Non-contact sealing piston with piston rod

Abstract

The invention relates to a piston for use in piston compressors for high-pressure applications, characterized in that the piston (3) comprises a piston body (20), on which at least one sleeve groove (24) is provided, and at least one piston sleeve (21), on which at least one sleeve groove (24) is provided, the at least one piston sleeve (21) is arranged in the sleeve groove (24) in a floating bearing manner in a radial direction transverse to the piston axis (KA) in such a way that a sleeve gap (HS) is provided between an inner circumferential surface (22) of the at least one piston sleeve (21) and an outer circumferential surface (23) of the sleeve groove (24), an axial sleeve length (HL) of the at least one piston sleeve (21) in the direction of the piston axis (KA) is smaller than an axial groove length (NL) of the at least one sleeve groove (24), a piston rod recess (40) is formed at a piston rod-side axial end of the piston body (20), in which piston rod recess an axial end of the piston rod (4) is arranged, wherein an axial end face (41) of the piston rod (4) axially bears against a bottom face (42) of the piston rod recess (40), and the piston body (20) is arranged on the piston rod (4) in a floating bearing manner in the radial direction in such a way that a piston rod gap (KS) is provided in the radial direction between the piston rod recess (40) and the piston rod (4), a bearing ring groove (29) is provided on the outer circumferential surface of the piston body (20) at the axial end of the piston rod side of the piston body (20), in which bearing ring groove a bearing ring (28) is arranged, which radially protrudes beyond the outer circumferential surface of the piston body (20).

Description

Technical Field

[0001] The present invention relates to a non-contact sealed piston with a piston rod for a reciprocating compressor, and the application of such a piston in the cylinder of a reciprocating compressor. Background Technology

[0002] As is well known, in a reciprocating compressor, the piston reciprocates within the cylinder chamber to compress the medium inside. To ensure sufficient and efficient compression of the medium, a sealing element is required between the moving piston and the cylinder or its sliding surface. Common sealing elements are contact seals, such as the well-known piston rings, which are embedded in grooves on the piston and abut against the cylinder's sliding surface. The piston rings' task is to seal the cylinder chamber and minimize leakage of the compressed medium through the piston. Multiple piston rings are typically provided on the piston.

[0003] Piston rings and cylinder sliding surfaces suffer from friction and wear due to contact, piston speed, and the applied lateral forces. This wear, in particular, limits the service life of piston rings. Therefore, piston rings are maintenance components that must be replaced periodically.

[0004] In modern reciprocating compressors, piston rings are almost entirely made of polymer-based materials. In this case, these are primarily high-performance polymers whose frictional properties are improved through filler materials in a way that minimizes friction and wear. This, in turn, extends the service life of the piston rings.

[0005] Piston compressors can also be categorized into oil-lubricated and oil-free applications. In oil-lubricated applications, a thin oil film forms an interface between the pressurized piston rings and the cylinder sliding surface. Compared to non-lubricated applications, this significantly reduces friction and wear. Furthermore, the oil film also provides a sealing effect. However, in many applications, oil-lubricated piston rings are not feasible because the lubricating oil would also contaminate the compressed medium, which is generally undesirable and unacceptable.

[0006] In non-lubricated, oil-free applications, tribology (friction and wear) is defined only by the properties of the compression medium, operating conditions (such as the average piston speed, the acting lateral forces, etc.), the material and coating of the cylinder sliding surface, and the material of the piston rings (e.g., polymeric materials and their filler materials). Here, a transfer film is formed between the piston ring material and the cylinder sliding surface.

[0007] For non-lubricated high-pressure applications with a final compression pressure >250 bar, typical high-performance polymers also reach their material limits, and piston ring friction and wear increase excessively. Although the local contact pressure between the piston ring and cylinder sliding surface can be reduced by appropriately modifying the piston ring design, this only increases the possible final compression pressure to approximately 350 bar in non-lubricated applications. Furthermore, piston rings cannot be used because they wear too quickly, resulting in an excessively short achievable piston ring life.

[0008] To achieve an industrially applicable service life, high-pressure pistons used in reciprocating compressors (for compressing final pressures <350 bar) are equipped with a large number of piston rings. In this case, 10 to 20 piston rings may be used. However, this results in an excessively long piston structure, which in turn leads to an excessively long reciprocating compressor structure, and may also reduce the heat dissipation of each piston ring, thus limiting the service life.

[0009] Each piston ring is installed in a groove in the piston. For high-pressure applications, metal support rings are also installed behind the piston rings (on the side facing away from the pressure) to support these piston rings against the cylinder pressure acting through them. The function of these support rings is to prevent the piston rings from being squeezed into the gap between the cylinder sliding surface and the piston diameter due to pressure.

[0010] For example, WO 2010 / 084071 A1 discloses a piston for high-pressure applications, which includes piston rings and a support ring. The specification also describes the use of the support ring on the piston for guiding and centering the piston within the cylinder.

[0011] The typical clearance width between the support ring and the cylinder sliding surface varies within a range of several 0.1 mm. Smaller clearance widths are not specified in industrial applications because of the risk of metal-to-metal contact between the cylinder sliding surface and the support ring due to tolerances and the lateral forces transmitted to the piston via the crankshaft connecting rod mechanism. Metal-to-metal contact is associated with high wear and can also damage the cylinder sliding surface.

[0012] An alternative known sealing solution between the piston and cylinder sliding surfaces is a non-contact seal, such as a labyrinth seal or a gap seal. In this case, the seal is achieved through a cylindrical sealing gap between the piston's outer diameter and the cylinder sliding surface. In a labyrinth seal design, this sealing gap creates an additional throttling effect through additional grooves. To achieve a sufficient seal in high-pressure applications >350 bar, this sealing gap must be in the range of <50 μm. This very small sealing gap presents a series of challenges.

[0013] The unavoidable non-uniformity of the piston and / or cylinder sliding surfaces due to manufacturing reasons, manufacturing tolerances on the piston and cylinder sliding surfaces, and the lateral forces acting from the crankshaft connecting rod mechanism almost inevitably lead to metal-to-metal contact between the piston and cylinder sliding surfaces when the seal clearance height is <50μm. This causes significant wear and damage to the reciprocating compressor components even after a short period of operation.

[0014] Due to the high load, pistons and cylinders are typically made of steel. Uneven temperature distribution between the piston and cylinder, caused by thermal expansion, alters the effective clearance height of the seal. This can lead to piston-cylinder contact even with small temperature differences, as is common in reciprocating compressors, because the effective clearance height approaches zero, or it can cause excessive leakage because the clearance height is increasing.

