A piston that seals without contact, having a piston rod.
The piston compressor design with a non-contact gap seal and floating support system addresses wear and sealing issues in high-pressure applications, ensuring reliable operation and extended service life by minimizing contact forces and maintaining a stable sealing gap.
Patent Information
- Application Number
- JP2025574205
- Authority / Receiving Office
- JP · JP
- Patent Type
- Applications
- Current Assignee / Owner
- Priority Date
- 2023-06-20
- Filing Date
- 2024-06-14
- Publication Date
- 2026-07-13
AI Technical Summary
Piston compressors face challenges in high-pressure applications due to rapid wear and limited service life of piston rings, especially in unlubricated conditions, leading to increased friction and wear, and the inability of non-contact gap seals to maintain a stable sealing gap without metal-to-metal contact.
A piston design with a non-contact gap seal utilizing a double radial floating support and piston sleeve, coupled with a radially floating piston rod, minimizes contact forces through gas dynamic forces and pressure compensation, allowing for a small sealing gap and effective sealing even at high pressures up to 1000 bar.
The solution enables reliable operation of piston compressors in high-pressure applications by reducing wear and maintaining a stable sealing gap, extending the service life and preventing damage to components, while avoiding contamination of the compression medium.
Smart Images
Figure 2026523167000001_ABST
Abstract
Description
Technical Field
[0001] The present invention relates to a piston for a piston compressor that seals in a non-contact manner having a piston rod, and to a method of using such a piston in a cylinder of a piston compressor.
Background Art
[0002] As is well known, in a piston compressor, a piston reciprocates in a cylinder chamber of a cylinder to compress a compression medium in the cylinder chamber. In order to enable sufficient and efficient compression of the compression medium in the cylinder chamber, a sealing member is required between the movable piston and the cylinder or the cylinder sliding surface. A normal sealing member is a contact sealing material that is inserted into a groove of the piston and abuts against the cylinder sliding surface, such as a well-known piston ring. The piston ring has the role of sealing the cylinder chamber and suppressing leakage of the compression medium through the piston as little as possible. Usually, a plurality of piston rings are provided on the piston.
[0003] The piston ring, and also the cylinder sliding surface, are subject to friction and wear that limit the service life of the piston ring, particularly based on contact, piston speed, and the acting lateral force. Therefore, the piston ring is a maintenance part that must be replaced regularly.
[0004] Almost all current piston compressors use piston rings made of polymer-based materials. These are particularly high-performance polymers that are modified with fillers regarding tribological properties so that friction and wear are minimized. Accordingly, the service life of the piston ring can be extended.
[0005] In piston compressors, a distinction can also be made between oil-lubricated and oil-free applications. In oil-lubricated applications, a thin oil film forms a boundary layer between the pressure-biased piston ring and the cylinder sliding surface. This significantly reduces friction and wear compared to unlubricated applications. Furthermore, the oil film also provides a sealing effect. However, in many applications, oil-lubricated piston rings are not feasible because the lubricating oil can contaminate the compression medium, which is often undesirable and unacceptable.
[0006] In unlubricated, oil-free applications, tribology (friction and wear) is defined solely by the properties of the compression medium, operating conditions such as average piston speed, acting lateral and other forces, the material and coating of the cylinder sliding surface, and the piston ring material (e.g., polymer material and its fillers). Furthermore, the formation of a transition film between the piston ring material and the cylinder sliding surface is also a factor.
[0007] For unlubricated high-pressure applications with final compression pressures exceeding 250 bar, even typical high-performance polymers reach their material limits, leading to disproportionately increased piston ring friction and wear. While appropriate modifications to the piston ring design can reduce localized contact pressure between the piston ring and the cylinder sliding surface, this only allows for a possible increase in final compression pressure up to approximately 350 bar, even in unlubricated applications. In addition, the piston rings wear out too rapidly, resulting in a significantly shorter achievable service life, rendering them unsuitable.
[0008] To achieve a useful service life for industrial applications, high-pressure pistons for piston compressors (for final compression pressures <350 bar) are equipped with numerous piston rings. In such cases, 10 to 20 piston rings may be provided. However, this leads to a very large overall length for the piston, and consequently for the piston compressor, as well as a reduction in the heat dissipation of individual piston rings, which in turn limits the service life.
[0009] Each piston ring is inserted into a groove in the piston. For high-pressure applications, an additional metal support ring is inserted behind the piston ring (on the side opposite to the pressure) to support it against the cylinder pressure acting through the piston ring. Such a support ring prevents the piston ring from being pushed out into the gap between the cylinder sliding surface and the piston diameter due to the pressure.
[0010] Pistons with piston rings and support rings for high-pressure applications are known, for example, from International Publication No. 2010 / 084071. It also describes the use of support rings in the piston to guide and center the piston within the cylinder.
[0011] The typical gap width between the support ring and the cylinder sliding surface varies within a range of several tens of millimeters. Smaller gap widths are not used in industry. This is because tolerances and lateral forces transmitted to the piston through the crank mechanism would create a risk of metal-to-metal contact between the cylinder sliding surface and the support ring. Metal-to-metal contact leads to high wear and can damage the cylinder sliding surface.
[0012] A well-known alternative sealing concept for sealing between the piston and cylinder sliding surface is a non-contact sealing material, such as a labyrinth seal or gap seal. In this case, the sealing action is achieved through a cylindrical sealing gap between the outer diameter of the piston and the cylinder sliding surface, and this sealing gap, in the embodiment as a labyrinth seal, generates an additional throttling effect through additional grooves. To achieve sufficient sealing action in high-pressure applications of >350 bar, such a sealing gap must be in the range of <50 μm. Such a very small sealing gap presents a series of difficulties.
[0013] Due to manufacturing reasons, the non-uniformity of the piston and / or cylinder sliding surfaces, manufacturing tolerances of the piston and cylinder sliding surfaces, and lateral forces acting from the crank mechanism almost inevitably lead to metal-to-metal contact between the piston and cylinder sliding surfaces, given a gap height of <50 μm in the sealing gap. This leads to significant wear and damage to each component of the piston compressor, even after a short period of operation.
[0014] Pistons and cylinders are typically made of steel due to the high loads they are subjected to. Uneven temperature distribution between the piston and cylinder leads to changes in the effective gap height of the sealing gap due to thermal expansion. This can lead to contact between the piston and cylinder, as the effective gap height approaches zero, or to disproportionate leakage, as the gap height increases, even under the low temperature differences that typically occur in piston compressors.
[0015] In particular, in high-pressure applications where the final compression pressure is >350 bar, deformation of the cylinder and / or piston due to pressure can occur, which can lead to changes in gap height in the range of 10 μm. This, too, can lead to similar negative effects, namely increased contact and gap height.
