A piston that seals without contact, having a piston rod.

Abstract

The piston for use in piston compressors for high-pressure applications is manufactured with a piston (3) as a cup-shaped piston body (20), the piston body (20) having a closed bottom region (21) and a hollow cylindrical region (22) that follows the bottom region (21) in the axial direction, the axial end face (24) of the piston rod (4) facing the piston (3) and / or the axial bottom surface (25) of the bottom region (21) facing the piston rod (4) are curved in a convex shape, and the axial direction of the piston rod (4) facing the piston (3) The outward end face (24) is positioned in the hollow cylindrical region (22) of the piston body (20) and contacts the bottom surface (25) of the bottom region (21) at a contact point (K). A piston rod gap (KS) is provided between the inner circumferential surface (26) of the hollow cylindrical region (22) and the piston rod (4) in a radial direction perpendicular to the piston axis (KA). The contact point (K) is located along the piston axis (KA) near the axial compression surface (28) of the bottom region (21) of the piston body (20) that faces the bottom surface (25), rather than the center of mass (M) of the piston body (20).

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 also relates 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 a 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, which are modified with respect to tribological properties by filler substances 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] International Publication No. 98 / 31936 describes a non-lubricated piston compressor comprising a metal cylinder sliding surface and a piston having a casing made of plastic material. Parallel movement of the piston within the cylinder is possible via a piston rod support that allows for radial movement of the piston rod.

[0009] To achieve a useful service life for industrial applications, the high-pressure pistons of piston compressors (for a final compression pressure <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.

[0010] 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 outer diameter of the piston due to the pressure.

[0011] 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.

[0012] 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.

[0013] 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.

[0014] 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.

[0015] 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.

[0016] In particular, in high-pressure applications of piston compressors where the final compression pressure is >350 bar, deformation of the cylinder and / or piston can occur due to the pressure, 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.

[0017] 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.

[0018] For these reasons, although gap seals for piston compressors are theoretically possible, no industrial implementation of such gap seals in piston compressors is known.

[0019] 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.

[0020] However, using a piston compressor even in high-pressure applications should be preferable because it is a well-established and particularly simple type of compressor. [Overview of the project] [Problems that the invention aims to solve]

[0021] Therefore, the object of the subject invention is to provide a piston for a piston compressor that enables the use of a piston compressor for high-pressure applications.

[0022] This is made possible by the non-contact sealing piston described in claim 1. A non-contact gap seal can be realized by a piston manufactured according to the present invention. Based on the embodiments, in particular, based on the floating support of the piston and based on the distance of the center of mass of the piston body to the contact point between the piston body and the piston rod, the contact force that may occur between the piston and the cylinder can be adjusted, so as to ensure reliable operation under small sealing gaps. Accordingly, a 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 final compression pressures may occur, particularly in hydrogen applications. Accordingly, it becomes possible for the first time to use a piston compressor for such final compression pressures.

[0023] The stabilizing effect of the piston body with respect to the inclined position is particularly favorable when the ratio of the contact distance between the contact point and the center of mass to the contact distance between the axial compression surface and the center of mass is less than 1, preferably between 0.5 and 0.8.

[0024] If at least one throttling groove is provided on the outer circumferential surface of the piston body, preferably multiple throttling grooves spaced apart in the axial direction, particles that may be present between the outer circumferential surface of the piston body and the cylinder sliding surface can be trapped. Consequently, scratches on the cylinder sliding surface and / or the outer circumferential surface of the piston body can be prevented. Furthermore, such throttling grooves can also improve the sealing effect.

[0025] When the piston rod is elastically coupled to the piston body via a radially or axially elastic coupling member in a hollow cylindrical region, it is possible to prevent the piston body from lifting off 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 towards the piston rod.

[0026] Next, by way of example, schematically and without limitation, the subject invention will be described in detail while referring to FIGS. 1 to 8 showing preferred embodiments of the present invention. The drawings show the following.

Brief Description of the Drawings

[0027] [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 piston for non-contact sealing according to the present invention having a radially elastic coupling member. [Figure 5] Showing a piston for non-contact sealing according to the present invention having a radially elastic coupling member. [Figure 6] Showing a piston for non-contact sealing according to the present invention having an axially elastic coupling member. [Figure 7] Showing a piston for non-contact sealing according to the present invention having an initial stress unit. [Figure 8] Showing a cylinder of a piston compressor having a piston according to the present invention.

