Contactlessly sealing piston with piston rod
Patent Information
- Authority / Receiving Office
- EP · EP
- Patent Type
- Applications
- Current Assignee / Owner
- HOERBIGER WIEN GMBH
- Filing Date
- 2024-06-14
- Publication Date
- 2026-04-29
AI Technical Summary
Current piston compressors face challenges in high-pressure applications due to the limitations of piston rings and gap seals, leading to short service life, contamination risks, and inefficiencies, particularly at pressures above 350 bar, which has prevented their industrial use for unlubricated high-pressure applications.
A contactless sealing piston design with a double radial floating bearing system, including a piston sleeve and piston rod, minimizes contact between the piston and cylinder by using a small sealing gap, which is maintained through axial and sleeve gaps, allowing for pressure equalization and compensation of transverse forces, enabling operation up to 1000 bar.
This design extends the service life of piston compressors, prevents contamination, and enhances sealing efficiency, making it possible to use piston compressors in high-pressure applications, such as hydrogen compression, by reducing friction, wear, and maintaining a consistent sealing gap despite thermal and pressure changes.
Smart Images
Figure EP2024066496_26122024_PF_FP_ABST
Abstract
Description
[0001] Contactless sealing piston with piston rod
[0002] The invention relates to a contactless sealing piston with piston rod for a piston compressor, as well as the use of such a piston in a cylinder of a piston compressor.
[0003] As is well known, in a piston compressor, a piston is moved back and forth within a cylinder chamber to compress a compression medium therein. To ensure sufficient and efficient compression of the compression medium in the cylinder chamber, a sealing element is required between the moving piston and the cylinder or cylinder surface. Common sealing elements are contact seals, such as piston rings, which are inserted into grooves on the piston and rest against the cylinder surface. The purpose of a piston ring is to seal the cylinder chamber and minimize the leakage of compression medium via the piston. Typically, a plurality of piston rings are provided on the piston.
[0004] Piston rings and the cylinder bore are subject to friction and wear due to contact, piston speed, and acting transverse forces, which, in particular, limits the service life of the piston rings. Piston rings are therefore maintenance parts that must be replaced regularly.
[0005] Today's piston compressors almost exclusively use piston rings made of polymer-based materials. These are primarily high-performance polymers whose tribological properties are modified by fillers to minimize friction and wear. This can extend the service life of piston rings.
[0006] With piston compressors, a distinction must also be made between oil-lubricated and oil-free applications. In oil-lubricated applications, a thin oil film forms a boundary layer between the pressurized piston rings and the cylinder bore. This significantly reduces friction and wear compared to unlubricated applications. Furthermore, the oil film also has a sealing effect. However, in many applications, oil-lubricated piston rings are not possible because the lubricating oil can also contaminate the compression medium, which is often undesirable and impermissible.
[0007] In unlubricated, oil-free applications, tribology (friction and wear) is determined solely by the properties of the compression medium, operating conditions such as average piston speed, acting transverse forces, etc., the material and coating of the cylinder bore, and the material of the piston rings (e.g., polymer material and its fillers). This results in the formation of a transfer film between the piston ring material and the cylinder bore.
[0008] For unlubricated high-pressure applications with a final compression pressure of >250 bar, even typical high-performance polymers reach their material limits, and friction and wear on the piston rings increase disproportionately. While suitable modifications to the piston ring design can reduce the local contact pressure between the piston ring and the cylinder bore, this only allows an increase in the possible final compression pressure to approximately 350 bar, even in unlubricated applications. Furthermore, piston rings are not suitable because they would wear too quickly, thus shortening the achievable service life of the piston rings.
[0009] To achieve industrial-grade service lives, high-pressure pistons (for final compression pressures <350 bar) for reciprocating compressors are equipped with a large number of piston rings. 10 to 20 piston rings can be provided. However, this results in very large overall lengths for the pistons and, consequently, the piston compressor, and can lead to reduced heat dissipation of the individual piston rings, which in turn limits the service life.