[0016] In particular, micrometer-sized particles carried with the compression medium may enter the sealing gap, which can lead to piston jamming or damage to the piston surface and / or cylinder sliding surfaces.
[0017] For these reasons, although gap seals for piston compressors are theoretically possible, no industrial implementation of such gap seals in piston compressors is known.
[0018] German Patent Application Publication No. 2160816 describes a piston for a dry piston compressor. A piston ring made of plastic is supported on the piston radially on the catch surface, thereby acting as an approximately wear-free labyrinth seal with a narrow gap against the cylinder wall. Straight-line guidance of the piston is ensured by the guide ring. However, even a slight tilt of the piston within the cylinder causes further wear on the piston ring, widening the gap and thus degrading the sealing effect.
[0019] International Publication No. 98 / 31936 describes a non-lubricated piston compressor comprising a piston having a metal cylinder sliding surface and a casing made of plastic material, which together form a gap ring seal. Parallel movement of the piston within the cylinder is possible via a piston rod support that allows the piston rod to pivot.
[0020] The problems mentioned above regarding piston rings and gap seals have traditionally hindered the industrial use of piston compressors for unlubricated high-pressure applications, particularly under final compression pressures of >350 bar. For such high-pressure applications, other compressor types have traditionally been employed, such as diaphragm compressors and ion compressors. In diaphragm compressors, the short lifespan of the diaphragm is a problem. In ion compressors, mixing of the ionic liquid and the compression medium can occur, which is generally undesirable. Even disregarding that, lifespan is still a problem in ion compressors.
[0021] Using a piston compressor in high-pressure applications is preferable because it is a well-established, particularly simple, type of compressor. [Overview of the project] [Problems that the invention aims to solve]
[0022] Therefore, the object of the invention in question is to provide a piston for a piston compressor that enables the use of the piston compressor for high-pressure applications.
Means for Solving the Problem
[0023] This is made possible by the piston that seals in a non-contact manner according to claim 1. By means of the piston manufactured based on the present invention, a non-contact gap seal can be realized. Based on an embodiment, in particular, based on a double radial floating support including a first floating support of the piston sleeve and a second floating support of the piston on the piston rod, the gas dynamic force generated between the piston and the cylinder reduces the force required to guide the piston to such an extent that the contact between the piston and the cylinder is minimized. Thereby, it can be operated while maintaining a very small sealing gap. At the same time, pressure compensation of the piston sleeve is generated by the axial gap and the sleeve gap, whereby the piston sleeve does not deform or hardly deforms based on the acting pressure, and accordingly the sealing gap is not impaired either. Accordingly, the piston having a piston rod according to the present invention is particularly applicable for high-pressure applications where the final compression pressure is up to 1000 bar or higher. Such a final compression pressure can particularly occur in the use of hydrogen. Accordingly, it becomes possible for the first time to use a piston compressor for such a final compression pressure.
[0024] In an alternative embodiment, the piston body may be provided with a plurality of sleeve grooves axially spaced in the direction of the piston axis, and the piston sleeve is arranged to be radially floatingly supported in each sleeve groove.
[0025] If at least one throttle groove is provided on the outer circumferential surface of the piston body, preferably, if a plurality of throttle grooves spaced apart from each other in the axial direction are provided, particles that may exist between the outer circumferential surface of the piston body and the cylinder sliding surface can be captured. Along with this, it is possible to prevent the occurrence of scratches on the cylinder sliding surface and / or the outer circumferential surface of the piston body. Furthermore, the sealing effect can also be improved by such throttle grooves.
[0026] If the axial end face of the piston rod and / or the bottom surface of the piston rod notch are manufactured to be convexly curved, it is possible to compensate for the angular offset that may occur between the piston and the cylinder, and thus prevent the high contact force caused by this.
[0027] If the piston rod is elastically coupled to the piston body via a radially or axially elastic coupling member within the piston notch, it is possible to prevent the piston body from rising from the piston rod when starting the piston compressor without impairing the floating support of the piston body. The same can be achieved by an initial stress unit that actively presses the piston body toward the piston rod.
[0028] Next, by way of example, schematically and without limitation, the subject invention will be described in detail while referring to FIGS. 1 to 9 showing preferred embodiments of the present invention. The drawings show the following.
Brief Description of the Drawings
[0029] [Figure 1] Showing a piston compressor. [Figure 2] Showing a piston for non-contact sealing according to the present invention. [Figure 3] Showing details of a piston for non-contact sealing according to the present invention. [Figure 4] Showing a radially movable support portion of the piston on the piston rod. [Figure 5]This shows the support portion that moves radially in the piston direction on the piston rod. [Figure 6] This shows the support portion that moves radially in the piston direction on the piston rod. [Figure 7] This shows the support portion that moves radially in the piston direction on the piston rod. [Figure 8] This shows the support portion that moves radially in the piston direction on the piston rod. [Figure 9] This shows the cylinder of a piston compressor having a piston according to the present invention. [Modes for carrying out the invention]
[0030] The basic operating principle of the piston compressor 1 is well known, but it will be briefly explained below with reference to Figure 1. Figure 1 shows a cross-sectional view of the piston compressor 1 in one preferred embodiment. The piston compressor 1 has a compressor housing 2, which contains at least one cylinder Z1, within which a piston 3 is movable to reciprocate. The piston 3 reciprocates within the cylinder Z1 between top dead center OT and bottom dead center UT. The piston 3 partially divides the cylinder chamber K1 within the cylinder Z1, within which a compressible medium, such as hydrogen, is compressed by the movement of the piston 3. The piston 3 is coupled to a reciprocating piston rod 4 and is driven by the piston rod 4 during operation. The end of the piston rod 4 opposite the piston 3 is coupled to a piston drive 12, which generates the reciprocating motion of the piston rod 4 and the piston 3.
[0031] In the embodiment shown in Figure 1, the piston drive unit 12 is manufactured as a crank mechanism having a crosshead 5, but the piston drive unit 12 may also be manufactured to have, for example, a cam mechanism. In this embodiment, the piston rod 4 is coupled at the end facing the piston 3 to a crosshead 5 supported by the compressor housing 2 so as to be movable in the axial direction. The crosshead 5 is driven to cause the piston rod 4 and the piston 3 to reciprocate. In the illustrated embodiment, the crosshead 5 is driven by a crank mechanism. For this purpose, the crosshead 5 is pivotally coupled to one end of a thrust rod 6, and the other end of the thrust rod 6 is pivotally coupled to the crankpin 7a of the crankshaft 7. The crankshaft 7 rotates when the piston compressor 1 is in operation. The crankshaft 7 is driven by a drive device not shown in Figure 1, for example, by an electromechanical device. Through the thrust rod 6, the rotational motion of the crankshaft 7 is converted into the translational motion of the crosshead 5 when the compressor 1 is in operation. The lateral force generated at that time is supported by the compressor housing 2 via the crosshead 5. This causes the piston rod 4 to perform a substantially purely translational reciprocating motion in the direction of the piston axis KA. However, the piston 3 may, of course, be driven by a method other than using a crank mechanism.
