder star are supported via static pressure-relieved piston shoes on a stroke ring arranged eccentrically to the cylinder star. The piston and the slipper are connected to each other via a ball joint and locked by a snap ring. Alternatively, a flange may be provided for the ball joint. The oil flow is fluidly connected with the cylinders of the cylinder star via a delivery channel and a discharge channel in the housing. When the cylinder star rotates, the piston performs a stroke movement due to the eccentric position of the stroke ring, and wherein oil is sucked from the delivery channel in the expansion phase and pushed into the discharge channel in the compression phase.

EP 0011145B 1 discloses in particular a slipper for a hydrostatic annular piston machine, in which a shank with a ball head is arranged in the spherical surface of the piston. In this case, the spherical surface is designed as a stepped bore penetrating the piston in the longitudinal direction. The piston itself is disposed in a cylinder bore located within the cylinder body. A snap ring secures the piston and the slipper together.

In radial piston machines, in particular cylindrical pistons are used. In order to reduce the friction of the pistons at the inner wall of the cylinder, these pistons must be embodied to be correspondingly long with respect to their diameter. This increases the volume of such radial piston machines, since the cylinder star must be embodied with a correspondingly sufficient diameter in order to provide a corresponding installation space for the cylinder.

Since the piston machine described, by virtue of its type of construction, determines the application of pressures of more than 300bar, the joint point between the piston and the shoe is exposed to high alternating loads. In particular, the joint itself can only withstand a certain maximum tensile force. If this force is exceeded, the joints separate, resulting in a failure of the piston machine.

The object of the invention is therefore to implement a radial piston machine in such a way that friction is reduced. Another object of the invention is to increase the allowable tensile forces at the piston or to extend the service life of the piston before defects occur.

In a radial piston machine, the object is solved in that the piston element is designed in the form of a sphere at least in the region of the piston element, which piston element causes a seal at the inner wall of the cylinder during the stroke movement. The spherical design of the piston element produces a sealing region which is annular, i.e. forms a closed circular line. The closed circular line results in much less friction than a face seal by a cylindrical piston. Although the position of the circular sealing line on the surface of the at least partially spherical piston changes during the rotation, or tilting movement, of the spherical piston element, the resulting gap between the inner wall of the cylinder and the spherical piston element is always exactly the same, irrespective of the position of the spherical piston element in the cylinder and irrespective of the tilt angle of the spherical piston element, because the diameter of the circular sealing line is constant on account of the spherical shape and the inner diameter of the cylinder is also constant. In this way, cost-intensive friction-reducing coatings or also tribological contours in the outer layer of the piston can be dispensed with. Likewise, the technical expenditure to produce a sufficiently perfect spherical shape is much lower than to produce a conventional longitudinal piston with comparable performance to be sufficiently perfect; or in other words, according to the prior art, a lower manufacturing quality can be selected for the spherical piston element than for the longitudinal piston, which achieves the same object.

Since in the ball the diameter of the great circle is however constant, regardless of the direction in which the ball rotates, the piston element does not become stuck in the cylinder during the stroke movement, since the diameter of the respective sealing circle line remains unchanged compared to the diameter of the cylinder. The at least partially spherical piston element is also self-centering by the pressure exerted by the face. Thereby reducing losses in the piston machine and wear in the piston guide in the cylinder.

Since the at least partial spherical shape of the spherical piston reduces the overall length of the piston on the piston axis, a smaller rotor diameter can be selected for the same output, thereby resulting in a more compact construction of the radial piston machine. At the same time, a higher working pressure can also be selected, since the disadvantages of radial piston machines are no longer determined by the joint connection. Thus, the circular piston eliminates the restriction of the internal pressure of the housing by tensile, compressive and lateral forces acting at the joint. In piston machines with longitudinal pistons, restrictions have been created with regard to the internal pressure of the housing or the back pressure in the external leakage line, which is caused, for example, by the increased position of the hydraulic tank, filter or cooler in the leakage line. With the elimination of this limitation, the range of applications of radial piston machines in this application is expanded or made possible for some applications.

Mathematically, the surface of the partially spherical piston that contacts the inner wall of the cylinder is a symmetrical spherical area. The spherical zone is a curved outer side of, for example, a spherical disc or a spherical ring. The spherical disc, or spherical layer as it is called, 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 in this case lie on different sides of the spherical center point and at the same time have the same distance from the spherical center point, this is a symmetrical spherical disk whose outer surface just produces a symmetrical spherical region. But symmetrical spherical zones are also obtained by means of radial holes through the sphere.