[0010] The individual piston rings are inserted into grooves in the piston. For high-pressure applications, additional metal support rings are inserted behind the piston rings (on the side facing away from the pressure) to support them against the cylinder pressure acting on the piston ring. These support rings are designed to prevent the pressure-induced extrusion of the piston ring into the gap between the cylinder bore and the piston diameter.
[0011] A piston with piston rings and support rings for high-pressure applications is known, for example, from WO 2010 / 084071 A1. This document also describes the use of a support ring on the piston to guide and center the piston in the cylinder.
[0012] The typical gap width between support rings and cylinder bores is in the range of a few 0.1 mm. Smaller gap widths are not used in industry, as tolerances and transverse forces transmitted to the piston via the crankshaft drive would otherwise pose the risk of metal-to-metal contact between the cylinder bore and the support rings. Metal-to-metal contact would result in significant wear and could also damage the cylinder bore.
[0013] An alternative, well-known sealing concept for the piston-cylinder surface seal is a non-contact seal, such as a labyrinth seal or gap seal. In this case, the sealing effect is achieved via a cylindrical sealing gap between the piston outer diameter and the cylinder surface, which, in the labyrinth seal design, creates an additional throttling effect through additional grooves. To achieve sufficient sealing effectiveness in high-pressure applications > 350 bar, this sealing gap must be in the range <50 μm. Such a very small sealing gap causes a number of difficulties.
[0014] Unavoidable irregularities in the piston and / or cylinder bore surface due to manufacturing reasons, manufacturing tolerances on the piston and cylinder bore surface, as well as transverse forces acting on the crankshaft drive, almost inevitably lead to metallic contact between the piston and the cylinder bore surface at sealing gap heights of < 50 pm. This leads to significant wear and damage to piston compressor components after even a short operating period.
[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. Even at small temperature differences, such as those common in piston compressors, this can lead to contact between the piston and cylinder because the effective gap height approaches zero, or to disproportionate leakage because the gap height increases.
[0016] Particularly in high-pressure applications with final compression pressures of >350 bar, pressure-induced deformation of the cylinder and / or piston can occur, which can lead to changes in the gap height in the 10 pm range. This can lead to similar negative effects, namely contact or an increase in the gap height.
[0017] Last but not least, particles in the micrometer range that are carried along with the compression medium can get into the sealing gap, which can lead to the piston sticking or damage the piston surface and / or the cylinder bore.
[0018] For these reasons, a gap seal for piston compressors is theoretically possible, but no industrial implementation of such a gap seal in a piston compressor is known to date.
[0019] DE 2 160 816 A shows a piston for a dry-running reciprocating compressor. Plastic piston rings are supported radially on retaining surfaces on the piston, so that they act as a virtually wear-free labyrinth seal with a narrow gap against the cylinder wall. Guide rings ensure the straight guidance of the piston. However, even slight inclination of the piston in the cylinder causes the piston rings to close further and the gap to open, thus impairing the sealing effect. WO 98 / 31936 A1 describes an unlubricated piston compressor with a metallic cylinder surface and a piston with a plastic skirt, which form a split ring seal. The piston rod bearing, which allows pivoting of the piston rod, enables parallel running of the piston in the cylinder.
[0020] The aforementioned problems associated with piston rings and gap seals have so far prevented the industrial use of piston compressors for unlubricated high-pressure applications, especially at final compression pressures of >350 bar. Other compressor designs, such as diaphragm compressors or ionic compressors, have previously been used for such high-pressure applications. The short service life of the diaphragm is a problem with diaphragm compressors. With ionic compressors, mixing of the ionic liquid with the compression medium can occur, which is fundamentally undesirable. Apart from this, service life is also a problem with ionic compressors.
[0021] The use of a piston compressor even in high-pressure applications would be advantageous because the piston compressor is a well-known, proven and, above all, simple type of compressor.
[0022] It is therefore an object of the present invention to provide a piston of a piston compressor which enables use in a piston compressor for high-pressure applications.