[0015] Especially in high-pressure applications with a final compression pressure >350 bar, pressure-induced deformation of the cylinder and / or piston may occur, resulting in a change in clearance height within the range of 10 μm. This could again lead to similar negative effects, namely contact or increased clearance height.

[0016] Last but not least, micron-sized particles entrained with the compressed medium may fall into the sealing gaps, which can cause piston jamming or damage the piston surface and / or cylinder sliding surfaces.

[0017] For these reasons, although clearance sealing in reciprocating compressors is theoretically possible, it is not known to have been industrially implemented in reciprocating compressors to date.

[0018] DE2160816A illustrates a piston for a dry-running piston compressor. Plastic piston rings are radially supported on a fixed surface on the piston, thus providing a virtually wear-free labyrinth seal with a small clearance relative to the cylinder wall. Guide rings ensure straight guidance of the piston. However, even slight tilting of the piston within the cylinder can lead to accelerated wear of the piston rings and increased clearance, thereby reducing the sealing effect.

[0019] WO 98 / 31936 A1 describes a non-lubricated reciprocating compressor with a metal cylinder sliding surface and a piston with a plastic sleeve, which together form a clearance annular seal. The piston can move parallel within the cylinder due to the piston rod's pivoting capability enabled by its support.

[0020] The aforementioned issues related to piston rings and clearance seals have historically limited the industrial application of reciprocating compressors in non-lubricated high-pressure applications, particularly those with a final compression pressure >350 bar. For such high-pressure applications, other compressor types, such as diaphragm compressors or ion compressors, have been used to date. In diaphragm compressors, the short lifespan of the diaphragm is a concern. In ion compressors, the ionic liquid may mix with the compressed medium, which is generally undesirable. Furthermore, the lifespan of ion compressors is also an issue.

[0021] The use of reciprocating compressors in high-pressure applications is also advantageous because reciprocating compressors are a known, reliable, and primarily simple type of compressor. Summary of the Invention

[0022] Therefore, the object of the present invention is to provide a piston for a reciprocating compressor that can be used in reciprocating compressors for high-pressure applications.

[0023] This is achieved using a non-contact sealing piston as described in claim 1. Non-contact gap sealing can be achieved using the piston designed according to the invention because its structural design—particularly employing a double radial floating support or mounting method: a first floating support for the piston sleeve and a second floating support for the piston on the piston rod—reduces the force required to guide the piston to a sufficiently low level that the aerodynamic force generated between the piston and the cylinder is sufficient to minimize contact between the piston and the cylinder. Therefore, extremely small sealing gaps can be selected and operated. Simultaneously, pressure balance is achieved in the piston sleeve through axial clearance and sleeve clearance, so that the piston sleeve does not deform or only slightly deforms due to the operating pressure, and thus does not affect the sealing gap. Therefore, the piston with piston rod of the present invention can be used particularly for high-pressure applications with final compression pressures up to 1000 bar and higher. Such final compression pressures are particularly common in hydrogen applications. Thus, for the first time, a reciprocating compressor can be used for such final compression pressures.

[0024] In an alternative design, the piston body may be provided with a plurality of sleeve grooves spaced axially along the piston axis, wherein a piston sleeve is arranged radially and floatingly in each sleeve groove.

[0025] If at least one throttling groove, preferably multiple axially spaced throttling grooves, is provided on the outer circumferential surface of the piston body, particles that may be present between the outer circumferential surface of the piston body and the cylinder sliding surface can be captured. This prevents scratches from forming on the cylinder sliding surface and / or the outer circumferential surface of the piston body. Furthermore, such throttling grooves can improve the sealing effect.

[0026] If the axial end face of the piston rod and / or the bottom surface of the piston rod groove are designed to be convex and arched, it can also compensate for possible angular deviations between the piston and the cylinder and avoid high contact forces caused by such angular deviations.

[0027] If the piston rod is elastically connected to the piston body in the piston groove by a radial or axial elastic connecting element, it can prevent the piston body from lifting off the piston rod when the reciprocating compressor starts, without affecting the floating support of the piston body. The same effect can be achieved using a preload unit that actively presses the piston body onto the piston rod. Attached Figure Description

[0028] The invention is now described in detail with reference to Figures 1 to 9, which illustratively and schematically, but not limitingly, show advantageous structural designs of the invention. Wherein:

[0029] Figure 1 shows a piston compressor;

[0030] Figure 2 and Figure 3 Each of the non-contact sealing pistons according to the present invention is shown;

[0031] Figure 3 Details of the non-contact sealing piston according to the present invention are shown;

[0032] Figures 4 to 8 The radially movable supports of the piston on the piston rod are shown respectively; and

[0033] Figure 9 The cylinder of a piston compressor having the piston of the present invention. Detailed Implementation

[0034] Although the operating principle of the reciprocating compressor 1 is generally well known, it will be briefly described below with reference to FIG1. ​​FIG1 shows a cross-sectional view of a reciprocating compressor 1 according to an advantageous embodiment. The reciprocating compressor 1 has a compressor housing 2, inside which at least one cylinder Z1 is provided, and a piston 3 is reciprocating within the cylinder. The piston 3 reciprocates within the cylinder Z1 between top dead center OT and bottom dead center UT. The piston 3 partially defines a cylinder cavity K1 within the cylinder Z1, in which a compressed medium (e.g., hydrogen) is compressed by the movement of the piston 3. The piston 3 is connected to and driven by a reciprocating piston rod 4 during operation. The end of the piston rod 4 opposite to the piston 3 is connected to a piston drive mechanism 12, which generates the reciprocating motion of the piston rod 4 and the piston 3.

[0035] The piston drive mechanism 12 is designed as a crankshaft connecting rod mechanism with a crosshead 5 in the design according to FIG. 1. However, the piston drive mechanism 12 can also adopt a different design, such as with a cam drive mechanism. In this design, the piston rod 4 is connected at the end opposite to the piston 3 to a crosshead 5 that is axially supported in the compressor housing 2. The crosshead 5 is driven to reciprocate the piston rod 4 and the piston 3. In the illustrated construction design, the crosshead 5 is driven by a crankshaft connecting rod mechanism. For this purpose, the crosshead 5 is hinged to the end of a connecting rod 6, and the other end of the connecting rod 6 is hinged to the connecting rod journal 7a of the crankshaft 7. The crankshaft 7 rotates during the operation of the reciprocating compressor 1. The crankshaft 7 can be driven by a drive device not shown in FIG. 1, such as an electric motor. During the operation of the compressor 1, the rotational motion of the crankshaft 7 is converted into linear motion of the crosshead 5 via the connecting rod 6. The lateral force generated in this case is supported on the compressor housing 2 via the crosshead 5. Therefore, the piston rod 4 performs a purely linear reciprocating motion substantially along the piston axis KA direction. However, the piston 3 can also be driven in a different way than by using a crankshaft connecting rod mechanism.

