US20220349394A1

US20220349394A1 - Radial reciprocating engine having a ball piston

Info

Publication number: US20220349394A1
Application number: US17/618,267
Authority: US
Inventors: Dirk Becher; Daniel Flach
Original assignee: Moog GmbH
Current assignee: Moog GmbH
Priority date: 2019-06-19
Filing date: 2020-06-18
Publication date: 2022-11-03
Also published as: EP3987183A1; DE102019116680A1; CN113994091A; WO2020254501A1

Abstract

The invention relates to a radial reciprocating engine (1) having cylinders (5) arranged in a cylinder carrier (16) and a piston element (21) arranged in each cylinder (5) that is connected to a guide element (22), wherein the guide element (22) runs on a slide surface (14), whereby a stroke movement is imposed on the piston element (21). Since the piston element (21) is spherical at least in the region of the piston element (21) which seals the inner walls (51) of the cylinder (5) during the stroke movements, a linear seal is created, which enables a more compact design in comparison to radial pumps with cylindrical piston elements.

Description

The invention relates to a radial reciprocating engine with pistons that perform a stroke movement in a cylinder. Such radial reciprocating engines can serve either as work machines, for example as a pump, or also as a motor. Common to all radial reciprocating engines is that cylinders are arranged in a rotor and a piston element is arranged in each cylinder, which piston element is connected to a guide element, wherein the guide element runs off on a sliding surface and thereby imposes a stroke movement on the piston element.
In radial reciprocating engines, cylinders with their longitudinal axes are arranged radially in a rotor. Hydraulic displacement machines, which include radial reciprocating engines, operate on the displacement principle. They can therefore be operated both as pumps and as motors if the flow of the pressure-transmitting medium is controlled accordingly. Pumps and motors generally have the same structural design.
In radial reciprocating engines, a distinction can then still further be made between internally pressurized and externally pressurized radial reciprocating engines. In the case of internally pressurized radial reciprocating engines, working spaces of the cylinders are filled with a pressure medium from the inside and emptied therefrom, i.e., for example, via a radial hollow shaft. The cylinders rotate around the radial hollow shaft in this case. Pistons arranged in the cylinders are supported in this case on an outer ring, which is why internally pressurized radial reciprocating engines applied are also referred to as externally supported radial reciprocating engines. The outer ring, on which the working pistons are supported, is located in an eccentric position in relation to the hollow shaft.
In contrast to this, in the case of externally pressurized radial reciprocating engines, the pressure medium is supplied radially to the cylinders from the outside, wherein the pistons arranged in the cylinders are supported on a centrally arranged eccentric shaft. Externally pressurized radial reciprocating engines are therefore also referred to as internally supported radial reciprocating engines.
In a commercially available radial piston pump, a drive torque is transmitted from a drive shaft onto a radial piston cylinder block, in which a plurality of radially aligned cylinders is arranged in a star shape. The radial piston cylinder block is rotatably mounted on a control stud. Pistons arranged radially in the cylinders of the radial piston cylinder block are supported on a thrust ring mounted eccentrically in relation to the piston cylinder block via hydrostatically relieved slider shoes. Piston and slider shoe are connected to one another via a ball joint and secured by a snap ring. Alternatively, a flanging of the ball joint can also be provided. An oil stream is in fluidic connection with the cylinders of the radial piston cylinder block via supply and discharge channels in the housing. When the radial piston cylinder block rotates, as a result of the eccentric position of the thrust ring, the pistons execute a stroke movement and in an expansion phase suck in the oil from the supply channels and in a compression phase push the oil into the discharge channels.
EP 0 011 145 B1 discloses in particular a slider shoe for hydrostatic annular piston engines, in which a shaft with a spherical head is mounted in a spherical surface of a piston. In this case, the spherical surface is formed in a stepped bore penetrating the piston in the longitudinal direction. The piston itself is arranged in a cylinder bore located in a cylinder body. A snap ring holds piston and slider shoe together.
In radial reciprocating engines, cylindrical pistons are used in particular. In order to reduce the friction of the pistons on the inner walls of the cylinders, the pistons must be designed to be correspondingly long in relation to their diameter. This increases the volume of such a radial reciprocating engine, since the radial piston cylinder block must be designed with a diameter sufficient to provide the appropriate installation space for the cylinders.
Since the piston engines are intended by their design to exert pressures up to over 300 bar, the joint points between piston and slider shoe are exposed to high alternating loads. In particular, the joint itself can only withstand a specific maximum tensile force. If this force is exceeded, the joint is then disconnected, which leads to failure of the piston engine.
The object of the invention is therefore to design a radial reciprocating engine in such a way that the friction is reduced. A further object of the invention is to increase the permissible tensile force on the piston or to extend the service life of a piston before a defect occurs.
This object is achieved with a radial reciprocating engine of the type mentioned at the outset in that the piston element is spherical in form, at least in the region of the piston element, which seals the inner walls of the cylinder during stroke movements. The spherical design of the piston element results in a sealing region that is annular, i.e. that forms a closed circular line. A closed circular line causes far lower frictional forces than a flat seal by a cylindrical piston. During a rotation or tilting movement of the ball piston element, although the position of the circular sealing line on the surface of the at least partially ball piston changes, since however the diameter of the sealing circular line is constant due to its spherical shape and the inner diameter of the cylinder is also constant, this always results in exactly the same clearance between the cylinder inner wall and the ball piston element, irrespective of the position of the ball piston element in the cylinder and irrespective of the tilt angle of the ball piston element. Costly friction-reducing coatings or even tribological contours in the outer casing of a piston can thereby be dispensed with. The technical effort to produce a sufficiently perfect spherical shape is also far lower than is the case with a conventional longitudinal piston with comparable performance and with a sufficient level of perfection; or in other words, to achieve the same objective, the manufacturing quality of a ball piston element can be selected to be lower than that of a longitudinal piston according to the prior art.