In a prototype of the radial piston pump according to the invention, a tilting angle α of the spherical region of about 9 ° to both sides is produced due to the geometry selected. It is advantageous to choose an angle of inclination a of about 10 deg., better still 12 deg., with corresponding tolerances. Thickness of the spherical disc or height H of the spherical disc and diameter d of the spherical disc_KThe ratio corresponds to a twice tangential function of the tilt angle alpha. If the inclination is 12 ° to both sides, a thickness H of the spherical disk and a diameter d of the spherical disk of about 0.4 result_KThe ratio of. It is naturally clear to the skilled person that this may be a prompt only. Depending on the geometry of the radial piston machine, in individual cases, larger values of the angle of inclination α may also be required, or smaller angles of inclination α may also be sufficient. A tilt angle α of up to 20 ° appears technically reasonable or achievable.

Advantageously, the at least partially spherical piston element can be connected to the guide element via a connecting element, wherein the connecting element is rigidly connected to the piston and/or to the sliding element. As long as the movement of the spherical piston element is not limited by the forced guidance or the connecting element does not come into contact with the cylinder wall, the spherical piston element is completely free to rotate in the cylinder by means of the spherical portion of the piston element, in order to implement, for example, a pitch, yaw or roll movement. Generally, the movement of the piston element is limited to pitch and stroke movements by forced guidance. In most applications, it is therefore no longer necessary to movably connect the sliding elements or piston elements to one another, since the necessary rotational movement can be carried out by means of piston elements which are designed to be at least partially spherical. In this case, the loss of the piston machine and the wear in the piston guide in the cylinder are reduced.

In this case, the skilled person will select the length and diameter of the connecting element corresponding to the desired kinematics and the desired tilt angle α. In particular, in radial piston machines, the slipper has a rounded profile matching the pivot angle and the cylinder bore. The profile of the slipper need only be able to reflect the kinematic pivoting motion of the slipper in relation to the type of construction.

The piston according to the invention with the at least partially spherical piston element can be produced, for example, by means of turning, milling or grinding. Alternatively, it is possible to produce correspondingly assembled individual components by means of casting, additive manufacturing, MIM technology, sintering or using shaped components.

In particular applications, the piston element can of course also be connected to the connecting element by means of a further joint and/or the connecting element can be connected to the guide element by means of a joint.

In this case, the radial piston machine may be an internally pressurized radial piston machine or an externally pressurized radial piston machine. In the case of internally pressurized radial piston machines, the guide element runs on a stroke ring arranged eccentrically to the radial axis of the cylinder support. In the case of externally pressurized radial piston machines, the guide element runs on an eccentric shaft which rotates eccentrically to the center point of the cylinder support in the interior of the cylinder support.

The invention will now be further described and explained with reference to embodiments depicted in the drawings. The figures show:

FIG. 1 is a radial piston machine with a circular piston according to the invention

FIG. 2 is a circular piston according to the present invention

FIGS. 3A, 3B and 3C are portions of a radial piston machine

FIG. 4 is a rotary piston machine with external support of the circular piston according to the invention

The invention relates to a new piston for a radial piston machine, wherein the radial piston machine can be a radial piston machine according to the prior art, in addition to the new piston. Due to the shape of the piston, it will be referred to as a circular piston in the following, although strictly speaking, the portion of the circular piston that is in contact with the inner wall of the hollow cylinder need only correspond to the spherical portion. In this sense, therefore, when a sphere is referred to hereinafter for simplicity of language, it is intended to include only a partial sphere.

Fig. 1 shows a simplified view of an internally pressurized and externally supported radial piston machine 1 in a partial sectional view. Generally, such radial piston machines can be operated both as pumps and as motors, as long as there are no design features to prevent. In the following, the operation of the pump is described as representative of these two operation modes. The principles of such radial piston pumps are well known to the skilled person and therefore only those essential parts which are necessary for an understanding of the present invention will be described herein. The radial piston pump 1 has a housing 10 which is designed like a pot and is closed by a housing cover which is not shown. A stroke ring 12 which is movable in the adjustment direction 3 is arranged in the interior 11 of the housing 10, said ring being arranged such that its lateral surfaces have a corresponding gap between the housing base and the projection of the housing cover. Due to the sectional view, only the side 13 of the stroke ring 12 facing the viewer, which is located on the inner side of the housing cover when the housing 10 is closed, is visible in fig. 1. In this figure, the projection of the housing base is hidden by the stroke ring 12, to be precise omitted for the sake of clarity.