[0032] The reciprocating piston rod 4 is typically inserted into a sealing unit 9, which is also located within the compressor housing 2. The sealing unit 9 seals the piston drive housing 8, i.e., the crank housing in Figure 1, in which the crosshead 5 also moves, to prevent the compressed medium from entering the piston drive housing 8. The sealing unit 9 is preferably manufactured as a well-known packing having multiple seal rings and optionally other rings, such as scraper rings, support rings, etc.
[0033] An embodiment of the piston compressor 1 having a piston rod 4 that performs reciprocating motion can also be manufactured as a dual-action piston compressor 1 in the embodiment shown in Figure 1, so that the piston compressor 1 can perform dual action as needed. In the dual-action piston compressor 1, a second cylinder Z2 having a second cylinder chamber K2 is configured on the side of the piston 3 facing at least one first cylinder chamber K1, and the piston 3 also partially partitions this second cylinder chamber K2. When the piston 3 moves back and forth, the cylinder volume of the first cylinder Z1 decreases (compression stroke), and at the same time the cylinder volume of the second cylinder Z2 increases (suction stroke), or vice versa.
[0034] The illustrated embodiment of a dual-action compressor should, of course, be understood as illustrative only. In the simplest case, the piston compressor 1 has only one cylinder Z1, separated by the end face of the piston 3 facing away from the piston drive 12.
[0035] The cylinders Z1 and Z2 of the piston compressor 1 are also equipped with valves SV1, DV1, SV2, and DV2 in a well-known manner, with at least one intake valve SV1, SV2 and at least one pressure valve DV1, DV2 in each cylinder Z1 and Z2. These valves may be manufactured as automatic valves, that is, controlled by the acting cylinder pressure. Alternatively, the valves may be actively controlled, for example, by a well-known unloader.
[0036] The embodiment shown in Figure 1 illustrates a dual-action piston compressor 1 in which both suction valves SV1 and SV2 are connected to a suction pipe SL through which the compression medium KM is drawn, and both pressure valves DV1 and DV2 are connected to a pressure pipe DL through which the compressed compression medium KM is discharged.
[0037] A multi-stage piston compressor 1 can also be realized if the piston surface on one side of piston 3 is manufactured to be smaller than the piston surface on the opposite side of piston 3. In this case, the suction piping of the first compressor stage is connected only to the suction valve SV1 of the first compressor stage. The pressure piping DL1 of the first compressor stage, which is connected to the pressure valve DV1 of the first compressor stage, is also the suction piping SL of the second compressor stage, and is therefore connected to the suction valve SV2 of the second compressor stage. The pressure valve DV2 of the second compressor stage will be connected to the pressure piping DL2 of the second compressor stage.
[0038] In the embodiment shown in Figure 1, piston rings 15 are positioned on the piston 3 to seal the respective cylinder chambers K1 and K2, and to contact the cylinder sliding surfaces 10 of the cylinders Z1 and Z2.
[0039] Figure 2 shows an embodiment of a non-contact sealing piston 3 according to the present invention, along with a portion of the piston rod 4 of a piston compressor 1 for high-pressure applications, in use in the cylinder Z1 of the piston compressor 1. A non-contact gap seal is provided to seal between the piston 3 and the cylinder Z1. As is common in piston compressors 1, the piston 3 reciprocates within the cylinder chamber K1 of the cylinder Z1. The piston drive 12 of the piston 3 is not important to the present invention and is therefore not shown in Figure 2 except for the reciprocating piston rod 4. Spacings and gaps are greatly exaggerated in Figure 2 for illustrative purposes. In the case of the gap seal, the piston 3 is manufactured to act only as a single unit, so the piston 3 only partitions the cylinder chamber K1.
[0040] The piston 3 according to the present invention consists of a piston body 20 and at least one hollow cylindrical piston sleeve 21. The piston body 20 is provided with at least one sleeve groove 24 in the form of a circumferential groove, and at least one piston sleeve 21 is positioned therein, supported in a floating manner laterally with respect to the piston shaft KA. Here, "supported in a floating manner" means that the piston sleeve 21 can move freely (excluding any frictional forces acting on it) radially laterally with respect to the piston shaft KA within the sleeve groove 24 with respect to the piston body 20.
[0041] The floating support of the piston sleeve 21 to the piston body 20 is achieved by providing a radial sleeve gap HS between the inner circumferential surface 22 of at least one piston sleeve 21 and the outer circumferential surface of the sleeve groove 24, and by making the axial sleeve length HL of at least one piston sleeve 21 shorter than the axial groove length NL of at least one sleeve groove 24 in the direction of the piston axis KA.
[0042] When the piston 3 is used in the piston compressor 1, a sealing gap DS is formed between the outer circumferential surface 27 of at least one piston sleeve 21 and the cylinder sliding surface 10 of the cylinder Z1 in which the piston 3 moves during use, thereby achieving a gap seal. Consequently, the maximum outer diameter of the piston body 20 in the region of the cylinder sliding surface 10 is smaller than the inner diameter of the cylinder sliding surface 10 and also smaller than the outer diameter of the outer circumferential surface 27 of at least one piston sleeve 21.
[0043] The radial freedom of movement of the piston sleeve 21 within the piston body 20 is, of course, limited to the design-defined range of motion margin, which in the case of piston 3 is substantially defined by the sleeve gap HS. However, the range of motion possible when using piston 3 within cylinder Z1 of piston compressor 1 is, of course, defined by the sealing gap DS. This range of motion is also guaranteed when using piston 3 within piston compressor 1.
[0044] To enable the manufacture of the piston 3, the piston body 20 may be manufactured in multiple parts and may consist of multiple piston body parts which are assembled to form the piston body 20. In the embodiment shown in Figure 2, the piston body 20 consists of a piston center 25 and a piston end 26 attached to the piston center 25, for example, screwed to the outside or inside. The piston end 26 faces the cylinder chamber K1 when the piston 3 is used in the piston compressor 1. The outer diameter of the piston center 25 is smaller than the outer diameter of the piston end 26 along at least one specific axial groove length NL in the direction of the piston axis KA. Accordingly, at least one region of the piston center 25 having a smaller outer diameter forms at least one sleeve groove 24 having groove length NL, in which the piston sleeve 21 is supported and positioned in a floating manner.