However, since in the case of a sphere the diameter of a great circle is the same no matter in which direction the sphere is rotated, the piston element cannot jam in the cylinder during the stroke movement, because the diameter of the respectively sealing circular line remains unchanged in relation to the diameter of the cylinder. The ball piston element, at least in sections, is also independently centered in the middle due to the planar pressure. The losses of the piston engine and wear within the piston guide in the cylinder are thus reduced.
Since the spherical shape of the ball piston at least in sections reduces the overall length of the piston on the piston axis, the rotor diameter can be selected to be smaller with the same power, so that a more compact design of the radial reciprocating engine thereby results. At the same time, a higher working pressure can also be selected, because the weak point of the radial reciprocating engine is no longer determined by an articulated joint. The ball piston therefore eliminates the limitation of the housing internal pressure due to the tensile, compressive and transverse forces acting on a joint. Piston engines with longitudinal pistons resulted in limitations with regard to housing internal pressures or accumulation pressures in an external leakage oil line, which were caused, for example, by the elevated position of the hydraulic tank, filters or coolers in the leakage oil line. With the elimination of such a limitation, the range of application of piston engines expands in such applications or now makes certain applications possible.
Viewed mathematically, the surface of the sectionally ball piston that comes into contact with the inner walls of the cylinder is a symmetrical spherical zone. A spherical zone is the curved outer side of a spherical disc or a spherical ring, for example. A spherical disk, or also called a spherical layer, is obtained as the center part of a full sphere if the full sphere is cut into three parts by two parallel planes. If the parallel planes in this case lie on different sides of the sphere center point and at the same time are at the same distance from the sphere center point, this will be a symmetrical spherical disk whose outer surface results in a symmetrical spherical zone. However, a symmetrical spherical zone is also obtained by a radial hole through a sphere.
In a prototype of a radial piston pump according to the invention, a tilt angle α of the spherical zone of approximately 9° on both sides resulted due to the selected geometric conditions. With a corresponding tolerance, it would be advantageous to choose a tilt angle α of approximately 10°, better 12°. The ratio of the thickness of the spherical disk or height H of the spherical disk to the diameter d_Kof the spherical disk corresponds to twice the tangent function of the tilt angle α. At 12° on both sides, this then results in a ratio of the thickness H of the spherical disk to the diameter d_Kof the spherical disk of approximately 0.4. Of course, it is clear to the person skilled in the art that this may only be a point of reference. Depending on the geometric dimensions of the radial reciprocating engine, in individual cases even larger values for the tilt angle α may be required, or a small tilt angle α may also be sufficient. Tilt angles a of up to 20° seem to be technically reasonable or achievable.
The piston element, which is at least sectionally spherical, can advantageously be connected to the guide element via a connecting element, wherein the connecting element is rigidly connected to the piston and/or rigidly connected to the slide element. Due to the spherical section of the piston element, the ball piston element, insofar as its movement is not limited by positive guidance, or the connecting element does not come into contact with the cylinder walls, has all degrees of freedom to rotate in the cylinder in order, for example, to execute a pitching, yawing or rolling movement. In general, the movement of the piston element is limited by positive guidance to a pitching movement and a stroke movement. In most applications, it is thus no longer necessary to movably connect the slide element or the piston element to one another, since the required rotational movements can be carried out by the piston element formed at least in sections in the shape of a sphere. The losses of the piston engine and wear within the piston guide in the cylinder are thereby reduced.
The length and diameter of the connecting element will be selected by the person skilled in the art in accordance with to the desired kinematics and the desired tilt angle α. In particular in radial reciprocating engines, the slider shoe has a rounded contour that is matched to the swivel angle and the cylinder inner diameter. The contour of the slider shoe only has to be able to reproduce the design-related kinematic swiveling movement of the slider shoe.
The piston according to the invention with the at least sectionally spherical piston element can be produced, for example, by turning, milling or grinding. Alternatively, production, even of individual parts that are assembled accordingly, is possible by casting, additive manufacturing, MIM technology, sintering or even by the use of formed parts.
In particular application cases, the piston element can, of course, also be connected to the connecting element by a joint and/or the connecting element can be connected to the guide element by a joint.
The radial reciprocating engine can in this case be an internally pressurized radial reciprocating engine or an externally pressurized radial reciprocating engine. In the case of an internally pressurized radial reciprocating engine, the guide elements run on a thrust ring eccentrically arranged in relation to the radial axis of the cylinder carrier. In the case of the externally pressurized radial reciprocating engine, the guide elements run on an eccentric shaft inside the cylinder carrier that rotates eccentrically in relation to the center of the cylinder carrier.