The inner circumference of the stroke ring 12 forms the sliding surface 14 for the shoes 22 on which the spherical pistons 21 are supported, which are guided displaceably in the radially extending bores 5 of the cylinder support. Since in this type of construction a rotary movement of the cylinder support is provided, the cylinder support is referred to hereinafter as rotor 16. These holes will be referred to hereinafter as piston bores 5 due to their interaction with the spherical piston 21.

The piston bores 5 are distributed rotationally symmetrically about a radial axis 17 of the rotor 16. The number of piston bores depends in part on the size of the rotor 16 or the displacement or absorption volume of the radial piston pump 1. In the example shown here, i.e. at a maximum working pressure of 350bar, with 19cm³In a/U-displacement or absorption volume radial piston pump, seven piston bores 5 are provided in the rotor 16.

The rotor 16 is arranged on a control shaft diameter 18 which is arranged fixedly in a control shaft diameter bore of the housing 10 and is rotated by a drive shaft. In the motor mode of operation, the torque generated by the spherical pistons 21 is transmitted to the drive shaft, which in the present technology is correctly referred to as the output shaft. In this case, the ball piston 21 seals the working chamber of the radial piston machine 1 from the interior 11 of the radial piston machine 1, wherein the working chamber in the piston bore 5 extends from the extension of the ball piston 21 in the direction of the control shaft diameter 18 to the control shaft diameter 18. In fig. 1, the control shaft diameter hole and the drive shaft are hidden by the control shaft diameter 18, i.e. on the truncated side of the image, and are therefore not visible.

The spherical piston 21 and the slipper 22 are connected to each other by a piston rod 23. In this case, the piston rod 23 can also be embodied as a connecting rod, i.e. it can be movably connected to the ball piston 21 via a joint arranged at the ball piston 21. Alternatively or in addition, a piston rod 23 or a connecting rod with a joint arranged at the slipper 22 may be movably connected with the slipper 22. However, a synergistic effect occurs in particular when the piston rod 23 is rigidly connected both to the ball piston 21 and to the slipper 22. In this case, the spherical shape of the spherical piston 21 is used not only as a seal between the working chamber and the interior 11 of the radial piston machine 1, but also as the sole joint. This is particularly advantageous in that weakening of the combination of the spherical piston 21, the piston rod 23 and the slipper 22 by a joint is avoided. In particular in the case of a rigid connection of the circular piston 21 and the sliding shoe 22, the piston rod 23 can be embodied as a taper of the circular piston 21 in such a way that it achieves high strength.

The spherical or spherical part shape allows the spherical piston 21 to be guided within the piston bore 5 by the slipper 22, in particular to perform a limited tilting movement in the direction of rotation of the rotor 16 or in the direction opposite thereto in its plane of rotation. In this case, the tilting movement of the circular piston 21 is limited by a piston rod 23 which can stop at the piston bore wall, in particular at the piston bore opening directed to the stroke ring 12. The ball piston 21 is not subjected to lateral movements by a forced guidance of both sides of the skid shoe 22 between not shown ring strips at the inner side of the stroke ring 12.

The stroke ring 12 can be displaced in the interior 11 transversely to the control shaft diameter 18 in the adjusting direction 3 by means of two adjusting

pistons

31, 32 for the purpose of varying the flow rate. The two adjusting

pistons

31, 32 act on the outer circumference of the stroke ring 12 at two diametrically opposed points with adjusting

shoes

33, 34.