[0023] This is made possible with a contactless sealing piston according to claim 1. With a piston designed according to the invention, a contactless gap seal can be realized because of the design, in particular the double radial floating bearing with a first floating bearing of the piston sleeve and a second floating bearing of the piston on the piston rod, which reduces the forces required to guide the piston to a level that means the gas dynamic forces occurring between the piston and cylinder are sufficient to minimize contact between the piston and cylinder. This means that a very small sealing gap can be selected and operated. At the same time, the axial gap and the sleeve gap create pressure equalization in the piston sleeve, so that the piston sleeve is not deformed or is only slightly deformed due to the acting pressure and the sealing gap is therefore not impaired either.Thus, the piston with piston rod according to the invention is particularly suitable for high-pressure applications with final compression pressures of up to 1000 bar and even higher. Such final compression pressures occur particularly in hydrogen applications. This makes it possible for the first time to use a piston compressor for such final compression pressures. In an alternative embodiment, several sleeve grooves can be provided on the piston body, axially spaced in the direction of the piston axis, with a piston sleeve arranged in each sleeve groove in a radially floating manner.
[0024] If at least one throttle groove is provided in an outer peripheral surface of the piston body, preferably several axially spaced throttle grooves, any particles located between the outer peripheral surface of the piston body and the cylinder bore can be trapped. This can prevent the formation of scratches on the cylinder bore and / or on the outer peripheral surface of the piston body. Furthermore, such a throttle groove can also enhance the sealing effect.
[0025] If the axial end face of the piston rod and / or the base surface of the piston rod recess is convexly curved, a possible angular misalignment between the piston and cylinder can be compensated and the high contact forces caused thereby can be prevented.
[0026] If the piston rod is elastically connected to the piston body in the piston recess via a radial or axial elastic connecting element, lifting of the piston body from the piston rod during start-up of the piston compressor can be prevented without compromising the floating bearing of the piston body. The same can be achieved with a preload unit that actively presses the piston body against the piston rod.
[0027] The present invention will be explained in more detail below with reference to Figures 1 to 9, which show exemplary, schematic and non-limiting advantageous embodiments of the invention.
[0028] Fig.1 a piston compressor,
[0029] Fig.2 and 3 each show a contactless sealing piston according to the invention, Fig.3 a detail of a contactless sealing piston according to the invention, Fig.4 to Fig.8 each show a radially movable bearing of the piston on the piston rod and
[0030] Fig.9 a cylinder of a piston compressor with a piston according to the invention.
[0031] Although the operating principle of a piston compressor 1 is generally well known, it will be briefly explained below with reference to Fig. 1. Fig. 1 shows a sectional view of a piston compressor 1 in an advantageous embodiment. The piston compressor 1 has a compressor housing 2 in which at least one cylinder Z1 is provided, in which a piston 3 can move reciprocally. The piston 3 is moved back and forth in the cylinder Z1 between a top dead center TDC and a bottom dead center BDC. The piston 3 partially delimits a cylinder chamber K1 in the cylinder Z1, in which a compression medium, for example hydrogen, is compressed by the movement of the piston 3. The piston 3 is connected 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 connected to a piston drive 12, which generates the reciprocating movement of the piston rod 4 and the piston 3.
[0032] 1, the piston drive 12 is designed as a crank drive with a crosshead 5, although the piston drive 12 can also be designed differently, for example with a cam drive. In this design, the piston rod 4 is connected at the end opposite the piston 3 to a crosshead 5 which is mounted axially movably in the compressor housing 2. The crosshead 5 is driven in order to move the piston rod 4 and the piston 3 back and forth. In the embodiment shown, the crosshead 5 is driven by a crank drive. For this purpose, the crosshead 5 is articulated to one end of a push rod 6 and the other end of the push rod 6 is articulated to a crank pin 7a of a crankshaft 7. The crankshaft 7 rotates during operation of the piston compressor 1. The crankshaft 7 can be driven by a drive device not shown in FIG. 1, e.g. by an electric machine.During operation of the compressor 1, the push rod 6 converts a rotary movement of the crankshaft 7 into a translational movement of the crosshead 5. The resulting transverse forces are supported on the compressor housing 2 via the crosshead 5. As a result, the piston rod 4 undergoes an essentially purely translational, reciprocating movement in the direction of the piston axis KA. The piston 3 can, of course, also be driven by means other than a crank drive.