The invention is now described and explained in more detail on the basis of exemplary embodiments depicted in the drawings. The figures show:

FIG. 1 a radial reciprocating engine with the ball pistons according to the invention

FIG. 2 a ball piston according to the invention

FIGS. 3A, 3B and 3C a section of a radial reciprocating engine

FIG. 4 an externally supported rotary piston engine with a ball piston according to the invention

The invention relates to a novel piston for a radial reciprocating engine, wherein the radial reciprocating engine, apart from the novel piston, can be a radial reciprocating engine according to the prior art. Due to the shape of the piston, this piston is referred to below as a ball piston, although the part of the ball piston that comes into contact with an inner wall of a hollow cylinder strictly speaking only has to correspond to a spherical section. In this sense, when for the sake of language simplification a spherical shape is spoken of below, a shape that is only spherical in sections should thus also be included.
FIG. 1 shows a simplified view of an internally pressurized and externally supported radial reciprocating engine 1, partially in section. In general, if no design features prevent this, such radial reciprocating engine can be operated both as a pump and as a motor. In the following, the operating mode as a pump is described as representative of the two operating modes. The principle of such a radial piston pump is known to the person skilled in the art, so that here only the essential parts are described insofar as they are necessary for understanding the invention. The radial piston pump 1 has a housing 10 that is approximately pot-shaped and is closed off by a housing cover (not shown). A thrust ring 12, which can be displaced in an adjustment direction 3, is mounted in the interior space 11 of the housing 10 and is mounted with its side faces between shoulders of the housing base and of the housing cover with a corresponding clearance. Due to the sectional view, only the side face 13 of the thrust ring 12 facing the observer is visible in FIG. 1, which side face comes to rest on the inside of the housing cover when the housing 10 is closed. The shoulders of the housing base are hidden by the thrust ring 12 in this illustration or are omitted for the sake of clarity.
The inner circumference of the thrust ring 12 forms a slide surface 14 for slider shoes 22 on which ball pistons 21 are supported, which are movably guided in radially running bores 5 of a cylinder carrier. Since, with this design, the cylinder carrier is set in rotary motion, the cylinder carrier is hereinafter referred to as rotor 16. Due to the interaction of these bores with the ball pistons 21, these bores are hereinafter referred to as piston bore 5.
The piston bores 5 are distributed rotationally symmetrically around the radial axis 17 of the rotor 16. The number of piston bores depends in part on the size of the rotor 16 or on the displacement volume of the radial piston pump 1. In the example shown here of a radial piston pump with a displacement volume of 19 cm³/U at a maximum operating pressure of 350 bar, seven piston bores 5 are provided in the rotor 16.
The rotor 16 is arranged on a control stud 18 fixedly arranged in a control stud hole in the housing 10 and is set in rotation by a drive shaft. In the operating mode as a motor, the torque generated by the ball pistons 21 is transferred to the drive shaft, which is now technically correctly referred to as the output shaft. The ball piston 21 here seals a working space of the radial reciprocating engine 1 against the interior space 11 of the radial reciprocating engine 1, wherein the working space extends within the piston bore 5 from the ball piston 21 in the extension of the piston bore in the direction of the control stud 18 and up to the control stud 18. In FIG. 1, the control stud hole and the drive shaft are hidden by the control stud 18 or are on the cutoff side of the diagram, and are therefore not visible.
Ball pistons 21 and slider shoe 22 are connected to one another by means of a piston rod 23. Here, the piston rod 23 can also be designed as a connecting rod, meaning that the connecting rod can be connected to the ball piston 21 so as to be movable via a joint arranged on the ball piston 21. Alternatively or additionally, the piston rod 23 or the connecting rod can be movably connected to the slider shoe 22 by a joint arranged on the slider shoe 22. However, a synergistic effect results in particular when the piston rod 23 is connected not only rigidly to the ball piston 21, but also rigidly to the slider shoe 22. In this case, the spherical shape of the ball piston 21 is used not only as a seal between the working space and the interior space 11 of the radial reciprocating engine 1, but also as a sole joint. In particular, this has the advantage that a weakening of the combination of ball piston 21, piston rod 23 and slider shoe 22 due to joints is avoided. In particular in the case of a rigid connection of the ball piston 21 and the slider shoe 22, the piston rod 23 can be designed as a tapering of the ball piston 21 in such a way that the piston rod achieves a high level of strength.