[0036] The reciprocating piston rod 4 typically passes through a sealing unit 9, which is also located within the compressor housing 2. This sealing unit 9 seals against the piston drive housing 8 (as shown in Figure 1, the crankcase 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 designed with known sealing packing, including multiple sealing rings and, if necessary, other rings such as oil scraper rings, support rings, etc.

[0037] The design of the reciprocating piston rod 4 in the piston compressor 1 also makes it possible to design the piston compressor 1 as a double-acting compressor when needed, such as in the embodiment of FIG1. ​​In the double-acting piston compressor 1, a second cylinder Z2 with a second cylinder K2 is constructed on the side of the piston 3 opposite to the at least one first cylinder chamber K1, wherein the piston 3 also partially defines the second cylinder chamber K2. Therefore, during the reciprocating motion of the piston 3, the cylinder volume of the first cylinder Z1 decreases (compression stroke), while the cylinder volume of the second cylinder Z2 increases (intake stroke), and vice versa.

[0038] The design shown as a double-acting compressor should, of course, be understood as merely exemplary. In the simplest case, the reciprocating compressor 1 has only one unique cylinder Z1, which is defined by the end face of the piston 3 facing away from the piston drive mechanism 12.

[0039] Valves SV1, DV1, SV2, and DV2 are also arranged in a known manner on the cylinders Z1 and Z2 of the reciprocating compressor 1, wherein at least one suction valve SV1 or SV2 and at least one pressure valve DV1 or DV2 are provided on one cylinder Z1 or Z2. The valves can be designed as automatic valves, that is, controlled by the pressure of the operating cylinder. However, the valves can also be actively controlled, for example, by a well-known lifting gripper.

[0040] The embodiment according to Figure 1 shows a double-acting reciprocating compressor 1, wherein two suction valves SV1, SV2 are connected to a suction pipe SL for suction of compressed medium KM, and two pressure valves DV1, DV2 are connected to a pressure pipe DL for output of compressed medium KM.

[0041] If the piston surface on one side of piston 3 is designed to be smaller than the piston surface on the opposite side of piston 3, a multi-stage reciprocating compressor 1 can be realized. In this case, the suction pipe of the first compressor stage is only connected to the suction valve SV1 of the first compressor stage. The pressure pipe DL1 of the first compressor stage, which is connected to the pressure valve DV1 of the first compressor stage, is also the suction pipe SL of the second compressor stage, and therefore is connected to the suction valve SV2 of the second compressor stage. The pressure valve DV2 of the second compressor stage is connected to the pressure pipe DL2 of the second compressor stage.

[0042] In the design according to Figure 1, a piston ring 15 is arranged on the piston 3, which abuts against the cylinder sliding surface 10 of the cylinder body Z1, Z2 to seal the corresponding cylinder chambers K1, K2.

[0043] Figure 2 The illustration shows a design of a non-contact sealing piston 3, along with a portion of a piston rod 4, according to the invention, for use in the cylinder Z1 of a reciprocating compressor 1 for high-pressure applications. A non-contact gap seal is provided for sealing between the piston 3 and the cylinder Z1. As is typically done in a reciprocating compressor 1, the piston 3 reciprocates within the cylinder cavity K1 of the cylinder Z1. Figure 2 In this diagram, apart from the reciprocating piston rod 4, the piston drive mechanism 12 of the piston 3 is not shown because it is not important to this invention. Figure 2 In the illustration, the distance and gap are shown in a significantly exaggerated manner. In the case of gap sealing, piston 3 is designed to act only in one direction, thus defining only one cylinder chamber K1.

[0044] The piston 3 according to the invention comprises a piston body 20 and at least one hollow cylindrical piston sleeve 21. At least one sleeve groove 24, which is in the form of an annular groove, is provided on the piston body 20, and the at least one piston sleeve 21 is arranged in the sleeve groove in a floating support manner in a transverse direction transverse to the piston axis KA. In this case, "floating support" means that the piston sleeve 21 can move freely (except for possible frictional forces) relative to the piston body 20 in a radial direction transverse to the piston axis KA within the sleeve groove 24.

[0045] The floating support of the piston sleeve 21 on the piston body 20 is achieved by providing a radial sleeve gap HS between the inner circumferential surface 22 of the at least one piston sleeve 21 and the outer circumferential surface of the sleeve groove 24, and the axial sleeve length HL of the at least one piston sleeve 21 along the piston axis KA direction is less than the axial groove length NL of the at least one sleeve groove 24.

[0046] When the piston 3 is used in the piston compressor 1, a sealing gap DS is formed between the outer peripheral surface 27 of the at least one piston sleeve 21 and the cylinder sliding surface 10 of the cylinder body Z1 (in which the piston 3 moves during use). This sealing gap provides a gap sealing effect. Therefore, 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 peripheral surface 27 of the at least one piston sleeve 21.

[0047] Of course, the piston sleeve 21 can only move freely radially within a structurally defined range of motion on the piston body 20, which is essentially determined by the sleeve clearance HS on the piston 3. However, when the piston 3 is used in the cylinder Z1 of the reciprocating compressor 1, the possible mobility is of course determined by the sealing clearance DS. This mobility is also guaranteed when the piston 3 is used in the reciprocating compressor 1.

[0048] In order to manufacture piston 3, piston body 20 can be designed to consist of several parts, and is composed of multiple piston body components assembled into piston body 20. According to... Figure 2In the design, the piston body 20 consists of a piston intermediate portion 25 and a piston end portion 26, the piston end portion being fastened to the piston intermediate portion 25, for example, screwed onto or tightened into the piston intermediate portion. The piston end portion 26 faces the cylinder cavity K1 when the piston 3 is used in the piston compressor 1. Along the piston axis KA and at least one specific axial groove length NL, the outer diameter of the piston intermediate portion 25 is smaller than the outer diameter of the piston end portion 26. Therefore, at least one region of the piston intermediate portion 25 with the smaller outer diameter constitutes the at least one sleeve groove 24 with groove length NL, in which the piston sleeve 21 is floatingly supported.