Modes for Carrying Out the Invention

[0028] 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.

[0029] 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.

[0030] 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.

[0031] 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.

[0032] 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.

[0033] 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.

[0034] 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.

[0035] 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.

[0036] In the embodiment shown in Figure 1, piston rings 15 are positioned on the piston 3 to contact the cylinder sliding surfaces 10 of cylinders Z1 and Z2 in order to seal the respective cylinder chambers K1 and k2.

[0037] Figure 2 shows an embodiment of a contactless 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 cylinder Z1 of the piston compressor 1. Other components of the piston compressor 1, such as the entire piston transmission 12, are not shown for simplification. If the piston rod 4 in the piston compressor 1 is manufactured in multiple parts (for example, shown in Figure 6), the “piston rod” includes at least the portion of the piston rod 4 that, in combination with the piston 3, 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.

[0038] A non-contact gap seal with a sealing gap DS is provided to seal between the piston 3 and the cylinder Z1. As is common in piston compressors 1, the piston 3 moves in a reciprocating motion within the cylinder chamber K1 of the cylinder Z1. The piston drive unit 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 a non-contact gap seal, the piston 3 is manufactured to act only in a single action, so the piston 3 only separates the cylinder chamber K1 on the side opposite to the piston drive unit 12.

[0039] The piston 3 according to the present invention consists of a cup-shaped piston body 20 having a closed bottom region 21 and a hollow cylindrical region 22 that follows the bottom region 21 in the axial direction. Accordingly, the piston 3 is axially separated by a compression surface 28 at the axial end facing the piston transmission device 12. When the piston 3 is used in the cylinder Z1 of the piston compressor 1, the compression surface 28 at least partially separates the compression chamber K1 of the cylinder Z1.

[0040] The piston body 20 is preferably manufactured as a single unit, that is, the bottom region 21 and the hollow cylindrical region 22 are made from a single piece. Alternatively, the piston body 20 may be manufactured in multiple parts, and the individual parts of the piston body 20 are appropriately joined together.

[0041] At least one sealing region DB extending in the axial direction is provided on the outer circumferential surface 27 of the piston body 20. The sealing region DB can extend in the direction of the piston axis KA along the entire axial length of the piston body 20. Alternatively, multiple such sealing regions DB may be provided, each spaced axially apart from the others in the direction of the piston axis KA.

[0042] When piston 3 is used in piston compressor 1, a sealing gap DS is formed between the outer circumferential surface 27 of the piston body 20 in the sealing region DB and the cylinder sliding surface 10 of the cylinder Z1 in which piston 3 moves, thereby achieving a non-contact gap seal. The sealing gap DS is typically between 1 μm and 50 μm in width to constitute a non-contact gap seal. The sealing gap DS is formed between the outer circumferential surface 27 of the piston body 20 in the sealing region DB and the inner diameter D of the cylinder sliding surface 10 of the cylinder Z1. Z This is defined by the minimum radial distance that occurs when the piston 3 is concentrically positioned within the cylinder Z1.

[0043] The outer diameter D of the outer circumferential surface 27 of the piston body 20 in the sealing region DB is preferably constant. At a minimum, the outer diameter D of the piston body 20 is such that a sealing gap DS is created along the sealing region DB and throughout the entire sealing region. The axial region of the piston body 20 outside the sealing region DB has an outer diameter that does not create a sealing gap DS when the piston 3 is used in any case.

[0044] In the embodiment shown in Figure 2, tapered axial regions are provided at both axial ends, but the tapered junction regions may be provided at only one end, or at neither end. The sealing region DB extends over the remaining axial length of the piston body 20.

[0045] In particular, one advantage of the piston 3 according to the present invention, shown in Figure 2, is that it generates almost no additional dead space. This is because, apart from a very small sealing gap DS, no gaps are provided.