In the control shaft diameter 18, for the introduction of pressure medium, a first low-pressure channel 41 and a second low-pressure channel 42 are formed, each extending longitudinally in parallel, eccentric to the center axis of the control shaft diameter 18, which low-pressure channels each open into a first circumferential groove at the control shaft diameter 21, which first circumferential groove is referred to as a low-pressure groove 45 hereinafter. Further, for discharging the pressure medium, a first high-pressure channel 43 and a second high-pressure channel 44 are formed, each extending longitudinally in parallel, eccentrically to the center axis of the control shaft diameter 21, and these high-pressure channels each lead to a second circumferential groove at the control shaft diameter 21, which is hereinafter referred to as a high-pressure groove 46. The low-pressure groove 45 and the high-pressure groove 46 are located in the passage area of the rotor 16 which accommodates the piston bore 5 of the circular piston 21. The first low-pressure channel 41 and the second low-pressure channel 42 end in the low-pressure connection of the radial piston pump 1, and the first high-pressure channel 43 and the second high-pressure channel 44 end in the high-pressure connection of the radial piston pump 1. The low-pressure connection and the high-pressure connection of the radial piston pump 1 are not visible in this figure, since they are located on the rear side of the housing base from the observer. In this exemplary embodiment, two low-

pressure channels

41, 42 and two high-

pressure channels

43, 44 are selected, since this, in combination with the known special geometry of the control shaft diameter 18, offers flow-technical advantages. However, a single low-pressure duct 41 and a single high-pressure duct 43 are sufficient to implement the basic principle of the radial piston machine 1.

In operation of the radial piston pump 1, that is to say when the rotor 16 is rotated, the ball pistons 21 are carried along in the piston bores 5 in the direction of the rotational movement. In this case, due to the rotational movement of the rotor 16, centrifugal forces act on the ball pistons 21 guided in the piston bores 5, so that the respective ball pistons 21 in the piston bores 5 are pressed radially outward until the shoes 22 of the ball pistons 21 abut with their ends remote from the control shaft diameter 18 against the stroke ring 12. Hereinafter, the expression "outward" indicates a direction pointing away from the rotation axis 17 of the rotor 16, and the expression "inward" indicates a direction pointing toward the rotation axis 17 of the rotor 16. Likewise, the term "outer" refers to the relative position of an object having a larger radial distance from the rotational axis 17 of the rotor 16 than an object having a smaller radial distance from the rotational axis 17 of the rotor 16.

In the case where the position of the stroke ring 12 is adjusted such that the imaginary axis of the stroke ring 12 is arranged in line with the rotation axis 17 of the rotor 16, the distance D between the outer side of the rotor 16 and the inner side of the stroke ring 12 is the same in each position of the rotor 16. In this case, the radial distance of the ball pistons 21 from the axis of rotation 17 does not change during the rotational movement of the rotor 16, so that the ball pistons 21 do not perform a stroke in the respective piston bores 5.

When the stroke ring 12 is arranged such that the imaginary axis 19 of the stroke ring 12 no longer coincides with the rotational axis 17 of the rotor 16, the distance D between the rotor 16 and the stroke ring 12 is periodically varied while the rotor 16 is cyclically reciprocated within the stroke ring 12. This change in distance results in the sliding surface 14 of the stroke ring 12 exerting a reaction force on the shoes 22, which presses the ball piston 21 inwards against the centrifugal force, when the distance D between the stroke ring 12 and the rotor 16 decreases. When the distance D between the stroke ring 12 and the rotor 16 increases, the stroke ring 12 no longer exerts a force on the shoes 22 of the respective spherical piston 21 and presses the spherical piston 21 radially outward by centrifugal force, also by compressive force, so that the respective shoes 22 do not lose contact with the sliding surface 14 of the stroke ring 12. Correspondingly, during the complete rotation of the rotor 16, the spherical piston 21 is forced in two periodic stroke movements, the first stroke movement being radially outward, wherein the volume of the working chamber increases continuously, and therefore is referred to hereinafter as the expansion phase, and the second stroke movement being radially inward, in the direction of the control shaft diameter 18, wherein the volume of the working chamber decreases continuously, and therefore is referred to hereinafter as the compression phase.

In this case, the low-pressure groove 45 is arranged on the control shaft diameter 18 in such a way that, during the expansion phase, the low-pressure groove 45 and thus also the first and second low-

pressure channels

41, 42 are in fluid connection with the respective ball piston 21. The radially outward stroke movement of the ball piston 21 thus generates a suction effect which sucks in pressure medium at the low-pressure connection and fills the working chamber of the piston bore 5 with pressure medium. Further, the high-pressure groove 46 is in this case arranged on the control shaft diameter 21 such that, in the compression phase, the high-pressure groove 46 and thus also the first and second high-

pressure channels

43, 44 are in fluid connection with the respective ball piston 21. The radially inward stroke movement of the ball piston 21 thus generates a pressure effect which pushes pressure medium which accumulates in the working chamber of the respective piston bore 5 via the high-pressure groove 46 into the high-

pressure channels

43, 44.