[0049] To achieve floating support of 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, thereby forming a sleeve gap HS between the inner circumferential surface 22 of the piston sleeve 21 and the outer circumferential surface 23 of the sleeve groove 24. The width of the sleeve gap HS is preferably between 0.1 mm and 1 mm.

[0050] Furthermore, the axial sleeve length HL of the at least one piston sleeve 21 along the piston axis KA direction is less than the groove length NL of the sleeve groove 24 (in which the piston sleeve 21 is floatingly supported), thereby forming an axial clearance AS between the piston sleeve 21 and the piston body 20. This axial clearance AS is important for the floatingly supported piston sleeve 21 to avoid axially clamping the piston sleeve 21. The size of the axial clearance AS should be appropriately selected, but should not be too large, as this axial clearance AS creates an undesirable dead zone.

[0051] Piston 3 experiences cylinder pressure p during operation zTherefore, the piston sleeve 21 is pressed against the end of the sleeve groove 24 opposite to the cylinder body Z1 along the piston axis KA direction. The cylinder pressure also acts on the end face of the piston sleeve 21 facing the cylinder body Z1 due to the axial clearance AS. However, the end face of the piston sleeve 21 is smaller than the cross-section of the cylinder body, so the axial force acting is very small. This also reduces the generated frictional force, which should not, or at least should not significantly, limit the movement of the piston sleeve 21 in the radial direction transverse to the piston axis KA. The floating support of the piston sleeve 21 in the transverse direction is functionally unaffected under any circumstances. The contact surface of the piston sleeve 21 on the piston body 20 should preferably be chosen to be as large as possible so as to keep the contact pressure and associated deformation between the two surfaces as small as possible. Appropriate measures for pressure balancing between the two surfaces can further reduce the contact force. Therefore, the sleeve clearance HS and axial clearance AS between the piston sleeve 21 and the piston body 20 ensure that the piston sleeve 21 is floatingly supported on the piston body 20 in the radial direction. Therefore, the piston sleeve 21 can compensate for small lateral deviations (maximum in the region of the sleeve clearance HS) during the operation of the piston 3 within the cylinder Z1 of the reciprocating compressor 1, without generating significant lateral forces between the piston sleeve 21 and the cylinder sliding surface 10. This reduces potential wear where contact between the piston sleeve 21 and the cylinder sliding surface 10 is unavoidable. However, the sleeve clearance HS, at 0.1 mm to 1 mm, is insufficient to generate a significant additional dead zone (space in the cylinder Z1 that cannot be used for compression). It is well known that an excessively large dead zone reduces the delivery capacity of the reciprocating compressor 1 and is therefore undesirable.

[0052] The sleeve clearance HS and axial clearance AS also ensure that, when piston 3 is running in reciprocating compressor 1, the cylinder pressure p z The pressure balance is achieved between the outer circumferential surface 27 and the inner circumferential surface 22 of the piston sleeve 21, through the sleeve gap HS between the cylinder cavity K1 and the piston sleeve 21 and the piston body 20. This pressure balance reduces the deformation of the piston 3 caused by the applied pressure and the resulting changes in the sealing gap DS between the outer circumferential surface 27 of the piston sleeve 21 and the cylinder sliding surface 10.

[0053] The effective cylinder pressure p z It also ensures that the piston sleeve 21 is centered and aligned within the cylinder Z1. Due to hydrodynamic effects, a pressure rise occurs on the side where the sealing gap DS between the piston sleeve 21 and the cylinder sliding surface decreases, which in turn ensures that the piston sleeve 21 is moved back to the center position.

[0054] To achieve effective gap sealing, the sealing gap DS between the outer peripheral surface 27 of the piston sleeve 21 and the cylinder sliding surface 10 is between 1 μm and 50 μm. The outer peripheral surface 27 of the piston sleeve 21 advantageously performs a sealing function along the entire length HL of the sleeve due to the formed sealing gap DS.

[0055] Figure 3 The diagram illustrates a design for a piston 3 with multiple piston sleeves 21. Here, multiple piston sleeves 21, each radially floating and supported in a sleeve groove 24, are used instead of the design shown in the diagram. Figure 2 In the design, a piston sleeve 21 is radially floating and supported in a sleeve groove 24. Therefore, the original sealing surface for non-contact piston sealing, namely the outer peripheral surface of the piston sleeve 21, is divided into multiple sealing surfaces.

[0056] In this embodiment, each piston sleeve 21 is provided with a sleeve groove 24, in which a piston sleeve 21 is radially floatingly supported and arranged in a sleeve groove in such a way that a sleeve gap HS (which can also be different) is provided between the inner circumferential surface 22 of each piston sleeve 21 and the outer circumferential surface 23 of each corresponding sleeve groove 24, and the sleeve length HL of each piston sleeve 21 along the piston axis KA is less than the axial groove length NL of the corresponding sleeve groove 24. The sleeve length HL of each individual piston sleeve 21 is not necessarily the same. The sleeve grooves 24 are axially spaced apart from each other along the piston axis KA. Two adjacent sleeve grooves 24 are axially separated from each other by a radial rib 31 on the piston body 20. The above refers to Figure 2 The same description applies to each piston sleeve 21 and each sleeve groove 24. Therefore, the piston body 20 has a sleeve clearance HS, an axial clearance AS, a groove length NL, and a sleeve length HL. The piston body 20 is based on... Figure 3 In the preferred embodiment, it is designed to consist of several parts, having multiple piston body segments assembled together.

[0057] Due to the piston drive mechanism 12, lateral forces affect the piston 3 through the connection between the piston rod 4 and the piston 3. These lateral forces, due to the small sealing clearance DS, result in contact forces between the piston 3, specifically the piston sleeve 21, and the cylinder sliding surface 10. This lateral force—for example, caused by axial misalignment between the cylinder 10 and the piston drive mechanism 12—can be compensated for by the radial mobility of the piston sleeve 21.

[0058] The axial sleeve length HL of the piston sleeve 21 is preferably in a ratio of HL / D = 1.5 to 3 relative to the outer diameter D of the piston sleeve 21. In the case of multiple piston sleeves 21, this ratio is related to the sum of the sleeve lengths HL of all piston sleeves 21.

[0059] 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 smaller this ratio, the better the decoupling of the lateral force is achieved, wherein the lower limit of the thickness d is determined by the necessary strength of the piston sleeve 21.