The spherical shape or the spherical section shape allows the ball piston 21 to perform a limited tilting movement in or counter to the direction of rotation of the rotor 16, in particular in its plane of rotation, doing so within the piston bore 5 and forcibly guided by the slider shoe 22. The tilting movement of the ball piston 21 is limited in this case by the piston rod 23, which can come to a stop at the piston bore walls, in particular the piston bore openings, which point towards the thrust ring 12. As a result of a bilateral positive guidance of the slider shoes 22, the ball piston 21 is not subjected to any lateral movements between ring strips (not shown) on the inside of the thrust ring 12.
For the purpose of changing the delivery rate, the thrust ring 12 can be displaced in the interior space 11 transversely to the control stud 18 by two adjustment pistons 31, 32 in the adjustment direction 3. At two diametrically opposite locations, the two adjustment pistons 31, 32 act on the external circumference of the thrust ring 12 with two sliding adjustment blocks 33, 34.
In the control stud 18, an eccentric first low-pressure channel 41 running longitudinally parallel to the center axis of the control stud 18 and a second low-pressure channel 42 are formed in each case for the supply of a pressure medium, which in each case open into a first circumferential groove, hereinafter referred to as a low-pressure slot 45, on the control stud 21. Furthermore, for the discharge of the pressure medium, an eccentric first high-pressure channel 43 running longitudinally parallel to the center axis of the control stud 21 and a second high-pressure channel 44 are formed in each case, each of which opens into a second circumferential groove, hereinafter referred to as a high-pressure slot 46, on the control stud 21. The low-pressure slot 45 and the high-pressure slot 46 are located in the outlet region of the piston bore 5 of the rotor 16 that accommodates the round pistons 21. The first low-pressure channel 41 and the second low-pressure channel 42 end in a 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 a high-pressure connection of the radial piston pump 1. Low-pressure connection and high-pressure connection of the radial piston pump 1 are not visible in this drawing, because, from the point of view of the observer, they are located on the rear side of the housing base. In this exemplary embodiment, two low- pressure channels 41, 42 and two high- pressure channels 43, 44 were selected, because this offers fluidic advantages in interaction with a known special geometric design of the control stud 18. However, in order to satisfy the basic principle of the radial reciprocating engine 1, a single low-pressure channel 41 and a single high-pressure channel 43 would already be sufficient.
During operation of the radial piston pump 1, i.e., when the rotor 16 is set in rotation, the ball pistons 21 in the piston bores 5 are taken along in the direction of the rotational movement. Due to the rotational movement of the rotor 16, a centrifugal force acts on the ball pistons 21 guided in the piston bores 5, whereby the ball pistons 21 are pressed radially outwards in the piston bores 5 until the slider shoes 22 of the ball piston 21 bear against the thrust ring 12 with their ends that face away from the control stud 18. In the following, the expression “outwards” refers to a direction that, starting from the axis of rotation 17 of the rotor 16, points away from the axis of rotation 17 of the rotor 16, while the expression “inwards” refers to a direction pointing in the direction of the axis of rotation 17 of the rotor 16. Similarly, the term “outside” refers to a relative position of an object at a greater radial distance from the axis of rotation 17 of the rotor 16 than an object at a radially smaller distance from the axis of rotation 17 of the rotor 16.
In the case where the position of the thrust ring 12 is set such that the imaginary axis of the thrust ring 12 is arranged identically to the rotational axis 17 of the rotor 16, the distance D between the outer side of the rotor 16 and the inner side of the thrust ring 12 is the same in each position of the rotor 16.
In this case, the radial distance of the ball pistons 21 with respect to the axis of rotation 17 does not change during a rotational movement of the rotor 16, so that the ball pistons 21 do not perform a stroke within the respective piston bore 5.