Thus, in the pump mode of operation, the periodic stroking movement of the spherical piston 21 forces pressure medium from the low-pressure channel to the high-pressure channel. In contrast, in the operating mode of the motor, the pressure medium flows from the high-pressure channel to the low-pressure channel. Depending on whether the radial piston machine 1 is operated as a pump or as a motor, in the pump mode pressure can be transmitted to the radial piston machine by means of a drive torque, or in the motor mode pressure energy can be extracted from the radial piston machine and converted into an output torque.

Fig. 2 shows a piston element 2 comprising a ball piston 21 and a slipper 22, which are rigidly connected to each other by means of a piston rod 23. The bore 29, through which the pressure medium located in the piston bore 5 is pressed onto the slipper bottom 26 during operation of the piston machine 1, extends through the longitudinal axis of the ball piston 21 and the piston axis 20. Thereby forcing the shoe sole 26 into a state of hydrostatic equilibrium in which a pressure medium film reducing friction is formed between the shoe sole and the sliding surface 14. In the radial piston machine 1, the shoe sole 26 is adapted to the geometry of the sliding surface 14, i.e. it is curved outwards. Further, as shown in fig. 3, the slipper 22 has a front slipper end 27, as seen in the direction of rotation about its piston axis 20, and opposite thereto a rear slipper end 28.

In principle, the spherical piston 21 can take on a completely spherical shape, with the exception of the piston rod 23. However, it is only necessary to have the ideal perfect spherical shape at the points of the spherical piston 21 which are in contact with the piston bore wall 51, or strictly speaking, the piston bore wall, in order to seal the piston bore space. In the following, the great circle of the spherical piston 21, which is perpendicular to the piston axis 20, is referred to as the middle circle 24. When the piston axis 20 coincides with the piston bore axis 50, the extension of the plane spanned by the central circle 24 outside the central circle 24 intersects the piston bore to form an intersecting circle, which is referred to as a sealing circle in the following because it actually seals the working chamber of the piston bore 5 from the interior 11 of the piston machine 1. In this case, the sum of all sealing circles that can be formed during a complete rotation defines the spherical region 250, i.e. the outer circumference of the spherical disk, wherein the at least partially spherical piston element 71 must correspond to the ideal spherical shape.

To prevent the ball piston 21 from jamming in the piston bore 5, the diameter d of the central circle 24_KIs chosen to be slightly smaller than the diameter of the piston bore 5. For example, the middle circle 24 is 10 μm smaller, a gap of 5 μm is created between the working circle and the middle circle when the middle circle 24 and the piston bore wall 51 are in the center position. Due to the viscosity of the pressure medium, an adequate seal is created between the working chamber of the piston machine 1 and the inner chamber 11. Of course, the skilled person selects the gap between the ball piston 21 and the piston bore wall 51 such that it is most suitable for a given size and corresponding use.

Figure 3A shows the spherical piston 21 at its outer dead centre,that is to say the transition from the expansion phase to the compression phase. At the outer dead center, the distance between the rotor 16 and the sliding surface 14 of the stroke ring 12 is the maximum distance D due to the eccentric position of the stroke ring 12_max. Due to the centrifugal forces, i.e. the ring strips, the pressing device and other pressure forces, the ball piston 21 is generally oriented at its outer dead center such that the piston axis 20 more or less coincides with the piston bore axis 50. The geometry of the individual elements of the annular piston machine 1 is selected such that, in this case, the center circle 24 of the spherical piston 21 is located in the vicinity of the outer piston bore opening 52, but still at a sufficient depth in the piston bore wall 51 to ensure safe guidance of the piston 2 in the piston bore 5 and sealing between the piston bore wall 51 and the center circle 24.