[0060] The potential angular deviation between the axis of piston 3 and the axis of piston rod 4, resulting in a tilted position of piston 3, leads to a high contact force between piston 3 and cylinder 10. For this reason, the connection between piston 3 and piston rod 4 is additionally designed as a radially floating support. In this case, "floating support" also means that piston 3 and piston rod 4 can move radially freely relative to each other (except for the acting frictional force). Of course, free radial movement is only possible within the possible range of motion determined by the design. Piston rod 4 should be able to move radially freely relative to piston 3 without radially jamming on piston 3. The radial movement clearance of piston rod 4 during operation depends, of course, on the design of piston drive mechanism 12, but since the structure of piston compressor 1 is known, it can be assumed to be known.

[0061] Figure 4 The diagram illustrates a design for a piston 3 radially movable and supported on a piston rod 4. A piston rod groove 40 with a receiving depth AT is provided on 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 arranged in this piston rod groove 40, such that the axial end face 41 of the piston rod 4 facing the piston abuts against the bottom surface 42 of the piston rod groove 40. The piston rod 4 is arranged in the piston rod groove 40 with an axial length corresponding to the receiving depth AT. Radial movable support is achieved by providing a piston rod clearance KS radially between the piston rod 4 and the piston rod groove 40. This piston rod clearance KS is achieved by the minimum inner diameter of the piston rod groove 40 being greater than the maximum outer diameter of the axial end of the piston rod 4 arranged within the piston rod groove 40. This piston rod clearance KS is determined by the minimum radial distance between the piston rod 4 and the piston rod groove 40, which occurs when the piston rod 4 and the piston rod groove 40 are concentrically arranged.

[0062] The piston rod clearance KS is preferably selected in the design range of 0.1mm to 10mm.

[0063] If the piston rod 4 is designed to consist of multiple parts, then the "piston rod" arranged in the piston rod groove 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 all portions of the piston rod 4.

[0064] The axial end face 41 of the piston rod 4 and / or the axial bottom face 42 of the piston rod groove 40 can also be convex (outwardly) arched. This firstly reduces the contact area between the piston rod 4 and the piston 3, and secondly, such a convex arch allows for slight pivoting movement between the piston 3 and the piston rod 4. Therefore, when the piston rod 4 pivots slightly relative to the piston 3, the piston 3 remains concentrically held within the cylinder Z1. This particularly makes it possible to compensate for the aforementioned possible angular deviation, which cannot be compensated for by the floating support of the piston sleeve 21.

[0065] Figure 4 The piston 3 shown is floatingly supported on the piston rod 4, or it can be arranged according to... Figure 2 or Figure 3 In the design.

[0066] Figure 5 An alternative radial floating support for piston 3 on piston rod 4 is shown, with only the piston rod side end of piston 3 shown. For illustrative purposes, Figure 6 It shows Figure 5 The enlarged view of the support design is shown. Piston 3, in particular, can be used as... Figure 2 or Figure 3 or Figure 4 Designed as described.

[0067] According to Figure 5 In the design, and according to Figure 4 Compared to the previous design, the piston rod 4 and piston 3 are additionally radially and elastically connected to each other in the piston rod groove 40 by a radially elastic connecting element 46. In the illustrated design, an elastic ring (such as an O-ring) is radially arranged between the piston rod 4 and the piston rod groove 40 as the radially elastic connecting element 46. For this purpose, a ring receiving groove 47a and 47b are respectively provided on the piston rod 4 and the piston rod groove 40. The elastic ring is supported and held in these two ring receiving grooves 47a and 47b when the piston rod 4 is arranged in the piston rod groove 40. However, the elasticity of the elastic ring and the dimensions of the ring receiving grooves 47a and 47b are chosen such that the force generated in the radial direction is negligible, thereby not affecting the radial floating support of the piston 3.

[0068] During the operation of the reciprocating compressor 1, the cylinder pressure p that acts in cylinder Z1 is... z The cylinder pressure acting on piston 3 also ensures that piston 3 moves together with piston rod 4 from top dead center OT during piston rod 4's return motion. Therefore, the radially elastic connecting element 46 does not function during piston 3's operation and does not affect the function of piston 3.

[0069] However, when starting the reciprocating compressor 1, for example after maintenance, it cannot be expected that the cylinder Z1 has been pressurized. If the piston drive mechanism 12 is driven in this state, the reciprocating motion of the piston 3 may cause a loss of contact between the piston rod 4 and the piston 3 at the top dead center OT. This may cause the piston 3 to periodically strike the cylinder head until the pressure built up by the compression motion is sufficient to compensate for this loss of contact. The radially resilient connecting element 46 also helps to prevent this loss of contact during the start-up of the reciprocating compressor 1. The radially resilient connecting element 46 forms a resilient connection between the piston rod 4 and the piston 3, through which the piston rod 4 can be easily used to pull the piston 3 back.

[0070] exist Figure 6 The image also shows a convex end face 42 of the piston rod 4.

[0071] Figure 5 and Figure 6 The arrangement of the piston 3 radially floatingly supported on the piston rod 4, as shown in the figure, can also be configured according to... Figure 2 or Figure 3 In the design.

[0072] Alternatively, an axial elastic connecting element 46 can be provided between the piston rod 4 and the piston 3 to replace the radial elastic connecting element 46 between the piston rod 4 and the piston 3, such as in... Figure 7 As explained in the text. This axially elastic connection also allows the piston 3 to be pulled back via the piston rod 4 without affecting the radially floating support arrangement of the piston 3 on the piston rod 4.

[0073] An axially elastic connecting element 46, which is in the form of an elastic body, is provided on the axial end of the piston rod 4 facing the piston 3. The condition that "the piston rod 4 is arranged in the piston rod groove 40 with piston clearance KS" remains unchanged. The bottom surface 42 of the piston rod groove 40 also abuts against the end face 41 of the piston rod 4. At least one of these surfaces may also be convexly arched as described above. The axially elastic connecting element 46 is merely an additional feature. For this purpose, a connecting element groove 48 can be provided at the axial end of the piston rod 4, in which the axially elastic connecting element 46 is arranged. A threaded pin 49 extends from each of the two axial ends of the axially elastic connecting element 46, one of which is screwed into a threaded hole 50 on the piston body 20, while the other threaded pin 49 is screwed into a threaded hole 50 inside the piston rod 4.

[0074] Figure 7 The piston 3 shown is radially floatingly supported on the piston rod 4, or it can be arranged according to... Figure 2 or Figure 3 In the design.