When the thrust ring 12 is adjusted so that the imaginary axis 19 of the thrust ring 12 is no longer identical to the rotational axis 17 of the rotor 16, the distance D between the rotor 16 and the thrust ring 12 is changed cyclically when the rotor 16 rotates within the thrust ring 12. This change in distance has the effect that, given a decreasing distance D between the thrust ring 12 and the rotor 16, the slide surface 14 of the thrust ring 12 exerts a counterforce on the slider shoes 22, which presses the ball pistons 21 inwards against the centrifugal force. In contrast, when the distance D between the thrust ring 12 and the rotor 16 increases, the thrust ring 12 relieves the load on the slider shoe 22 of the relevant ball piston 21 and the ball piston 21 is pressed radially outwards by the centrifugal force as well as by a compressive force, so that the relevant slider shoe 22 does not lose the contact with the slide surface 14 of the thrust ring 12. Accordingly, two periodic stroke movements are forced on the ball piston 21 during a full revolution of the rotor 16, a first stroke movement radially outwards, with which the volume of the working space continuously increases and which is therefore referred to below as the expansion phase, and a second stroke movement radially inwards in the direction of the control stud 18, with which the volume of the working space is continuously reduced and which is therefore referred to below as the compression phase.
The low-pressure slot 45 is arranged on the control stud 18 in such a way that in the expansion phase the low-pressure slot 45 and thus also the first and second low- pressure channels 41, 42 are in fluidic connection with the respective ball piston 21. The radially outward stroke movement of the ball piston 21 therefore generates a suction effect, which sucks in the pressure medium present at the low-pressure connection and fills the working chamber of the piston bore 5 with pressure medium. Furthermore, the high-pressure slot 46 is arranged on the control stud 21 in such a way that, during the compression phase, the high-pressure slot 46 and thus also the first and second high- pressure channels 43, 44 are fluidically connected to the respective ball piston 21. The radially inward stroke movement of the ball piston 21 therefore generates a pressure effect, which pushes the pressure medium accumulated in the working space of the relevant piston bore 5 into the high- pressure channels 43, 44 via the high-pressure slot 46.
In the operating mode as a pump, the cyclic stroke movement of the ball pistons 21 thus forces a pressure medium stream from the low-pressure channel to the high-pressure channel. In contrast, in the operating mode as a motor, a pressure-medium stream from the high-pressure to the low-pressure channel takes place depending on the operating mode of the radial reciprocating engine 1 as a pump or as a motor, a pressure energy is given to the radial reciprocating engine during pump operating mode by a drive torque, or in motor operating mode a pressure energy is withdrawn and converted into an output torque.
FIG. 2 shows a piston element 2 that comprises the ball piston 21 and the slider shoe 22, which are rigidly connected to one another by means of the piston rod 23. Through a longitudinal axis of the ball piston 21, of the piston axis 20 a bore 29 runs through which pressure medium in the piston bore 5 is pressed onto a slider shoe sole 26 during operation of the piston machine 1. The slider shoe sole 26 is thereby forced into a hydrostatic equilibrium, in which a friction-reducing film of pressure medium forms between the sliding sole and sliding surface 14. In a radial reciprocating engine 1, the slider shoe sole 26 is matched to the geometry of the slide surface 14; that is, it is outwardly curved.
Furthermore, as shown in FIG. 3, the slider shoe 22, viewed in the direction of rotation relative to its piston axis 20, has a slider shoe leading end 27 and, opposite thereto, a slider shoe trailing end 28.
In principle, apart from the piston rod 23, the ball piston 21 could assume a fully spherical shape. However, an ideally perfect spherical shape is required to come into contact with the piston bore wall 51 only at the points of the ball piston 21 in order to seal the piston bore space or, more precisely, as good as come into contact with it. The great circle of the ball piston 21, which is perpendicular to the piston axis 20, is referred to below as the center circle 24. If the piston axis 20 coincides with the piston bore axis 50, the extension of the plane spanned by the center circle 24 will, beyond the center circle 24, intersect the piston bore in an intersection circle, which is hereinafter referred to as a sealing circle, because it effectively closes off the working space of the piston bore 5 with respect to the interior space 11 of the piston machine 1. The sum of all sealing circles that can form in the case of a full rotation here defines a spherical zone 250, i.e., the outer circumference of a spherical disk in which the at least partially spherical piston element 71 must correspond to an ideal spherical shape.