On further rotation, starting from the outer dead center, the force acting through the piston bore wall 51 on the center circle 24 of the spherical piston 21 then carries the piston 2 together in the direction of rotation 15 of the rotor 16. Since the distance between the stroke ring 12 and the rotor 16 is smaller and smaller during the compression phase, the piston 2 is guided deeper inwards in the respective piston bore 5. The distance between the stroke ring 12 and the rotor 16 at the position of the leading shoe end 27 is different from the distance between the stroke ring 12 and the rotor 16 at the position of the trailing shoe end 27, limited by the eccentricity between the stroke ring 12 and the rotor 16. Thereby, a force acting from the front shoe end 27 to the rear shoe end 28 in the compression phase is established. The piston 2 of the ball piston 21 is thereby tilted about its centre point Z in the compression phase in the direction opposite to the direction of rotation 15 of the rotor 16. By this tilting movement, the middle circle 24 rotates in the direction of the outer piston bore opening 52 at the piston bore wall 51 facing away from the direction of rotation and in the direction of the inner piston bore opening 53 at the piston bore wall 51 facing the direction of rotation. The centre circle 24 of the ball piston 21 thus loses contact with the piston bore wall 51. Due to the spherical shape of the ball piston 21, the new great circle of the ball piston 21 now becomes the sealing circle 25, i.e. the great circle at the corresponding position of the ball piston 21 in the piston bore 5 perpendicular to the piston bore axis 50. As a geometrical consequence, the mid-point of each sealing circle coincides with the spherical mid-point Z of the spherical piston 21.

At the inner dead point IT of the slave piston 2In the transition to the outer dead point AT, i.e. in the expansion phase, the distance between the outside of the rotor 16 and the inside of the stroke ring 12 is from D_minContinuously increases to D_max. In the expansion phase, the distance D at the front shoe end 27 is therefore always greater than the distance at the rear shoe end 28. Thus, the reaction force component exerted by the stroke ring 12 on the piston 2 at the front shoe end 27 is smaller than the reaction force component at the rear shoe end 28. The piston 2 is thus deflected in the direction of the front shoe end 27 compared to the direction of rotation 15 of the rotor 16. Thus, in the expansion phase, the angle of inclination α leads the rotary movement 15. In the compression phase, i.e. in the transition between the outer dead point AT to the inner dead point IT, these relationships are reversed and the angle of inclination α lags behind the rotary movement 15 of the rotor 16. By the tilting movement, the piston axis 20 is tilted by an angle α relative to the piston bore axis 50. This tilting movement has on the one hand the largest extent, as shown in fig. 3B, in about half the movement between the inner dead point IT and the outer dead point AT of the piston 2, or in about half the movement of the piston 2 between the outer dead point AT and the inner dead point IT.

Finally, FIG. 3C shows when the distance between the stroke ring 12 and the rotor 16 reaches a minimum distance D_minA spherical piston 2 at its inner dead point. As can be seen, the piston base 210 adjoining the partially spherical piston 21 on the side remote from the piston rod 23 is of substantially conical design, so that the piston base 210 in the inner dead center IT can be adapted as far as possible to the funnel-shaped transition from the piston bore 5 to the control shaft diameter 18. This advantageously reduces the dead volume in the inner dead center IT.

According to the laws of geometry, the angle α is also equal to the angle α between the plane of the sealing circle and the plane of the mid-circle. For reasons of manufacturing technology, but also for possible changes in the direction of rotation, the ball surface of the ball piston 21 is embodied symmetrically, so that the spherical portion, i.e. the spherical region 250 as shown in fig. 2, comprises at least one angle between- α and + α. Further, depending on the chosen geometry of the circular piston machine 1 and the piston 2, the taper is chosen at the position of the piston rod 23 such that it does not abut the piston bore wall 51 or the rotor 16 during operation, in order to avoid damage to the piston bore rod 23 or the rotor 16 and the piston bore wall 51.

The invention is also applicable to internally supported radial piston machines. Fig. 4 shows the basic structure of the internally supported radial piston machine 6, wherein the low-pressure channel and the high-pressure channel are arranged on the outside 66 of the piston carrier 61, which is fixed in this case. The design of the low-pressure and high-pressure channels in the internally supported radial piston machine 6 is familiar to the skilled person and is therefore not explicitly shown in fig. 4 for the sake of clarity.

The piston holder 61 forms a receptacle for a plurality of piston bores 5, which are arranged radially around the center point 60 of the piston holder 61 at the same distance from one another, so that the extensions of the longitudinal axes 50 of the piston bores 5 intersect at the center point 60 of the piston holder 61. For the sake of clarity, only three piston bores 5 are shown in fig. 4. Typically an odd number between 3 and 9, but the skilled person will select the appropriate number of piston bores depending on the size and performance of the radial piston pump.