[0033] The reciprocating piston rod 4 is typically guided through a sealing unit 9, which in turn is arranged in the compressor housing 2. The sealing unit 9 seals against the piston drive housing 8, such as the crankcase in Fig. 1, in which the crosshead 5 also moves, in order to prevent compression medium from entering the piston drive housing 8. The sealing unit 9 is preferably designed as a well-known sealing packing with a plurality of sealing rings and, if necessary, additional rings, such as scraper rings, support rings, etc.
[0034] The design of the piston compressor 1 with a reciprocating piston rod 4 also makes it possible to design the piston compressor 1 as a double-acting one if required, as in the exemplary embodiment according to Fig. 1 as a double-acting piston compressor 1. In a double-acting piston compressor 1, a second cylinder Z2 with a second cylinder chamber K2 is formed on a side of the piston 3 opposite the at least one first cylinder chamber K1, wherein the piston 3 also partially delimits the second cylinder chamber K2. During the reciprocating movement of the piston 3, the cylinder volume of the first cylinder Z1 is reduced (compression stroke) and simultaneously the cylinder volume of the second cylinder Z2 is increased (suction stroke), or vice versa.
[0035] The illustrated design as a double-acting compressor is, of course, only an example. In the simplest case, the piston compressor 1 has only a single cylinder Z1, which is defined by the end face of the piston 3 facing away from the piston drive 12.
[0036] Valves SV1, DV1, SV2, DV2 are also arranged in a known manner on a cylinder Z1, Z2 of the piston compressor 1, with at least one suction valve SV1, SV2 and at least one pressure valve DV1, DV2 being provided on each cylinder Z1, Z2. The valves can be designed as automatic valves, i.e., controlled by the acting cylinder pressure. However, the valves can also be actively controlled, for example, by means of known lifting grippers.
[0037] In the embodiment according to Fig.1, a double-acting piston compressor 1 is shown, in which both suction valves SV1, SV2 are connected to a suction line SL, via which compression medium KM is sucked in, and the two pressure valves DV1, DV2 are connected to a pressure line DL, via which compressed compression medium KM is discharged.
[0038] If the piston area on one side of the piston 3 is smaller than the piston area on the opposite side of the piston 3, a multi-stage piston compressor 1 can also be realized. In this case, the suction line of the first compressor stage would only be connected to the suction valve SV1 of the first compressor stage. The pressure line DL1 of the first compressor stage, connected to the pressure valve DV1 of the first compressor stage, would simultaneously be the suction line SL of the second compressor stage and thus connected to the suction valve SV2 of the second compressor stage. The pressure valve DV2 of the second compressor stage would be connected to the pressure line DL2 of the second compressor stage.
[0039] In the embodiment according to Fig.1, a piston ring 15 is arranged on the piston 3, which rests on the cylinder surface 10 of the cylinder Z1, Z2 in order to seal the respective cylinder chamber K1, k2.
[0040] Fig. 2 shows an embodiment of a contactless sealing piston 3 according to the invention with part of a piston rod 4 of a piston compressor 1 for high-pressure applications in use in a cylinder Z1 of a piston compressor 1. A contactless gap seal is provided to seal between the piston 3 and the cylinder Z1. As is usual with piston compressors 1, the piston 3 is moved back and forth in the cylinder chamber K1 in the cylinder Z1. The piston drive 12 of the piston 3 is not shown in Fig. 2 apart from the reciprocating piston rod 4 because this is irrelevant for the invention. Distances and gaps are greatly exaggerated in Fig. 2 for the purpose of illustration. In the case of a gap seal, the piston 3 is only single-acting, so that the piston 3 only delimits one cylinder chamber K1.