[0045] To achieve floating support for the piston sleeve 21 on the piston body 20, the inner diameter of the piston sleeve 21 is larger than the outer diameter of the sleeve groove 24 on the piston body 20. As a result, a sleeve gap HS is created between the inner circumferential surface 22 of the piston sleeve 21 and the outer circumferential surface 23 of the sleeve groove 24. The sleeve gap HS is preferably between 0.1 mm and 1 mm in width.
[0046] In addition, the axial sleeve length HL of at least one piston sleeve 21 is smaller than the groove length NL of the sleeve groove 24 in which the piston sleeve 21 is floatingly supported and positioned in the direction of the piston axis KA, thereby creating an axial gap AS between the piston sleeve 21 and the piston body 20. The axial gap AS is important for the floating support of the piston sleeve 21 in order to avoid pinching of the piston sleeve 21 in the axial direction. The size of the axial gap AS should be appropriately selected, but it should not be too large, because an axial gap AS will create undesirable dead space.
[0047] In the direction of the piston axis KA, the piston 3 acts on the end face of the piston sleeve 21 facing the cylinder Z1 during use, based on the axial gap AS, and the cylinder pressure p acts on it. Z Based on this, it is pressed against the end of the sleeve groove 24 facing the cylinder Z1. However, the end face of the piston sleeve 21 is small compared to the cylinder cross-section, and therefore the axial force acting on it is small. This reduces the frictional force that is generated and should not restrict, or at least not significantly restrict, the radial movement of the piston sleeve 21 laterally relative to the piston axis KA. In any case, the floating support of the piston sleeve 21 laterally is not affected in terms of function. The contact surface of the piston sleeve 21 on the piston body 20 is preferably selected to be as wide as possible in order to minimize the contact pressure between these two surfaces and the resulting deformation. A suitable measure for pressure compensation between these two surfaces may lead to a further reduction in the contact force. Accordingly, the sleeve gap HS and axial gap AS between the piston sleeve 21 and the piston body 20 ensure the floating support of the piston sleeve 21 radially on the piston body 20. Consequently, the piston sleeve 21 can compensate for small (maximum within the sleeve gap range) lateral deviations of HS during the operation of the piston 3 in the cylinder Z1 of the piston compressor 1, so that no significant lateral force is generated between the piston sleeve 21 and the cylinder sliding surface 10. This reduces the wear that may occur in cases where unavoidable contact occurs between the piston sleeve 21 and the cylinder sliding surface 10. Even so, the sleeve gap HS is 0.1 mm to 1 mm, which is small enough not to create much additional dead space (space in the cylinder Z1 that cannot be used for compression). As is well known, too much dead space reduces the output capacity of the piston compressor 1 and is therefore undesirable.
[0048] The sleeve gap HS and the axial gap AS are determined by the cylinder pressure p during the operation of the piston 3 of the piston compressor 1. ZThis also ensures that pressure compensation is possible between the outer circumferential surface 27 and the inner circumferential surface 22 of the piston sleeve 21, as the pressure from the cylinder chamber K1 reaches the sleeve gap HS between the piston sleeve 21 and the piston body 20. Such pressure compensation can reduce the deformation of the piston 3 based on the acting pressure, and, associated with this, the change in the sealing gap DS between the outer circumferential surface 27 of the piston sleeve 21 and the cylinder sliding surface 10.
[0049] Acting cylinder pressure p Z This also acts to center the piston sleeve 21 within the cylinder Z1. Based on the hydrodynamic effect, pressure is generated on the side where the sealing gap DS between the piston sleeve 21 and the cylinder sliding surface narrows, which in turn acts to move the piston sleeve 21 back to its central position.
[0050] To achieve an effective gap seal, the sealing gap DS between the outer circumferential surface 27 of the piston sleeve 21 and the cylinder sliding surface 10 is between 1 μm and 50 μm in size. Preferably, the outer circumferential surface 27 of the piston sleeve 21 performs the sealing function based on the sealing gap DS formed along the entire sleeve length HL.
[0051] Figure 3 shows an embodiment of a piston 3 having multiple piston sleeves 21. Unlike the embodiment shown in Figure 2, where a single piston sleeve 21 is supported radially floating within the sleeve groove 24, here multiple piston sleeves 21, each radially floating within the sleeve groove 24, are used. Consequently, the original sealing surface for the non-contact piston seal, i.e., the outer circumferential surface of the piston sleeve 21, is divided into multiple sealing surfaces.
[0052] In this embodiment, each piston sleeve 21 is provided with a sleeve groove 24, in which one piston sleeve 21 is radially supported and positioned, having a sleeve gap HS (which may differ) between the inner circumferential surface 22 of the piston sleeve 21 and the outer circumferential surface 23 of each corresponding sleeve groove 24, and the axial sleeve length HL of each piston sleeve 21 is smaller than the axial groove length NL of the corresponding sleeve groove 24 in the direction of the piston axis KA. The sleeve lengths HL of the individual piston sleeves 21 do not have to be equal. The individual sleeve grooves 24 are spaced apart from each other axially in the direction of the piston axis KA. Two adjacent sleeve grooves 24 are axially separated from each other by the radial web 31 of the piston body 20. The above description applies mutatis mutandis to each piston sleeve 21 and each sleeve groove 24 with respect to Figure 2. Accordingly, the sleeve gap HS, axial gap AS, groove length NL, and sleeve length HL exist in the piston body 20. In the embodiment shown in Figure 3, the piston body 20 is preferably manufactured in multiple parts so as to have a plurality of combined piston body segments.
[0053] Based on the piston transmission device 12, a lateral force may be applied to the piston 3 via the connection of the piston rod 4 to the piston 3. Due to the small sealing gap DS, such a lateral force will result in a contact force between the piston 3, specifically the piston sleeve 21, and the cylinder sliding surface 10. Such a lateral force, for example, due to the axial offset between the cylinder 10 and the piston transmission device 12, can be compensated for by the radial motion of the piston sleeve 21.
[0054] The axial sleeve length HL of the piston sleeve 21 is preferably in a size ratio of HL / D = 1.5 to 3 with respect to the outer diameter D of the piston sleeve 21. In the case of multiple piston sleeves 21, this size ratio applies to the combined sleeve length HL of all piston sleeves 21.
[0055] The radial thickness d of the piston sleeve 21 is preferably less than 30% of the outer diameter D of the piston sleeve 21. The lower this ratio, the better the lateral force is divided, and in this case, the lower limit of the thickness d is set by the required strength of the piston sleeve 21.