[0046] Based on the piston drive system 12, a lateral force may be applied to the piston 3 via the connection of the piston rod 4 to the piston 3, perpendicular to the piston axis KA. Due to the small sealing gap DS, such a lateral force causes contact between the cylinder sliding surface 10 and the piston 3. If the resulting contact force becomes too large, this causes excessive wear on both frictional mating parts, and as a further consequence, failure of each component. Therefore, the lateral force transmitted between the piston rod 4 and the piston 3 must be reduced as much as possible. For this reason, the piston 3 is positioned to be floatingly supported on the piston rod 4, radially perpendicular to the piston axis KA. "Floatingly supported" means that the piston 3 and the piston rod 4 are freely movable relative to each other radially with respect to the piston axis KA (except for any frictional forces that may act on them in the design). Free movement is, of course, only possible within the range of the possible movement margins defined in the design. It is preferable that the piston rod 4 can move freely radially relative to the piston 3 without contacting the piston 3 radially. The degree of radial movement of the piston rod 4 during operation naturally depends on the embodiment of the piston transmission device 12, but can be assumed to be known based on the known design of the piston compressor 1.

[0047] To achieve such radially floating support, the axial end of the piston rod 4 facing the piston 3 is positioned in the hollow cylindrical region 22 of the piston body 20 according to the present invention and abuts against the bottom region 21 (facing the compression surface 28 in the axial direction) at contact point K. Accordingly, the axial end face 24 of the piston rod 4 facing the piston 3 and the bottom surface 25 of the bottom region 21 abut each other at contact point K. Radially floating support with respect to the piston axis KA is achieved by providing a radial piston rod gap KS between the piston rod 4 and the inner circumferential surface 26 of the hollow cylindrical region 22 of the piston body 20. The piston rod gap KS may be formed by the minimum inner diameter of the hollow cylindrical region 22 being larger than the maximum outer diameter of the axial end of the piston rod 4 positioned within the hollow cylindrical region 22. The piston rod gap KS is defined by the minimum radial distance between the piston rod 4 and the hollow cylindrical region 22, which occurs when the piston rod 4 and the piston body 20 are arranged concentrically.

[0048] The piston rod gap KS is preferably selected within the range of 0.1 mm to 10 mm for design purposes.

[0049] Therefore, the piston rod 4 can move radially within the piston body 20 by the amount of the piston gap KS. Consequently, the piston body 20 is also radially movable within the cylinder Z1 by up to twice the amount of the sealing gap DS when the piston 3 is operating.

[0050] Acting cylinder pressure p Z This also acts to center the piston body 20 within the cylinder Z1 when the piston 3 is in operation. Based on the hydrodynamic effect, pressure is generated on the side where the sealing gap DS between the piston body 20 and the cylinder sliding surface narrows, which in turn causes the piston body 20 to move back to its central concentric position.

[0051] The piston rod 4 may have a slight tilt position and / or slight concentricity defect with respect to the piston axis KA during operation. If such tilt position and / or concentricity defect of the piston rod 4 is fixedly transmitted to the piston body 20, a high contact force will be generated in the cylinder Z1 between the piston body 20 and the cylinder sliding surface 10, which will result in significant wear. To prevent this, the axial end face 24 of the piston rod 4 facing the piston, and / or the axial bottom surface 25 of the bottom surface region 21 are curved convexly (outward) according to the present invention, as shown in Figure 3. In the embodiment of Figure 3, the axial end face 24 of the piston rod 4 facing the piston and the axial bottom surface 25 of the bottom surface region 21, which face each other, are curved convexly, but it is sufficient if only one of these surfaces is manufactured in a convex shape.

[0052] The convex curvature of at least one of these axial surfaces allows for compensation of angle defects in the piston rod 4. Consequently, the possible inclination position of the piston rod 4 relative to the piston axis KA is not transmitted to the piston body 20, and the piston body 20 is kept concentric within the cylinder Z1.

[0053] The piston body 20 operates under the cylinder pressure p acting within the cylinder Z1. Z This causes pressure toward the piston rod 4, thereby generating a contact force at the contact point K that produces a lateral frictional force.

[0054] Such frictional forces depend on the unavoidable pressure difference in piston 3, as well as the coefficient of friction between the two contact surfaces, and can be large in high-pressure applications. By selecting appropriate geometry and materials within the capabilities of a craftsman skilled in the art, aiming to minimize these coefficients of friction, the resulting lateral forces required for centering can be controlled, thereby ensuring that the necessary maneuverability is maintained.