[0075] exist Figure 8 Another structural design is described below, in which the piston 3 is radially floating and supported on the piston rod 4. In this design, the axial end of the piston rod 4 is also arranged in the piston rod groove 40 on the piston body 20 with a piston rod clearance KS in the transverse direction. The bottom surface 42 of the piston rod groove 40 and the end face 41 of the piston rod 4 also abut against each other axially. At least one of these surfaces may also be convex arched as described above.

[0076] According to Figure 8 In the design, the piston rod 4 is additionally pressed axially against the bottom surface 42 of the piston rod groove 40 by a preload unit 51 without affecting the floating support, and in particular without affecting the piston rod clearance KS.

[0077] In the illustrated embodiment, the preload unit 51 is designed with a clamping sleeve 52. This clamping sleeve 52 is fastened (e.g., as in…) Figure 8 The piston rod is fastened to the axial end of the piston body 20 by means of threaded connection, bonding, press-fitting, or other methods. A radially inwardly protruding shoulder 53 is provided on the axial end of the clamping sleeve 52 opposite to the piston 3. A radial rib 54 is provided on the piston rod 4 between the shoulder 53 and the piston rod groove 40. This radial rib 54 can be formed by a ring fastened to the piston rod 4, but it can also be integrally constructed with the piston rod 4. The radial rib 54 is fixedly connected to the piston rod 4 and does not move relative to the piston rod 4. A spring element 55 is provided between the radial rib 54 and the shoulder 53, which preloads the radial rib 54 and thus the piston rod 4 toward the piston 3. This ensures a certain axial contact force between the piston rod 4 and the piston 3, which is particularly advantageous in the case of force reversal, as it prevents loss of contact between the piston rod 4 and the piston 3 during the pressureless start-up phase of the reciprocating compressor 1.

[0078] Figure 8 The piston 3 shown is radially floatingly supported on the piston rod 4, or it can be arranged according to... Figure 2 or Figure 3 In the design.

[0079] Because the piston 3 is radially floatingly supported on the piston rod 4 and the piston sleeve 21 is floatingly supported on the piston 3, the piston body 20 can move radially, which may cause the piston body 20 to contact the cylinder sliding surface 10. To prevent this, a support ring 28 is arranged on the piston body 20 at the piston rod side end of the piston 3, as shown in... Figure 2 , 3As exemplarily illustrated in 4 and 5, the support ring 28 is arranged within a support ring groove 29 in the outer peripheral surface of the piston body 20, wherein the support ring groove 29 is axially spaced along the piston axis KA from the sleeve groove 24 for the at least one piston sleeve 21, which is closest to the piston rod end of the piston 3. The support ring 28 protrudes radially from the outer peripheral surface of the piston body 20.

[0080] The support ring 28 is designed to withstand any residual lateral forces acting on the cylinder sliding surface 10 during piston 3 operation, and in particular to ensure that the piston body 20 rotates concentrically along the cylinder axis (which coincides with the piston axis KA). Therefore, the support ring 28 rests against the cylinder sliding surface 10 during piston 3 operation and is advantageously made of a tribologically advantageous and wear-resistant material, particularly a polymeric material. Because the support ring 28 is located on the piston rod side end (i.e., the side away from pressure during piston 3 operation) and only needs to withstand a small lateral force, wear on the support ring 28 is minimal. Furthermore, the support ring 28 can be designed to be pressure-balanced, so that no force is generated on the support ring 28 during piston 3 operation due to any possible pressure differential acting on it. This also helps to keep wear on the support ring 28 minimal.

[0081] The outer diameter of the support ring 28 can be equal to the inner diameter of the cylinder sliding surface 10, and can be less than or greater than the outer diameter of the at least one piston sleeve 21.

[0082] To prevent the piston body 20 from contacting the cylinder sliding surface 10 during piston 3 operation, the maximum outer diameter of the piston body 20 (at least the portion of the piston body 20 that slides in the cylinder sliding surface 10 during operation) is always smaller than the outer diameter of the piston sleeve 21.

[0083] Reference Figure 4 Further possible features of the piston 3 of the present invention will be described. These additional features are independent of the design of arranging the piston 3 as a floating support on the piston rod 4.

[0084] According to Figure 4In the design of the piston 3, a throttling groove 30, which is in the form of an annular groove, is provided on the outer peripheral surface 27. This throttling groove 30 is optional, and only one throttling groove 30 may be provided. If a throttling groove is present, the at least one throttling groove 30 is used to capture and collect particles that may be present between the outer peripheral surface 27 of the piston sleeve 21 and the cylinder sliding surface 10. This prevents scratches from forming on the cylinder sliding surface 10 and / or the outer peripheral surface 27 of the piston sleeve 21, or prevents such particles from scraping away debris. Furthermore, the throttling groove 30 can also be designed to create a throttling effect similar to that in a non-contact labyrinth seal, which improves the non-contact sealing performance on the piston sleeve 21 during piston 3 operation. Turbulence of the compressed medium flowing along the outer peripheral surface 27 is generated within the annular groove 30, which leads to losses, which in turn reduce the operating pressure. This throttling effect is particularly good when the ratio of the sealing gap DS to the radial groove depth of the annular groove is at least 1.5, preferably at least 2.

[0085] In such Figure 3 In the design shown where the piston body has multiple piston sleeves 21, due to the axial separation of the sleeve grooves 24, a groove is automatically formed between two adjacent piston sleeves 21. Figure 2 The throttling groove in the middle has the same effect as the throttling groove 30. Nevertheless, according to... Figure 3 In the design, at least one throttling groove 30 may be additionally provided on the outer peripheral surface 27 of at least one piston sleeve.

[0086] According to Figure 4 In the design, a clearance ring 43, preferably made of a polymer such as PTFE, is provided in the axial clearance AS between the piston body 20 and the piston sleeve 21 on the axial end side away from the crankshaft 4. This clearance ring 43 is designed so that the piston sleeve 21 is fitted into the sleeve groove 24 without axial clamping or with a very small preload less than 10% of the unit area pressure applied by the final compression pressure, thus preventing the piston sleeve 21 from losing its floating support in the sleeve groove 24. This can be achieved by designing the clearance ring 43 to be elastic, wherein the elastic force acting along the piston axis KA direction is negligible. If the clearance ring 43 is designed to be rigid, its axial length can be at most equal to the axial length of the axial clearance AS, and preferably less than the axial length of the axial clearance AS. Furthermore, the clearance ring 43 is intended to ensure that the sleeve clearance HS connects to the outer circumferential surface of the piston 3 or the outer circumferential surface 27 of the piston sleeve 21 so that the cylinder pressure p z It can reach the sleeve clearance HS. For this purpose, the clearance ring 43 is structurally designed to form this connection, or to form this connection by means that the axial length of the clearance ring 43 is less than the axial length of the axial clearance AS.