[0056] Any angular offset that may occur between the axes of the piston 3 and the piston rod 4 will result in a high contact force between the piston 3 and the cylinder 10, depending on the resulting tilt position of the piston 3. For this reason, the connection between the piston 3 and the piston rod 4 is manufactured with additional radial floating support. In this case, "floating support" means that the piston 3 and the piston rod 4 are freely movable relative to each other in the radial direction (excluding the frictional force acting on them). The possibility of free radial movement is, of course, only within the range of possible movement margins defined by the design. It is preferable that the piston rod 4 can move freely radially relative to the piston 3 without radially contacting the piston 3. The radial movement margin of the piston rod 4 during operation is, of course, determined by the embodiment of the piston transmission 12, but can be assumed to be known based on the known design of the piston compressor 1.
[0057] Figure 4 shows an embodiment of a piston 3 supported by a piston rod 4 so as to be movable in the radial direction. A piston rod notch 40 with a housing depth AT is provided at the axial end of the piston 3 facing the piston rod 4. The axial end of the piston rod 4 facing the piston 3 is positioned within the piston rod notch 40, so that the axial end face 41 of the piston rod 4 facing the piston abuts against the bottom face 42 of the piston rod notch 40. Accordingly, the piston rod 4 is positioned within the piston rod notch 40 over an axial length corresponding to the housing depth AT. The radial mobility of the support is achieved by providing a piston rod gap KS in the radial direction between the piston rod 4 and the piston rod notch 40. The piston rod gap KS can be achieved by the minimum inner diameter of the piston rod notch 40 being larger than the maximum outer diameter of the axial end of the piston rod 4 positioned within the piston rod notch 40. The piston rod gap KS is defined by the minimum radial distance between the piston rod 4 and the piston rod notch 40 that occurs when the piston rod 4 and the piston rod notch 40 are arranged concentrically.
[0058] The piston rod gap KS is preferably selected within the range of 0.1 mm to 10 mm for design purposes.
[0059] If the piston rod 4 is manufactured in multiple parts, the “piston rod” positioned within the piston rod notch 40 includes at least the portion of the piston rod 4 that forms the axial end facing the piston 3, but may also include other portions of the piston rod 4, or even the entirety of the piston rod 4.
[0060] The axial end face 41 of the piston rod 4 and / or the axial bottom face 42 of the piston rod notch 40 may be convex (outward) curved. This allows, on the one hand, to narrow the contact surface between the piston rod 4 and the piston 3, and on the other hand, such a convex curve also allows for some pivoting motion between the piston 3 and the piston rod 4. Consequently, even if the piston rod 4 pivots slightly relative to the piston 3, the piston 3 remains concentric within the cylinder Z1. This, in particular, allows for compensation of the aforementioned angular offset that may occur, which cannot be compensated for by the floating support of the piston sleeve 21.
[0061] The arrangement of the piston 3, which is supported in a floating manner by the piston rod 4, as shown in Figure 4, may also be provided in the embodiment shown in Figure 2 or Figure 3.
[0062] Figure 5 shows an alternative radially floating support for the piston 3 on the piston rod 4, with only the piston rod-side end of the piston 3 shown. Figure 6 shows an enlarged view for illustrative purposes of the embodiment of the support in Figure 5. The piston 3 may be manufactured as described in particular with respect to Figure 2, or Figure 3, or Figure 4.
[0063] In the embodiment shown in Figure 5, compared to the embodiment in Figure 4, the piston rod 4 and the piston 3 are additionally elastically connected to each other radially within the piston rod notch 40 via a radially elastic coupling member 46. In the illustrated embodiment, an elastic ring, such as an O-ring, is positioned radially between the piston rod 4 and the piston rod notch 40 as the radially elastic coupling member 46. For this purpose, the piston rod 4 and the piston rod notch 40 are provided with ring housing grooves 47a and 47b, respectively. When the piston rod 4 is positioned within the piston rod notch 40, the elastic ring is housed and held within both of these ring housing grooves 47a and 47b. However, the elasticity of the elastic ring and the dimensions of the ring housing grooves 47a and 47b are selected so that the radial force generated becomes small enough to be negligible, thereby not impairing the floating support of the piston 3 radially.
[0064] When the piston compressor 1 is operating, the cylinder pressure p acting inside cylinder Z1 Z This force acts on the piston 3, and this force also acts to cause the piston 3 to move together with the piston rod 4 from top dead center OT when the piston rod 4 returns to its original position. Therefore, when the piston 3 is in operation, the radially elastic coupling member 46 has no function and does not impair the function of the piston 3.
[0065] However, when starting the piston compressor 1, for example after maintenance, it cannot be expected that the cylinder Z1 is already pressure-driven. When the piston drive 12 is driven in this state, the reciprocating motion of the piston 3 may cause contact loss between the piston rod 4 and the piston 3 at top dead center OT. This continues until the piston 3 periodically strikes the cylinder cover and the pressure generated by the compression motion is sufficient to compensate for such contact loss. The radially elastic coupling member 46 also serves to prevent such contact loss when starting the piston compressor 1. The radially elastic coupling member 46 forms an elastic connection between the piston rod 4 and the piston 3, thereby allowing the piston 3 to be easily pulled back together with the piston rod 4.
[0066] Figure 6 also shows the convex end face 42 of the piston rod 4.
[0067] The arrangement of the piston 3, which is radially floating on the piston rod 4 as shown in Figures 5 and 6, may also be provided in the embodiment shown in Figure 2 or Figure 3.
[0068] Instead of the radially elastic coupling member 46 between the piston rod 4 and the piston 3, an axially elastic coupling member 46 may be provided between the piston rod 4 and the piston 3, as described with reference to Figure 7. Such an axially elastic coupling also allows the piston 3 to be pulled back through the piston rod 4, but does not impair the arrangement of the piston 3, which is radially floating on the piston rod 4.
[0069] An axially elastic coupling member 46 is provided in the form of an elastic body at the axial end of the piston rod 4 facing the piston 3. The arrangement of the piston rod 4 with the piston rod gap KS within the piston rod notch 40 does not change in any way. The bottom surface 42 of the piston rod notch 40 and the end surface 41 of the piston rod 4 abut each other in the same way. At least one of these surfaces may be curved in a convex shape as described above. The axially elastic coupling member 46 is merely an additional addition. For this purpose, a coupling member notch 48 may be provided at the axial end of the piston rod 4, and the axially elastic coupling member 46 is arranged therein. Threaded pins 49 protrude from both ends of the axially elastic coupling member 46, with one threaded pin 49 screwed into a threaded hole 50 in the piston body 20, and the opposite threaded pin 49 screwed into a threaded hole 50 in the piston rod 4.