In order to prevent the ball piston 21 from jamming in the piston bore 5, the diameter d_Kof the center circle 24 is selected to be slightly smaller than the diameter of the piston bore 5. With a center circle 24 smaller by, for example, 10 μm, when the center circle 24 and the piston bore wall 51 are in a centric position, a clearance of 5 μm all round between working circle and center circle results. Due to the viscosity of the pressure medium this results in an adequate seal between the working space and the interior space 11 of the piston machine 1. Of course, the person skilled in the art will select the clearance between the ball piston 21 and the piston bore wall 51, such that under the given dimensions and the respective intended purpose it is optimally suitable.
FIG. 3A shows the ball piston 21 at its outer dead center, i.e., during the transition from the expansion phase into the compression phase. At the outer dead center, the distance between the rotor 16 and the slide surface 14 of the thrust ring 12 is at a maximum distance Dmax due to the eccentric position of the thrust ring 12. Due to centrifugal force, or ring strips, hold-down elements and other compressive forces, the ball piston 21 generally aligns itself at its outer dead center in such a way 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 engine 1 are selected such that, here, the center circle 24 of the ball piston 21 is located close to the outer piston bore opening 52, but still sufficiently deep in the piston bore wall 51 to ensure a secure guidance of the piston 2 within the piston bore 5 along with a seal between the piston bore wall 51 and the center circle 24.
During further rotation from the outer dead center, the piston 2 is then taken along in the direction of rotation 15 of the rotor 16 by the forces acting on the center circle 24 of the ball piston 21 from the piston bore wall 51. Since, in the compression phase, the distance between the thrust ring 12 and the rotor 16 is increasingly reduced, the piston 2 is guided deeper inwardly into the relevant piston bore 5. Due to the eccentricity between thrust ring 12 and rotor 16, the distance between thrust ring 12 and rotor 16 at the position of the slider shoe leading end 27 is different from the distance between thrust ring 12 and rotor 16 at the position of the slider shoe trailing end 27. This builds up a force, which in the compression phase acts from the slider shoe leading end 27 to the slider shoe trailing end 28. As a result, in the compression phase the piston 2 of the ball piston 21 is tilted about its center point Z opposite to the direction of rotation 15 of the rotor 16. As a result of this tilting movement, the center circle 24 on the piston bore wall 51 facing away from the rotational direction swivels in the direction of the outer piston bore opening 52 and on the piston bore wall 51 facing the rotational direction swivels in the direction of the inner piston bore opening 53. The center circle 24 of the ball piston 21 thereby loses contact with the piston bore wall 51. Due to the spherical shape of the ball piston 21, a new great circle of the ball piston 21 now becomes the sealing circle 25, namely the great circle of which comes to rest in a respective position of the ball piston 21 in the piston bore 5 perpendicular to the piston bore axis 50. As a geometric consequence, the center point of each sealing circle is identical to the spherical center point Z of the ball piston 21.
During the transition from inner dead center IT to outer dead center AT the piston 2, i.e., in the expansion phase, the distance between the outside of the rotor 16 and the inside of the thrust ring 12 increases continuously from Dmin to Dmax. In the expansion phase, the distance D at the slider shoe leading end 27 is therefore always greater than at the slider shoe trailing end 28. As a result, the counter-force component exerted on the piston 2 by the thrust ring 12 is lower at the slider shoe leading end 27 than at the slider shoe trailing end 28. Consequently, the piston 2 moves away in the direction of the slider shoe leading end 27 than in the direction of the direction of rotation 15 of the rotor 16. Viewed in this way, the tilt angle α leads ahead of the rotational movement 15 in the expansion phase. In the compression phase, that is, during the transition between the outer dead center AT to the inner dead center IT, these relationships reverse and the tilt angle α trails behind the rotational movement 15 of the rotor 16. As a result of this tilting movement, the piston axis 20 is tilted by an angle α with respect to the piston bore axis 50. This tilting movement has its greatest extent on the one hand, as shown in FIG. 3B, approximately at about halfway of the movement of the piston 2 between inner dead center IT and outer dead center AT, or approximately at about halfway of the movement of the piston 2 between outer dead center AT and inner dead center IT.