[0041] The piston 3 according to the invention consists of a piston body 20 and at least one hollow-cylindrical piston sleeve 21. At least one sleeve groove 24 in the form of a circumferential groove is provided on the piston body 20, in which the at least one piston sleeve 21 is arranged in a floating manner in the transverse direction transverse to the piston axis KA. "Floatingly mounted" here means that the piston sleeve 21 can move freely (apart from any possible frictional forces) in the sleeve groove 24 relative to the piston body 20 in the radial direction transverse to the piston axis KA.
[0042] The floating mounting of the piston sleeve 21 on the piston body 20 is achieved in that a radial sleeve gap HS is provided 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 in that the axial sleeve length HL of the at least one piston sleeve 21 in the direction of the piston axis KA is smaller than the axial groove length NL of the at least one sleeve groove 24.
[0043] When the piston 3 is used in a 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 surface 10 of the cylinder Z1, in which the piston 3 is moved during use, which creates the gap seal. The largest outer diameter of the piston body 20 in the region of the cylinder surface 10 is thus smaller than the inner diameter of the cylinder surface 10 and also smaller than the outer diameter of the outer peripheral surface 27 of the at least one piston sleeve 21.
[0044] The free movement in the radial direction of the piston sleeve 21 on the piston body 20 naturally only occurs within the design-defined range of motion, which is essentially defined on the piston 3 by the sleeve gap HS. However, when the piston 3 is used in a cylinder Z1 of a piston compressor 1, the possible mobility is naturally determined by the sealing gap DS. This mobility is also ensured when the piston 3 is used in a piston compressor 1.
[0045] In order to manufacture the piston 3, the piston body 20 can be constructed in several parts and consist of several piston body parts that are assembled to form the piston body 20. In the embodiment according to Fig. 2, the piston body 20 consists of a piston center part 25 and a piston end part 26, which is fastened to the piston center part 25, for example, screwed onto or into it. When the piston 3 is used in a piston compressor 1, the piston end part 26 faces the cylinder chamber K1. The outer diameter of the piston center part 25 is smaller than the outer diameter of the piston end part 26 in the direction of the piston axis KA along at least one specific axial groove length NL. The at least one region of the piston center part 25 with the smaller outer diameter thus forms the at least one sleeve groove 24 with groove length NL, in which the piston sleeve 21 is arranged in a floating manner.
[0046] To achieve the floating bearing 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, so that a sleeve gap HS is formed in the sleeve groove 24 between the inner circumferential surface 22 of the piston sleeve 21 and the outer circumferential surface 23 of the sleeve groove 24. The sleeve gap HS is preferably between 0.1 mm and 1 mm wide.
[0047] In addition, the axial sleeve length HL of the at least one piston sleeve 21 in the direction of the piston axis KA is smaller than the groove length NL of the sleeve groove 24 in which the piston sleeve 21 is arranged in a floating manner, so that an axial gap AS results between the piston sleeve 21 and the piston body 20. The axial gap AS is important for the floating bearing of the piston sleeve 21 in order to avoid axial clamping of the piston sleeve 21. The size of the axial gap AS is selected appropriately but should not be too large, since the axial gap AS creates undesirable dead space.
[0048] In the direction of the piston axis KA, the piston 3 is pressed against the end of the sleeve groove 24 opposite the cylinder Z1 due to the acting cylinder pressure pz, which, due to the axial gap AS, also acts on the end face of the piston sleeve 21 facing the cylinder Z1. However, the end face of the piston sleeve 21 is small compared to the cylinder cross-section, which is why the acting axial forces are low. This also reduces occurring frictional forces, which should not, or at least not significantly, restrict the mobility of the piston sleeve 21 in the radial direction transverse to the piston axis KA. In any case, the floating bearing of the piston sleeve 21 in the transverse direction remains unaffected in its function. The contact area of the piston sleeve 21 on the piston body 20 should preferably be as large as possible in order to keep the contact pressure between these two surfaces and the associated deformations as low as possible.Suitable measures for pressure equalization between these two surfaces can lead to a further reduction in the contact forces. The sleeve gap HS and the axial gap AS between the piston sleeve 21 and the piston body 20 thus ensure the floating bearing of the piston sleeve 21 in the radial direction on the piston body 20. This means that small lateral deviations HS (at most in the area of the sleeve gap) can be compensated for by the piston sleeve 21 during operation of the piston 3 in a cylinder Z1 of a piston compressor 1, without significant transverse forces occurring between the piston sleeve 21 and the cylinder running surface 10. This reduces potential wear in the event of unavoidable contact between the piston sleeve 21 and the cylinder running surface 10. However, the sleeve gap HS is small enough at 0.1 mm to 1 mm not to generate any significant additional dead space (space in the cylinder Z1 that cannot be used for compression).It is known that an excessively large clearance space reduces the delivery capacity of the piston compressor 1 and is therefore undesirable.