[0087] According to Figure 2This type of gap ring 43 can also be set in the design. According to... Figure 3 In the design, a clearance ring 43 can be provided in at least one axial clearance AS between the piston sleeve 21 and the piston body 20, preferably in all piston sleeves 21.

[0088] According to Figure 4 In the design, an elastic sealing element 44 is also provided between the piston rod side end face 45 of the piston sleeve 21 and the axial end of the piston rod side of the sleeve groove 24. This sealing element 44, according to... Figure 4 In the design, it is arranged in a groove on the axial end of the sleeve groove 24, but it can also be simply inserted between the piston sleeve 21 and the axial end of the sleeve groove 24.

[0089] Sealing element 44 improves the seal on the sleeve clearance HS, where cylinder pressure p exists. z As explained earlier, the cylinder pressure p z The piston sleeve 21 is pressed against the axial end of the sleeve groove 24 on the piston rod side. This also seals the sleeve clearance HS, preventing a leakage path for the compressed medium from forming through this clearance HS. The sealing element 44 can improve or make this seal more reliable. In particular, the sealing element 44 can also compensate for any possible tilt of the piston sleeve 21 relative to the piston axis KA, and prevent a leakage path from opening at the axial end of the piston sleeve 21 on the piston rod side if such tilt occurs.

[0090] As an alternative, the resilient sealing element 44 can also be arranged between the cylinder-side axial end of the sleeve groove 24 and the piston sleeve 21. In this design, the clearance ring 43 can be omitted or replaced by the resilient sealing element 44. The resilient sealing element 44 at the opposite axial end (as in...) Figure 4 (As shown in the diagram) This structural design can also be omitted, or intentionally designed to allow gas passage. The effect of this design is that the pressure acting on the inner diameter of the piston sleeve 21 is less than the pressure acting on the outer diameter of the piston sleeve 21, thus the piston sleeve 21 is slightly compressed by the resulting pressure difference. This, while increasing the sealing clearance DS during operation, also... Figure 4 The design shown increases leakage compared to the previous one, but its advantage is that it ensures that, under extreme conditions, the pressure on the inner diameter of the piston sleeve 21 will not rise to a level higher than the pressure on the outer diameter of the piston sleeve 21, thereby preventing the piston sleeve 21 from deforming radially outward and preventing the piston sleeve 21 from contacting the cylinder sliding surface 10.

[0091] The sealing element 44 as described above can also be provided according to... Figure 2 In the design. Based on Figure 3In the design, a sealing element 44 can be provided in at least one piston sleeve 21, preferably in all piston sleeves 21. This sealing element 44 can serve as a supplement to the clearance ring 43, as in... Figure 4 However, it can also be set without the gap ring 43.

[0092] The materials of the cylinder sliding surface 10 and the piston sleeve 21 are selected such that thermal deformation caused by different coefficients of thermal expansion, together with corresponding deformation caused by the present pressure, can maintain a functional sealing gap DS. That is, it is ensured that the sealing gap DS does not become too large or too small due to thermal expansion and deformation caused by load. In this case, if the coefficients of thermal expansion of the materials used differ by a maximum of 1 x 10⁻⁶, then... -6 A coefficient of thermal expansion of around 1 / K is advantageous. Preferred, but not mandatory, is that the material of the cylinder sliding surface 10 has a greater coefficient of thermal expansion than the material of the piston sleeve 21. This prevents the piston sleeve 21 from expanding more than the cylinder sliding surface 10, thus reliably preventing thermal closure of the sealing gap DS between the piston sleeve 21 and the cylinder sliding surface 10. Possible material combinations to ensure this are: the cylinder sliding surface 10 is ceramic and the piston sleeve 21 is Invar alloy (an iron-nickel alloy with a very low coefficient of thermal expansion); or the cylinder sliding surface 10 is steel and the piston sleeve 21 is ceramic. Similarly, in a suitable design, a steel piston sleeve 21 can be combined with a ceramic cylinder sliding surface 10. The design of two ceramic contact mating parts is also possible.

[0093] Figure 9 The cylinder Z1 of a reciprocating compressor 1 is shown, in which the piston 3 of the present invention, together with the piston rod 4, reciprocates. In this design, the cylinder chamber K1 in the cylinder Z1 is defined by a cylinder head 60, in which an intake valve SV1 and a pressure valve DV1 are also arranged. The piston 3 with the piston rod 4 can be referenced as above. Figures 2 to 8 The design is described below. In this embodiment, the cylinder sliding surface 10 is constructed on the cylinder liner 61, which is arranged in the cylinder block Z1, which conforms to the usual construction design.

[0094] According to Figure 9In its structural design, the cylinder block Z1 is designed to be cooled. A temperature regulating pipe 62 (through which the temperature regulating medium flows) is provided within the cylinder housing of the cylinder block Z1 and around the cylinder sliding surface 10 or around the cylinder sleeve 41. The temperature regulating pipe 62 and the temperature regulating medium allow the cylinder sliding surface 10 to be adjusted to a specific temperature, thereby controlling the thermal expansion of the cylinder sliding surface 10 and thus controlling the sealing gap DS between the cylinder sliding surface 10 and the at least one piston sleeve 21. For this purpose, a temperature sensor 63 can also be provided in the cylinder block Z1 to detect the temperature in the area of ​​the cylinder sliding surface 10. This temperature sensor 63 can be integrated into a closed-loop control circuit so that the temperature in the area of ​​the cylinder sliding surface 10 can be controlled by adjusting the temperature of the temperature regulating medium. The temperature regulating medium can be controlled for cooling and heating. Cooling is particularly necessary when the cylinder block Z1 heats up due to the compression of the compression medium during operation of the reciprocating compressor 1. Heating of the temperature regulating medium may also be necessary during the start-up of the reciprocating compressor 1 or during operation at low ambient temperatures. The temperature of the temperature-regulating medium can also be adjusted according to the current sealing gap DS, so that the sealing gap DS is maintained within the desired range by the thermal expansion of the cylinder sliding surface 10 and the piston sleeve 21. The sealing gap DS can also be measured for this purpose, for example, using ultrasonic testing.