Finally, FIG. 3C shows the ball piston 2 at its inner dead center when the distance between thrust ring 12 and rotor 16 has reached a minimum distance D_min. As can be seen, a piston base 210, which adjoins the piston 21, which is designed to be spherical in sections, on the side facing away from the piston rod 23, is designed to be essentially frustum-shaped, so that the piston base 210, at the inner dead center IT, can largely adapt to a funnel-shaped transition from the piston bore 5 to the control stud 18. This advantageously minimizes the dead volume in the inner dead center IT.
According to geometric laws, this angle α will also be equal to the angle α between the sealing circle plane and the center circle plane. For manufacturing reasons, but also because of a possible change in direction of rotation, the spherical surface of the ball piston 21 is designed symmetrical such that the spherical section or the spherical zone 250, as shown in FIG. 2, comprises at least one angle between −α and +α. Furthermore, depending on the selected geometry of the radial reciprocating engine 1 and the piston 2, the taper at the position of the piston rod 23 is to be selected such that, during operation, the piston rod does not contact the piston bore walls 51 or the rotor 16 in order to avoid damage to the piston bore rod 23 or the rotor 16 and piston bore walls 51.
The invention is also suitable for internally supported radial reciprocating engines. FIG. 4 shows the basic structure of an internally supported radial reciprocating engine 6, with which low-pressure channels and high-pressure channels are arranged on the outside 66 of one, in this case stationary, piston carrier 61. The design of thelow-pressure channelsand high-pressure channels in an internally supported radial reciprocating engine 6 are familiar to the person skilled in the art, so that, for the sake of clarity, they are not explicitly shown in the drawing FIG. 4.
The piston carrier 61 forms a receptacle for a plurality of piston bores 5, which are arranged radially around a center point 60 of the piston carrier 61 at an equal 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 carrier 61. For the sake of clarity, FIG. 4 shows only three piston bores 5. An odd number between three and nine is customary, but the person skilled in the art will select a suitable 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 an eccentric shaft 63 rotates around the piston carrier center 60. The eccentric center point E is located at a distance D from the piston carrier center point 60, as a result of which the eccentric center point E rotates around the piston carrier center point 60 on a circular path 64.
A ball piston according to the invention is arranged in each of the piston bores 5 and is essentially formed by a spherical piston element 71, a piston rod 73 and a slider shoe 72. Each round piston is pressed in the direction of the eccentric shaft 63 or of the eccentric center point E, i.e. into the interior of the radial reciprocating engine, by a restoring element, which in this exemplary embodiment are designed as coil springs 74. In the case of the internally supported radial reciprocating engine 6, the at least partially spherical piston element 71 seals off the working space of the radial reciprocating engine 6 to the inside. The working space in which the pressure medium is supplied or discharged is thus the 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 the eccentric shaft 63 forms a slide surface 65 on which the slider shoes 72 of the ball pistons 7 are supported with slide soles 76. Since the slide surface 65 of the eccentric shaft is convex in this exemplary embodiment, the slider shoes 76 are correspondingly concave. During a full rotation of the eccentric shaft 63, the ball piston 7 is forced through the sliding of the slider shoes 72 on the eccentric shaft 63 into two periodic stroke movements, a first stroke movement radially outwards, in which the volume of the working space is continuously reduced and which is therefore referred to below as the compression phase, and a second stroke movement radially inwards, in the direction of the piston carrier center point 60, with which the volume of the working space continuously increases, and which is therefore referred to below as an expansion phase. The distance D between the eccentric center point E and the piston carrier center point 60 determines the amplitude of the piston stroke.
As in the case of the externally supported radial reciprocating engine 1 described in the first exemplary embodiment, the ball pistons 7 perform a tilting movement during the expansion and compression phases. As a result of the at least partially spherical piston element 71, a circular sealing circle is also maintained here at all times between an inner wall 51 of the piston bore 5 and a spherical zone 75 of the ball piston 7. In this exemplary embodiment as well, a clearance between the ball piston 7 and the piston bore wall 51 prevents the ball piston 7 from jamming.