[0055] Based on known design configurations of the piston body 20, a specific mass center M of the piston 3 is derived (see Figures 2 and 3), and its position can be assumed to be known. From there, based on known design embodiments of the piston body 20, an axial contact gap A of the contact point K with respect to the mass center M is generated along the piston axis KA. According to the present invention, the piston body 20 is manufactured such that the position of the contact point K along the piston axis KA is located in the direction of the piston axis KA, closer to the compression surface 28 of the piston 3 than the mass center M.

[0056] Due to the position of the contact point K relative to the center of mass M, during the compression stroke in which the maximum force is generated (i.e., when the piston 3 is moving towards top dead center OT), the contact force (compression force) acting on the piston 3 between the piston rod 4 and the piston body 20 when the piston body 20 is in an inclined position exerts a return action on the piston body 20, and this return action attempts to return the piston body 20 to a concentric position. Consequently, the cylinder pressure p Z The torque acting on the piston body 20, which may occur based on the compression surface 28, is also reduced. Accordingly, the position of the contact point K along the piston axis KA according to the present invention improves the stability of the angular fidelity of the piston body 20 (relative to the piston axis). This is particularly important in high-pressure pistons where the final pressure is up to 1000 bar, and in gap seals, because even a slight inclination can lead to high contact force and large wear between the piston body 20 and the cylinder sliding surface 10 of the cylinder Z1. In contrast, the contact point K on the opposite side of the center of mass M has an destabilizing effect and further intensifies the inclination.

[0057] Typically, since the piston 3 is manufactured to be rotationally symmetric, the center of mass M will be located on the piston axis KA. In cases where the center of mass M is not located on the piston axis KA, the position of the radial projection of the center of mass M onto the piston axis KA with respect to the compression surface 28 serves as a guide. Similarly, when the contact point K is not located on the piston axis KA (for example, in the embodiment of Figure 6), the position of the radial projection of the contact point K onto the piston axis KA is considered.

[0058] For stabilization purposes, the ratio A / AF between the contact spacing A and the contact surface spacing AF between the center of mass M and the compression surface 28 is preferably as large as possible along the piston axis KA, provided that this ratio A / AF is always less than 1. Preferably, this ratio A / AF = 0.5 to 0.8.

[0059] With respect to the outer diameter D of the outer circumferential surface 27 of the piston body 20, the geometric ratio in the piston 3 is selected such that AF / D = 0.7 to 1.5 and A / D = 0.35 to 1.2 (which leads to the above ratio A / AF = 0.5 to 0.8). The axial length L of the piston body 20 with respect to the outer diameter D also follows a preferred size ratio, namely L / D = 1.5 to 3.

[0060] Figures 4 and 5 show alternative floating supports for the piston 3 on the piston rod 4. Figure 5 is an enlarged view of the floating support in Figure 4 for illustrative purposes. Figure 5 also shows the convex curvature of the axial end face 24 of the piston rod 4.

[0061] In the embodiments shown in Figures 4 and 5, the piston rod 4 and the piston 3 are elastically coupled radially to each other via a radially elastic coupling member 46. The piston rod gap KS is not affected by this. In the illustrated embodiments, an elastic ring, such as an O-ring, is positioned radially between the piston rod 4 and the hollow cylindrical region 22 of the piston body 20 as the radially elastic coupling member 46. For this purpose, ring housing grooves 47a and 47b are provided on the outer circumferential surface of the piston rod 4 and the inner circumferential surface of the hollow cylindrical region 22, respectively, as shown in Figures 4 and 5. When the piston rod 4 is positioned in the hollow cylindrical region 22, the elastic ring is housed and held within both of these ring housing grooves 47a and 47b. However, the elasticity of the radially elastic coupling member 46, and optionally the dimensions of the ring housing grooves 47a and 47b, are selected such that the force generated by the coupling member 46 is small enough to be negligible in the radial direction, and thereby the floating support of the piston 3 in the radial direction is not impaired.