[0070] The arrangement of the piston 3, which is radially floating on the piston rod 4 as shown in Figure 7, may also be provided in the embodiment shown in Figure 2 or Figure 3.
[0071] Figure 8 illustrates another embodiment of the arrangement of a piston 3 supported radially floating on a piston rod 4. In this embodiment as well, the axial end of the piston rod 4 is positioned within the piston rod notch 40 of the piston body 20 with a piston rod gap KS in the lateral direction. The bottom surface 42 of the piston rod notch 40 and the end surface 41 of the piston rod 4 also abut each other in the axial direction in the same manner. At least one of these surfaces may be curved in a convex shape, as described above.
[0072] In the embodiment shown in Figure 8, the piston rod 4 is actively and additionally pressed axially toward the bottom surface 42 of the piston rod notch 40 by the initial stress unit 51, without impairing the floating support, particularly the piston rod gap KS.
[0073] In the illustrated embodiment, the initial stress unit 51 is manufactured to have a clamp sleeve 52. The clamp sleeve 52 is attached to the axial end of the piston body 20, for example, by screw connection as shown in Figure 8, or by adhesive, press-fitting, or other methods. The axial end of the clamp sleeve 52 facing away from the piston 3 is provided with a shoulder 53 projecting radially inward. The piston rod 4 is provided with a radial web 54 between the shoulder 53 and the piston rod notch 40. The radial web 54 may be formed by a ring attached to the piston rod 4, or it may be integrally formed with the piston rod 4. The radial web 54 is fixedly coupled to the piston rod 4 and does not move relative to the piston rod 4. Between the radial web 54 and the shoulder 53 is provided a spring member 55 that applies an initial stress to the radial web 54 and consequently to the piston rod 4 in the direction of the piston 3. Accordingly, a specific axial contact force can be ensured between the piston rod 4 and the piston 3, which is particularly preferable when changing the direction of force. This prevents contact loss between the piston rod 4 and the piston 3 during the unpressurized starting phase of the piston compressor 1.
[0074] The arrangement of the piston 3, which is radially floating on the piston rod 4 as shown in Figure 8, may also be provided in the embodiment shown in Figure 2 or Figure 3.
[0075] Based on the piston 3, which is radially floatingly supported on the piston rod 4, and the piston sleeve 21, which is floatingly supported on the piston 3, the piston body 20 may slide radially, which may lead to contact between the piston body 20 and the cylinder sliding surface 10. To prevent this, a support ring 28 is positioned on the piston body 20 at the piston rod-side end of the piston 3, as shown in Figures 2, 3, 4, and 5. The support ring 28 is positioned in a support ring groove 29 on the outer circumferential surface of the piston body 20, which is, of course, axially spaced from the sleeve groove 24 for at least one piston sleeve 21, which is closest to the piston rod-side end of the piston 3, in the direction of the piston axis KA. The support ring 28 protrudes radially from the outer circumferential surface of the piston body 20.
[0076] The support ring 28 supports any residual lateral forces that may arise during the operation of the piston 3 against the cylinder sliding surface 10, and in particular, ensures that the piston body 20 slides in the center along the cylinder axis (which coincides with the piston axis KA). Accordingly, the support ring 28 abuts against the cylinder sliding surface 10 when the piston 3 is operating, and therefore it is preferable that it be made from a tribologically advantageous wear-resistant material, particularly a polymer material. The support ring 28 is located at the piston rod side end and, consequently, on the opposite side of the pressure when the piston 3 is operating, and therefore only needs to withstand a small lateral force, resulting in minimal wear on the support ring 28. In addition, the support ring 28 may be manufactured with pressure compensation, so that no force based on the pressure difference that may act on the support ring 28 when the piston 3 is operating does not occur on the support ring 28. This also contributes to minimizing wear on the support ring 28.
[0077] The outer diameter of the support ring 28 is at most equivalent to the inner diameter of the cylinder sliding surface 10, and may be smaller than or larger than the outer diameter of at least one piston sleeve 21.
[0078] To prevent the piston body 20 from contacting the cylinder sliding surface 10 when the piston 3 is in operation, the maximum outer diameter of the piston body 20 (at least the portion of the piston body 20 that slides on the cylinder sliding surface 10 during operation) is in any case smaller than the outer diameter of the piston sleeve 21.
[0079] Referring to Figure 4, other possible components of the piston 3 according to the present invention will be described. These additional components are not related to embodiments of the arrangement of the piston 3, which is floatingly supported on the piston rod 4.
[0080] In the embodiment of the piston 3 shown in Figure 4, a throttling groove 30 in the form of a circumferential groove is provided on the outer circumferential surface 27. These throttling grooves 30 are optional, and only one throttling groove 30 may be provided. At least one throttling groove 30, if present, serves to capture and collect any particles that may be present between the outer circumferential surface 27 of the piston sleeve 21 and the cylinder sliding surface 10. This prevents scratching on the cylinder sliding surface 10 and / or the outer circumferential surface 27 of the piston sleeve 21, or the detachment of chips caused by such particles. In addition, the throttling groove 30 may be configured to provide a throttling effect that enhances the non-contact sealing performance at the piston sleeve 21 when the piston 3 is in operation, similar to the case of a non-contact labyrinth seal. Within the circumferential groove 30, vortex formation occurs in the compression medium flowing along the outer circumferential surface 27 of the piston sleeve 21, which leads to losses, and consequently reduces the acting pressure. This throttling effect is particularly effective when the ratio of the sealing gap DS to the radial groove depth of the circumferential groove is at least 1.5, preferably at least 2.
[0081] As shown in Figure 3, in an embodiment having a piston body with multiple piston sleeves 21, a throttling groove 30 having the same function as the throttling groove in Figure 2 is naturally formed between two adjacent piston sleeves 21 based on the axial separation of the sleeve grooves 24. Nevertheless, in the embodiment of Figure 3 as well, at least one additional throttling groove 30 may be provided on the outer circumferential surface 27 of at least one piston sleeve.