Claims

1. Radial reciprocating engine comprising:

cylinders arranged in a cylinder carrier and a ball piston arranged in each cylinder;

the ball piston rigidly connected to a guide element via a connecting element and the guide element rigidly connected to the connecting element;

wherein the guide element runs on a slide surface;

wherein a stroke movement is imposed on the ball piston;

wherein the ball piston is spherical at least in a region of the ball piston which seals an inner walls of the cylinder during the stroke movement; and

wherein the region of the ball piston which seals the inner wall of the cylinder during the stroke movement comprises a symmetrical spherical zone.

2. (canceled)

3. (canceled)

4. Radial reciprocating engine according to claim 1, wherein the guide element is a slider shoe, which runs over the slide surface.

5. Radial reciprocating engine according to claim 1, wherein the radial reciprocating engine is an externally supported radial reciprocating engine.

6. Radial reciprocating engine according to claim 5, wherein the guide elements runs on a thrust ring.

7. Radial reciprocating engine according to claim 1, wherein the radial reciprocating engine is an internally supported radial reciprocating engine.

8. Radial reciprocating engine according to claim 7, wherein the guide elements runs on an eccentric shaft.

9. (canceled)

10. Radial reciprocating engine according to claim 1, wherein the symmetrical spherical zone comprises in each case at least one tilting range of 20° towards both sides.