[0062] When the piston compressor 1 is operating, the cylinder pressure p acts inside the cylinder. Z This force acts on the piston 3, and this force also causes the piston 3 to move together with the piston rod 4 when the piston rod 4 returns from its top dead center OT. 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.

[0063] 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 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.

[0064] 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 6. 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 floating on the piston rod 4.

[0065] 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 rods 4 with the piston rod gap KS in the hollow cylindrical region 22 remains unchanged. The bottom surface 25 of the bottom region 21 and the end surface 24 of the piston rod 4 similarly contact each other at the contact point K. In this embodiment, annular contact is made around the piston axis KA based on the axially elastic coupling member 46 and the convexly curved end surface 24 / or bottom surface 25. The contact point K is considered to be the radial projection onto the piston axis KA. The axially elastic coupling member 46 is merely an additional addition and does not impair the stabilizing effect of the contact point K described above. A coupling member notch 48 may be provided at the axial end of the piston rod 4 for the axially elastic coupling member 46, and the axially elastic coupling member 46 may be positioned therein. Threaded pins 49 protrude from both ends of the axially elastic coupling member 46. One of the threaded pins 49 is screwed into a threaded hole 50 of the piston body 20, and the other threaded pin 49 is screwed into a threaded hole 50 of the piston rod 4.

[0066] Figure 6 also shows an embodiment of the piston rod 4, which consists of multiple parts, with each part of the piston 4 being screwed together with the others.

[0067] Figure 7 illustrates another embodiment of the arrangement of a piston 3 that is radially floatingly supported on a piston rod 4. In this embodiment as well, the axial end of the piston rod 4 is positioned in the hollow cylindrical region 22 of the piston body 20 with a piston rod gap KS in the lateral direction. Similarly, the bottom surface 25 of the bottom region 21 and the end surface 24 of the piston rod 4 abut each other in the axial direction at the contact point K.

[0068] In the embodiment shown in Figure 7, the piston rod 4 is actively and additionally pressed axially toward the bottom surface 25 of the bottom region 21 by the initial stress unit 51, without impairing the floating support, particularly the piston rod gap KS.

[0069] 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, specifically the hollow cylindrical region 22, by screw connection, or by bonding, press-fitting, or other methods, for example, as shown in Figure 7. 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 body 20. 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, there is a spring member 55 that applies initial stress in the direction of the piston 3 to the radial web 54 and consequently to the piston rod 4. Consequently, a specific axial contact force can be guaranteed at the contact surface 23 between the piston rod 4 and the piston 3, which is particularly preferable when changing the direction of force. This is because contact loss between the piston rod 4 and the piston 3 can be prevented during the unpressurized starting phase of the piston compressor.

[0070] In the embodiments of the piston 3 shown in Figures 4, 6, and 7, a throttling groove 30 is provided on the outer circumferential surface 27 of the piston 3 in the form of a circumferential groove. These throttling grooves 30 are optional, and only one throttling groove 30 may be provided. In the embodiments of Figure 2 or Figure 3, at least one throttling groove 30 may be provided.

[0071] 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 body 20 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 body 20, 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 of the piston body 20 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 compressive medium flowing along the outer circumferential surface 27, which leads to loss, and this loss reduces the pressure on which it acts. Such a throttling effect is particularly effective when the ratio of the sealing gap DS to the radial groove depth of the throttling groove 30 is at least 1.5, preferably at least 2.

[0072] The materials of the cylinder sliding surface 10 and the piston body 20 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 body 20. This prevents the piston body 20 from expanding more than the cylinder sliding surface 10, and consequently ensures that thermally-induced closure of the sealing gap DS between the piston 3 and the cylinder sliding surface 10 is prevented. Possible material combinations that guarantee 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 body 20, or steel for the cylinder sliding surface 10 and ceramic for the piston body 20. Similarly, in a preferred design, a piston body 20 made of steel can be combined with a ceramic cylinder sliding surface 10. Embodiments of both components made of ceramic material are also possible.

[0073] Figure 8 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 6. 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.