The interior of the piston carrier 61 is designed as a cavity in which the eccentric shaft 63 rotates about the piston carrier midpoint 60. The eccentric center point E is located at a distance D from the piston holder center point 60 and thus rotates on a circular path 64 around the piston holder center point 60.

In each of the piston bores 5, a circular piston according to the invention is arranged, which is essentially formed by a spherical piston element 71, a piston rod 73 and a sliding shoe 72. Each round piston is pressed in the direction of the eccentric shaft 63 or the eccentric center E, i.e. towards the interior of the radial piston machine, by a respective return element, which in this embodiment is designed as a helical spring 74. In the internally supported radial piston machine 6, the at least partially spherical piston element 71 seals the working chamber of the radial piston machine 6 inwardly. The working chamber, in which pressure medium is fed or discharged, is thus part of the piston bore 5, which extends between the outer circumference 66 of the piston carrier 61 and the at least partially spherical piston element 71.

The outer circumference of eccentric shaft 63 constitutes a sliding surface 65 on which a shoe 72 of circular piston 7 is supported by a sliding bottom 76. Since in this embodiment the sliding surface 65 of the eccentric shaft is shaped convex, the sliding bottom 76 is correspondingly shaped concave. During the complete rotation of the eccentric shaft 63, the ball piston 7 is forced by sliding of the sliding shoe 72 on the eccentric shaft 63 in two periodic stroke movements, the first stroke movement being radially outward, in which the volume of the working chamber is decreasing, and therefore is referred to hereinafter as the compression phase, and the second stroke movement being radially inward, in the direction of the piston support mid-point 60, in which the volume of the working chamber is increasing, and therefore is referred to hereinafter as the expansion phase. In this case, the distance D between the eccentric midpoint E and the piston cradle midpoint 60 determines the amplitude of the piston stroke.

Just like the externally supported radial piston machine 1 described in the first embodiment, the circular piston 7 performs a tilting motion in the expansion phase and the compression phase. Here, due to the at least partially spherical piston element 71, a circular sealing circle is also maintained at all times between the inner wall 51 of the piston bore 5 and the spherical region 75 of the round piston 7. Also in this embodiment, the play between the ball piston 7 and the piston bore wall 51 prevents the ball piston 7 from becoming stuck.

Claims

1. A radial piston machine (1) having cylinders (5) arranged in cylinder supports (16) and a piston element (21) arranged in each cylinder (5) and connected to a guide element (22), wherein the guide element (22) runs on a sliding surface (14) and thereby forces the piston element (21) to stroke, characterized in that,

the piston element (21) is designed in the form of a ball at least in the region of the piston element (21), said piston element causing a seal at an inner wall (51) of the cylinder (5) during the stroke movement.

2. Radial piston machine (1) according to claim 1,

it is characterized in that the preparation method is characterized in that,

the piston element (21) is connected to the guide element (22) via a connecting element (23), wherein the connecting element (22) is rigidly connected to the piston element (21).

3. Radial piston machine (1) according to claim 1,

it is characterized in that the preparation method is characterized in that,

the piston element (21) is connected to the guide element (22) via a connecting element (23), wherein the connecting element (23) is rigidly connected to the guide element (22).

4. Radial piston machine (1) according to one of claims 1, 2 or 3,

the guide element is a sliding shoe (22) which runs on the sliding surface (14).

5. Radial piston machine according to one of claims 1 to 4,

it is characterized in that the preparation method is characterized in that,

the radial piston machine is an externally supported radial piston machine.

6. Radial piston machine (1) according to claim 5,

it is characterized in that the preparation method is characterized in that,

the radial piston machine is a radial piston machine in which the guide element (22) runs on a stroke ring (12).

7. Radial piston machine according to one of claims 1 to 4,

it is characterized in that the preparation method is characterized in that,

the radial piston machine is an internally supported radial piston machine.

8. Radial piston machine (1) according to claim 7,

it is characterized in that the preparation method is characterized in that,

in the internally supported radial piston machine, the guide element (22) runs on an eccentric shaft (63).

9. Radial piston machine according to one of claims 1 to 8,

it is characterized in that the preparation method is characterized in that,

the area of the piston element (21), which results in a seal at the inner wall (51) of the cylinder (5) during the stroke movement, is a symmetrical spherical area (250), at least in the area of the piston element (21).

10. A radial piston machine according to claim 9,

characterized in that said symmetrical spherical zone comprises at least a 20 ° inclination to either side.