[0049] The sleeve gap HS and the axial gap AS also ensure that, during operation of the piston 3 in the piston compressor 1, the cylinder pressure pz from the cylinder chamber K1 passes into the sleeve gap HS between the piston sleeve 21 and the piston body 20 in order to enable pressure equalization between the outer peripheral surface 27 of the piston sleeve 21 and the inner peripheral surface 22 of the piston sleeve 21. This pressure equalization can reduce the deformation of the piston 3 due to the acting pressure and associated changes in the sealing gap DS between the outer peripheral surface 27 of the piston sleeve 21 and the cylinder bore 10.
[0050] The acting cylinder pressure pz also ensures that the piston sleeve 21 is centered in the cylinder Z1. Due to hydrodynamic effects, a pressure buildup occurs on one side where the sealing gap DS between the piston sleeve 21 and the cylinder bore becomes smaller, which in turn ensures that the piston sleeve 21 is moved back to a central position.
[0051] To achieve an effective gap seal, the sealing gap DS between the outer peripheral surface 27 of the piston sleeve 21 and the cylinder bore 10 is between 1 μm and 50 μm. The outer peripheral surface 27 of the piston sleeve 21 advantageously assumes the sealing function along the entire sleeve length HL due to the formed sealing gap DS.
[0052] Fig. 3 shows a design of the piston 3 with multiple piston sleeves 21. Instead of a single piston sleeve 21 mounted radially floating in a sleeve groove 24, as in the design according to Fig. 2, several piston sleeves 21 are used here, each mounted radially floating in a sleeve groove 24. The actual sealing surface for the contactless piston seal, the outer peripheral surface of the piston sleeve 21, is thus divided among multiple sealing surfaces.
[0053] In this embodiment, a sleeve groove 24 is provided for each piston sleeve 21, in which a respective piston sleeve 21 is arranged in a radially floating manner, in that a sleeve gap HS (which could also be different) is provided between the inner circumferential surface 22 of each piston sleeve 21 and the outer circumferential surface 23 of each associated sleeve groove 24, and the axial sleeve length HL of each piston sleeve 21 in the direction of the piston axis KA is smaller than the axial groove length NL of the associated sleeve groove 24. The sleeve lengths HL of the individual piston sleeves 21 do not have to be the same. The individual sleeve grooves 24 are axially spaced from one another in the direction of the piston axis KA. Two adjacent sleeve grooves 24 are each axially separated from one another by a radial web 31 on the piston body 20. The above explanation for Fig. 2 applies analogously to each piston sleeve 21 and each sleeve groove 24.Thus, there is a sleeve gap HS, an axial gap AS, a groove length NL and a sleeve length HL on the piston body 20. In the embodiment according to Fig.3, the piston body 20 is preferably designed in several parts with several piston body segments that are put together.
[0054] Due to the piston drive 12, transverse forces could be applied to the piston 3 via the connection of the piston rod 4 to the piston 3. Due to the small sealing gap DS, such transverse forces will cause contact forces between the piston 3, specifically the piston sleeve 21, and the cylinder bore 10. Such transverse forces, for example, due to axial misalignment between the cylinder 10 and the piston drive 12, can be compensated by the radial mobility of the piston sleeve 21.