[0074] In the configuration shown in Figure 8, 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 61, 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, and consequently, the thermal expansion of the cylinder sliding surface 10, and the sealing gap DS between the cylinder sliding surface 10 and the piston body 20, can also be controlled. 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 depending on 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 body 20. 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) is manufactured as a cup-shaped piston body (20), the piston body (20) having a closed bottom region (21) and a hollow cylindrical region (22) that follows the bottom region (21) in the axial direction, the axial end face (24) of the piston rod (4) facing the piston (3) and / or the axial bottom surface (25) of the bottom region (21) facing the piston rod (4) are curved in a convex shape, and the axial end face (24) of the piston rod (4) facing the piston (3) is A piston (3) that seals in a non-contact manner, having a piston rod (4) which is positioned in the hollow cylindrical region (22) of the piston body (20) and in contact with the bottom surface (25) of the bottom region (21) at a contact point (K), wherein a piston rod gap (KS) is provided between the inner circumferential surface (26) of the hollow cylindrical region (22) and the piston rod (4) in a radial direction laterally with respect to the piston axis (KA), and the contact point (K) along the piston axis (KA) is located near the axial compression surface (28) of the bottom region (21) of the piston body (20) that faces the bottom surface (25), rather than the center of mass (M) of the piston body (20).

2. A piston (3) for non-contact sealing having a piston rod (4) according to claim 1, characterized in that the ratio A / AF of the contact distance (A) between the contact point (K) and the center of mass (M) and the contact surface distance (AF) between the axial compression surface (28) and the center of mass (M) is less than 1, preferably between 0.5 and 0.

8.

3. A piston (3) that seals in a non-contact manner, 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 the piston body (20), preferably a plurality of throttling grooves (30) spaced apart from each other in the axial direction.

4. A piston (3) for non-contact sealing having a piston rod (4) according to any one of claims 1 to 3, characterized in that the piston rod (4) is elastically coupled to the piston body (20) in the hollow cylindrical region (22) via a radial or axial elastic coupling member (46).

5. A piston (3) that seals in a non-contact manner, having the piston rod (4) according to claim 4, characterized in that an elastic ring (46) is provided between the piston rod (4) and the piston body (20) as the radially elastic coupling member.

6. A piston (3) that seals in a non-contact manner, having a piston rod (4) according to claim 5, characterized in that ring housing grooves (47a, 47b) are provided in the piston body (20) and the piston rod (4), respectively, and the elastic ring is arranged in the ring housing grooves (47a, 47b).

7. The piston (3) having a piston rod (4) according to claim 4, 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, threaded pins (49) protruding in the axial direction are provided at both axial ends of the axially elastic coupling member (46), one of the threaded pins (49) is screwed into a threaded hole (50) of the piston body (20), and the threaded pin (49) opposite to it is screwed into a threaded hole (50) of the piston rod (4).

8. A piston (3) for non-contact sealing having a piston rod (4) as described in any one of claims 1 to 3, 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) facing away from the piston (3), a radial web (54) is provided on the piston rod (4) between the shoulder (53) and the piston body (20) in the axial direction, and a spring member (55) is arranged between the radial web (54) and the shoulder (53) to press the piston rod (4) toward the piston (3).

9. 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 8 is arranged in a reciprocating motion within the cylinder (Z1), the piston (3) having a compression surface (28) at least partially demarcates a compression chamber (K1) formed within the cylinder (Z1), and a sealing gap (DS) is provided between the outer circumferential surface (27) of the piston body (20) and the cylinder sliding surface (10) of the cylinder (Z1).

10. The cylinder according to claim 9, characterized in that a temperature control pipe (62) is arranged around the cylinder sliding surface (10) inside the cylinder (Z1), and a temperature-controlled temperature control medium is guided through it.

11. The cylinder according to claim 9 or 10, characterized in that, as materials, ceramic is provided for the cylinder sliding surface (10) and Invar for the piston body (20), or steel is provided for the cylinder sliding surface (10) and ceramic for the piston body (20), or ceramic is provided for the cylinder sliding surface (10) and steel for the piston body (20), or ceramic is provided for the cylinder sliding surface (10) and ceramic for the piston body (20).

12. A piston compressor having at least one cylinder (Z1) according to any one of claims 9 to 11, wherein the piston rod (4) is coupled to a piston transmission device (12) of the piston compressor (1) at its axial end opposite to the piston (3).

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