Claims

1. A non-contact sealed piston (3) with a piston rod (4) for a reciprocating compressor (1), characterized in that, The piston (3) includes a piston body (20) and at least one piston sleeve (21). At least one sleeve groove (24) is provided on the piston body (20). The at least one piston sleeve (21) is arranged in the sleeve groove in a floating, radially transverse direction to the piston axis (KA) such that: a sleeve clearance (HS) is provided between the inner circumferential surface (22) of the at least one piston sleeve (21) and the outer circumferential surface (23) of the sleeve groove (24), and the axial sleeve length (HL) of the at least one piston sleeve (21) along the piston axis (KA) is less than the axial groove length (NL) of the at least one sleeve groove (24) to form an axial clearance (AS); in the piston body (20)... A piston rod groove (40) is constructed on the axial end of the piston rod side, and one axial end of the piston rod (4) is arranged in the piston rod groove. One axial end face (41) of the piston rod (4) axially abuts against the bottom surface (42) of the piston rod groove (40), and the piston body (20) is arranged on the piston rod (4) in a radially floating manner, that is, a piston rod gap (KS) is provided between the piston rod groove (40) and the piston rod (4) in the radial direction; and a support ring groove (29) is provided on the outer circumferential surface of the piston body (20) on the axial end of the piston rod side of the piston body (20), and a support ring (28) is arranged in the support ring groove, which protrudes radially from the outer circumferential surface of the piston body (20).

2. The non-contact sealing piston (3) with piston rod (4) according to claim 1, characterized in that, The piston body (20) is provided with a plurality of sleeve grooves (24) spaced axially along the piston axis (KA), wherein a piston sleeve (21) is arranged in a radially floating support in each sleeve groove (24).

3. The non-contact sealing piston (3) with piston rod (4) according to claim 1 or 2, characterized in that, At least one throttling groove (30) is provided in the outer peripheral surface (27) of at least one piston sleeve (21).

4. The non-contact sealing piston (3) with piston rod (4) according to claim 1, characterized in that, A plurality of throttling grooves (30) spaced apart from each other are provided on the outer peripheral surface (27) of the at least one piston sleeve.

5. The non-contact sealing piston (3) with piston rod (4) according to any one of claims 1 to 4, characterized in that, The support ring groove (29) is arranged axially between the axial end of the piston rod side of the piston body (20) and the axial end of the sleeve groove (24) closest to the axial end of the piston rod side of the piston body (20).

6. The non-contact sealing piston (3) with 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 groove (40) are designed to be convex and arched.

7. The non-contact sealing piston (3) with piston rod (4) according to any one of claims 1 to 6, characterized in that, The piston rod (4) is elastically connected to the piston body (20) in the piston rod groove (40) by a radial or axial elastic connecting element (46).

8. The non-contact sealing piston (3) with piston rod (4) according to claim 7, characterized in that, An elastic ring is provided between the piston rod (4) and the piston rod groove (40) as a radial elastic connecting element (46).

9. The non-contact sealing piston (3) with piston rod (4) according to claim 8, characterized in that, A ring receiving groove (47a, 47b) is provided in the piston rod groove (40) and on the piston rod (4), and the elastic ring is arranged in the ring receiving groove (47a, 47b).

10. The non-contact sealing piston (3) with piston rod (4) according to claim 7, characterized in that, The axial elastic connecting element (46) is designed as an elastic body, wherein a connecting element groove (48) is provided on the axial end of the piston rod (4) facing the piston (3), the axial elastic connecting element (46) is arranged in the connecting element groove, wherein an axially protruding threaded pin (49) is provided on each of the two axial ends of the axial elastic connecting element (46), and one threaded pin (49) is screwed into a threaded hole (50) on the piston body (20), and the opposite threaded pin (49) is screwed into a threaded hole (50) on the piston rod (4).

11. The non-contact sealing piston (3) with piston rod (4) according to any one of claims 1 to 6, characterized in that, A clamping sleeve (52) is fastened to the piston body (20) at the axial end of the piston rod side of the piston body (20), wherein a shoulder (53) protrudes radially inward at the axial end of the clamping sleeve (52) away from the piston (3), wherein a radial rib (54) is provided on the piston rod (4) along the axial direction between the shoulder (53) and the piston rod groove (40), wherein a spring element (55) is arranged between the radial rib (54) and the shoulder (53), the spring element pressing the piston rod (4) toward the piston (3).

12. The non-contact sealing piston (3) with piston rod (4) according to any one of claims 1 to 11, characterized in that, An elastic sealing element (44) is arranged between the piston rod side end face (45) of the piston sleeve (21) and the axial end of the piston rod side of the sleeve groove (24).

13. The non-contact sealing piston (3) with piston rod (4) according to any one of claims 1 to 12, characterized in that, A clearance ring (43) is arranged in the axial clearance (AS) between the piston body (20) and the piston sleeve (21) on the axial end side away from the piston rod (4), which establishes the connection between the sleeve clearance (HS) and the outer peripheral surface of the piston (3).

14. A cylinder for a reciprocating compressor (1), wherein a non-contact sealed piston (3) with a piston rod (4) according to any one of claims 1 to 13 is reciprocally arranged in the cylinder (Z1), and the piston (3) at least partially defines a compression chamber (K1) constructed in the cylinder (Z1), wherein a sealing gap (DS) is provided between the outer peripheral surface (27) of the at least one piston sleeve (21) or the plurality of outer peripheral surfaces (27) of the plurality of piston sleeves (21) and the cylinder sliding surface (10) of the cylinder (Z1).

15. The cylinder block according to claim 14, characterized in that, A temperature regulating pipe (62) is arranged around the cylinder sliding surface (10) inside the cylinder body (Z1), and the temperature regulating medium flows through the temperature regulating pipe.

16. The cylinder block according to claim 14 or 15, characterized in that, As materials, the cylinder sliding surface (10) is made of ceramic and the piston sleeve (21) is made of Invar alloy, or the cylinder sliding surface (10) is made of steel and the piston sleeve (21) is made of ceramic, or the cylinder sliding surface (10) is made of ceramic and the piston sleeve (21) is made of steel, or the cylinder sliding surface (10) is made of ceramic and the piston sleeve (21) is made of ceramic.

17. A piston compressor comprising at least one cylinder block (Z1) according to any one of claims 14 to 16, characterized in that, The piston rod (4) is connected to the piston drive mechanism (12) of the piston compressor (1) at its axial end away from the piston (3).

Images