[0082] In the embodiment shown in Figure 4, a gap ring 43, preferably made of a polymer such as PTFE, is provided in the axial gap AS between the piston body 20 and the axial end face of the piston sleeve 21 facing away from the crank rod 4. The gap ring 43 is manufactured such that, on the one hand, the piston sleeve 21 is not pinched axially in the sleeve groove 24, or that it is assembled with a small initial stress of <10% of the surface pressing force applied by the final compression pressure, thereby preventing the loss of floating support for the piston sleeve 21 in the sleeve groove 24. This can be achieved by manufacturing the gap ring 43 elastically, and the acting elastic force is small enough to be negligible in the direction of the piston axis KA. In the case of the rigid gap ring 43 embodiment, the axial length of the gap ring 43 may at most correspond to the axial length of the axial gap AS, and preferably is smaller than the axial length of the axial gap AS. On the other hand, the gap ring 43 is subjected to cylinder pressure p Z To ensure that the piston can reach the sleeve gap HS, a connection is made between the sleeve gap HS and the outer circumferential surface of the piston 3, or between the outer circumferential surface 27 of the piston sleeve 21. For this purpose, the gap ring 43 may be manufactured in a manner that such a connection occurs, or this connection may occur by the axial length of the gap ring 43 being smaller than the axial length of the axial gap AS.
[0083] Such a gap ring 43 may also be provided in the embodiment shown in Figure 2. In the embodiment shown in Figure 3, the gap ring 43 may be provided in at least the axial gap AS between the piston sleeve 21 and the piston body 20, preferably in all piston sleeves 21.
[0084] Furthermore, in the embodiment shown in Figure 4, an elastic sealing member 44 is provided between the piston rod-side end face 45 of the piston sleeve 21 and the piston rod-side axial end of the sleeve groove 24. In the embodiment shown in Figure 4, the sealing member 44 is positioned in a groove at the axial end of the sleeve groove 24, but it may also simply be inserted between the piston sleeve 21 and the axial end of the sleeve groove 24.
[0085] The sealing member 44 is at cylinder pressure p Z The sealing of the sleeve gap HS to which the pressure p is applied is improved. The piston sleeve 21 is as explained much earlier, Z This causes the sleeve groove 24 to be pressed toward the axial end on the piston rod side. This also ensures that the sleeve gap HS is sealed, thereby preventing the formation of a leakage path for the compressed medium through the sleeve gap HS. The sealing member 44 improves or makes such sealing more reliable. In particular, the sealing member 44 can also compensate for the inclined position of the piston sleeve 21 with respect to the piston shaft KA, and under such an inclined position, it is possible to prevent a leakage path from opening at the axial end of the piston sleeve 21 on the piston rod side.
[0086] The elastic sealing member 44 may alternatively be positioned between the cylinder-side axial end of the sleeve groove 24 and the piston sleeve 21. In this embodiment, the gap ring 43 is omitted or replaced by the elastic sealing member 44. The elastic sealing members 44 at the opposing axial ends (shown in Figure 4) are also omitted in this embodiment or can be manufactured to be appropriately gas permeable. This embodiment has the effect that the pressure applied at the inner diameter of the piston sleeve 21 is lower than the pressure applied at the outer diameter of the piston sleeve 21, and in this way the resulting pressure difference causes the piston sleeve 21 to be slightly compressed. Although this increases the sealing gap DS and the resulting leakage during operation compared to the embodiment shown in Figure 4, this embodiment has the advantage that, in boundary cases, it is guaranteed that pressure generation higher than the pressure at the outer diameter of the piston sleeve 21 cannot occur at the inner diameter of the piston sleeve 21. Such pressure generation can lead to radial deformation of the piston sleeve 21 outward, which could lead to contact between the cylinder sliding surface 10 and the piston sleeve 21.
[0087] The sealing member 44 may be provided in the embodiment of Figure 2, as described above. In the embodiment of Figure 3, the sealing member 44 may be provided in at least one piston sleeve 21, preferably all piston sleeves 21. The sealing member 44 may be provided in addition to the gap ring 43, as shown in Figure 4, or it may be provided without the gap ring 43.
[0088] The materials of the cylinder sliding surface 10 and the piston sleeve 21 are selected so that thermal deformation due to their different coefficients of thermal expansion, along with deformation due to the applied pressure, allows for the maintenance of a functional sealing gap DS. In other words, it is ensured that the sealing gap DS is neither too large nor too small, based on thermal expansion and deformation due to the load. In this case, the coefficient of thermal expansion of the materials used is a maximum of 1 × 10⁻⁶ -6A difference of only 1 / K is preferable. Preferably, but not required, the thermal expansion coefficient of the material of the cylinder sliding surface 10 is greater than that of the material of the piston sleeve 21. This prevents the piston sleeve 21 from expanding more than the cylinder sliding surface 10, and consequently ensures that thermal closure of the sealing gap DS between the piston sleeve 21 and the cylinder sliding surface 10 is prevented. Possible material combinations that ensure this are ceramic for the cylinder sliding surface 10 and Invar (an iron-nickel alloy with a very low thermal expansion coefficient) for the piston sleeve 21, or steel for the cylinder sliding surface 10 and ceramic for the piston sleeve 21. Similarly, in a preferred design, a piston sleeve 21 made of steel can be combined with a ceramic cylinder sliding surface 10. Embodiments in which both contact mating surfaces are made of ceramic material are also possible.
[0089] Figure 9 shows the cylinder Z1 of a piston compressor 1 in which a piston 3, comprising a piston rod 4 according to the present invention, reciprocates. The cylinder chamber K1 within the cylinder Z1 is separated by a cylinder head 60 in this embodiment, and an intake valve SV1 and a pressure valve DV1 are also located therein. The piston 3, comprising the piston rod 4, may be manufactured as described above with reference to Figures 2 to 8. In this embodiment, the cylinder sliding surface 10 is formed on a cylinder sliding bush 61 located inside the cylinder Z1, which corresponds to a conventional configuration.
[0090] In the configuration shown in Figure 9, cylinder Z1 is manufactured as a cooled type. A temperature control pipe 62 is provided inside the cylinder housing of cylinder Z1 and around the cylinder sliding surface 10 or cylinder sliding bush 41, through which a temperature-controlled temperature control medium is guided. The temperature control pipe 62 and temperature control medium allow the cylinder sliding surface 10 to be temperature-controlled to a specific temperature, thereby controlling the thermal expansion of the cylinder sliding surface 10 and, consequently, the sealing gap DS between the cylinder sliding surface 10 and at least one piston sleeve 21. For this purpose, a temperature sensor 63 may also be provided inside cylinder Z1 to detect the temperature in the region of the cylinder sliding surface 10. The temperature sensor 63 may be incorporated into a closed control loop to control the temperature in the region of the cylinder sliding surface 10 by adjusting the temperature of the temperature control medium. For this purpose, it is possible to not only cool but also heat the temperature control medium while controlling its temperature. Cooling is particularly necessary when cylinder Z1 is heated based on the compression of the compression medium during the operation of the piston compressor 1. During the starting process of the piston compressor 1, or when operating at low ambient temperatures, it may be necessary to heat the temperature control medium. The temperature of the temperature control medium can be controlled in addition to the current sealing gap DS in order to maintain the sealing gap DS within a desired range through thermal expansion of the cylinder sliding surface 10 and the piston sleeve 21. For this purpose, the sealing gap DS can be measured, for example, by ultrasonic measurement.
Claims
1. A non-contact sealing piston (3) having a piston rod (4) for a piston compressor (1), wherein the piston (3) includes a piston body (20) and at least one piston sleeve (21), the piston body (20) is provided with at least one sleeve groove (24), in which at least one of the piston sleeves (21) is supported radially and floating laterally with respect to the piston shaft (KA), wherein a sleeve gap (HS) is provided between the inner circumferential surface (22) of at least one of the piston sleeves (21) and the outer circumferential surface (23) of the sleeve groove (24), and the axial sleeve length (HL) of at least one of the piston sleeves (21) is smaller than the axial groove length (NL) of at least one of the sleeve grooves (24) to form an axial gap (AS) in the direction of the piston shaft (KA). A piston (3) having a piston rod (4) that seals in a non-contact manner is characterized in that a piston rod notch (40) is formed at the axial end of the piston rod side of the piston body (20) on the piston rod side, the axial end face (41) of the piston rod (4) abuts in the axial direction against the bottom surface (42) of the piston rod notch (40), the piston body (20) is supported radially in a floating manner and positioned on the piston rod (4) by providing a piston rod gap (KS) radially between the piston rod notch (40) and the piston rod (4), and a support ring groove (29) is provided at the axial end of the piston rod side of the piston body (20) on the piston rod side of the piston body (20) on the outer circumferential surface of the piston body (20) in which a support ring (28) projecting radially from the outer circumferential surface of the piston body (20) is positioned.
2. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to claim 1, characterized in that the piston body (20) is provided with a plurality of sleeve grooves (24) spaced apart in the axial direction in the direction of the piston shaft (KA), and one piston sleeve (21) is supported radially and floatingly within each of the sleeve grooves (24).
3. A non-contact sealing piston (3) having a piston rod (4) according to claim 1 or 2, characterized in that at least one throttling groove (30) is provided on the outer circumferential surface (27) of at least one of the piston sleeves (21).
4. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to claim 1, characterized in that a plurality of throttling grooves (30) spaced apart in the axial direction are provided on the outer circumferential surface (27) of at least one of the piston sleeves.
5. The piston (3) having a piston rod (4) according to any one of claims 1 to 4, characterized in that the support ring groove (29) is positioned in the axial direction between the axial end of the piston body (20) on the piston rod side and the axial end of the sleeve groove (24) located closest to the axial end of the piston body (20) on the piston rod side.
6. A piston (3) for non-contact sealing having a piston rod (4) according to any one of claims 1 to 5, characterized in that the axial end face (41) of the piston rod (4) and / or the bottom face (42) of the piston rod notch (40) are manufactured to be curved in a convex shape.
7. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to any one of claims 1 to 6, characterized in that the piston rod (4) is elastically coupled to the piston body (20) within the piston rod notch (40) via a radial or axial elastic coupling member (46).
8. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to claim 7, characterized in that an elastic ring is provided between the piston rod (4) and the piston rod notch (40) as the radially elastic coupling member (46).
9. A piston (3) for contactless sealing having a piston rod (4) according to claim 8, characterized in that a ring housing groove (47a, 47b) is provided on each piston rod (4) within the piston rod notch (40), and the elastic ring is arranged in the ring housing groove (47a, 47b).
10. The piston (3) having a piston rod (4) according to claim 7, characterized in that the axially elastic coupling member (46) is manufactured as an elastic body, a coupling member notch (48) is provided at the axial end of the piston rod (4) facing the piston (3) in which the axially elastic coupling member (46) is positioned, the axially elastic coupling member (46) is provided with threaded pins (49) that protrude in the axial direction at both axial ends, one threaded pin (49) is screwed into a threaded hole (50) in the piston body (20), and the threaded pin (49) opposite to it is screwed into a threaded hole (50) in the piston rod (4).
11. A piston (3) for non-contact sealing having a piston rod (4) as described in any one of claims 1 to 6, characterized in that a clamp sleeve (52) is attached to the piston body (20) at the axial end of the piston body (20) on the piston rod side, a shoulder (53) protrudes radially inward at the axial end of the clamp sleeve (52) that faces away from the piston (3), a radial web (54) is provided on the piston rod (4) between the shoulder (53) and the piston rod notch (40) in the axial direction, and a spring member (55) that presses the piston rod (4) toward the piston (3) is arranged between the radial web (54) and the shoulder (53).
12. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to any one of claims 1 to 11, characterized in that an elastic sealing member (44) is disposed between the piston rod side end face (45) of the piston sleeve (21) and the piston rod side axial end of the sleeve groove (24).
13. A piston (3) having a piston rod (4) according to any one of claims 1 to 12, characterized in that a gap ring (43) is arranged in the axial gap (AS) between the piston body (20) and the axial end face side of the piston sleeve (21) facing away from the piston rod (4), thereby establishing a connection between the outer circumferential surface of the piston (3) and the sleeve gap (HS).
14. A cylinder for a piston compressor (1), wherein a non-contact sealing piston (3) having a piston rod (4) according to any one of claims 1 to 13 is arranged in the cylinder (Z1) to perform a reciprocating motion, the piston (3) at least partially partitioning a compression chamber (K1) formed in the cylinder (Z1), and a sealing gap (DS) is provided between the outer circumferential surface (27) of at least one of the piston sleeves (21) or the outer circumferential surfaces (27) of a plurality of the piston sleeves (21) and the cylinder sliding surface (10) of the cylinder (Z1).
15. The cylinder according to claim 14, characterized in that a temperature control pipe (62) is arranged around the cylinder sliding surface (10) within the cylinder (Z1), and a temperature-controlled temperature control medium is guided through it.
16. The cylinder according to claim 14 or 15, characterized in that, as materials, ceramic is used for the cylinder sliding surface (10) and Invar for the piston sleeve (21), or steel is used for the cylinder sliding surface (10) and ceramic for the piston sleeve (21), or ceramic is used for the cylinder sliding surface (10) and steel for the piston sleeve (21), or ceramic is used for the cylinder sliding surface (10) and ceramic for the piston sleeve (21).
17. A piston compressor having at least one cylinder (Z1) according to any one of claims 14 to 16, characterized in that the piston rod (4) is coupled to the piston transmission device (12) of the piston compressor (1) at the axial end opposite to the piston (3).