[0055] The axial sleeve length HL of the piston sleeve 21 is related to the outer diameter D of the piston sleeve 21 in a preferred size ratio of HL / D = 1.5 to 3. In the case of several piston sleeves 21, this size ratio refers to the summed sleeve length HL of all piston sleeves 21.
[0056] The radial thickness d of a 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 transverse forces works, whereby the lower limit of the thickness d is determined by the required strength of the piston sleeve 21.
[0057] A possible angular misalignment between the axes of piston 3 and piston rod 4 would lead to high contact forces between piston 3 and cylinder 10 due to the resulting inclined position of piston 3. For this reason, the connection between piston 3 and piston rod 4 is additionally designed with a radially floating bearing. In this case, too, “floating bearing” means that piston 3 and piston rod 4 can move freely radially relative to one another (apart from any frictional forces). Free radial movement naturally only occurs within the design-defined possible range of movement. Piston rod 4 should be able to move freely radially relative to piston 3 without radially contacting piston 3. The radial movement play of piston rod 4 during operation naturally depends on the design of piston drive 12, but can be assumed to be known due to the known design of piston compressor 1. In Fig.4 is a version of the piston mounted radially movable on the piston rod 4.
[0058] 3. At the axial end of the piston 3 facing the piston rod 4, a piston rod recess 40 with a receiving depth AT is provided. The axial end of the piston rod 4 facing the piston 3 is arranged in the piston rod recess 40 such that the axial end face 41 of the piston rod 4 facing the piston rests against the base surface 42 of the piston rod recess 40. The piston rod 4 is thus arranged in the piston rod recess 40 over an axial length corresponding to the receiving depth AT. Radial mobility of the bearing is achieved by providing a piston rod gap KS in the radial direction between the piston rod 4 and the piston rod recess 40. The piston rod gap KS can be achieved by making the smallest inner diameter of the piston rod recess 40 larger than the largest outer diameter of the axial end of the piston rod 4 arranged in the piston rod recess 40.The piston rod gap KS is defined by the smallest radial distance between piston rod 4 and piston rod recess 40, which results from a concentric arrangement of piston rod 4 and piston rod recess 40.
[0059] The piston rod gap KS is preferably selected in the range between 0.1 mm and 10 mm.
[0060] If the piston rod 4 is designed in several parts, then the “piston rod” which is arranged in the piston rod recess 40 comprises at least the part of the piston rod 4 which forms the axial end facing the piston 3, but can also comprise further parts of the piston rod 4 or all parts of the piston rod 4.
[0061] The axial end face 41 of the piston rod 4 and / or the axial base surface 42 of the piston rod recess 40 can also be convexly curved (outwardly). This allows, on the one hand, the contact area between the piston rod 4 and the piston 3 to be reduced; on the other hand, such a convex curvature also allows a slight pivoting movement between the piston 3 and the piston rod 4. Thus, the piston 3 remains concentric in the cylinder Z1 even if the piston rod 4 pivots slightly relative to the piston 3. This enables, in particular, compensation for the previously mentioned possible angular offset, which cannot be compensated for by the floating bearing of the piston sleeve 21.
[0062] The floating arrangement of the piston 3 on the piston rod shown in Fig.4
[0063] 4 can also be provided in the embodiment according to Fig.2 or Fig.3.
[0064] Fig. 5 shows an alternative radially floating bearing of the piston 3 on the piston rod 4, whereby only the piston rod-side end of the piston 3 is shown. Fig. 6 shows an enlarged view of the bearing design of Fig. 5 for illustrative purposes. The piston 3 can be designed in particular as described for Fig. 2 or Fig. 3 or Fig. 4.
[0065] In the embodiment according to Fig. 5, in contrast to the embodiment according to Fig. 4, the piston rod 4 and the piston 3 are additionally radially elastically connected to one another in the piston rod recess 40 via a radial elastic connecting element 46. In the embodiment shown, an elastic ring, such as an O-ring, is arranged radially between the piston rod 4 and the piston rod recess 40 as a radial elastic connecting element 46. For this purpose, a ring receiving groove 47a, 47b is provided on the piston rod 4 and on the piston rod recess 40. When the piston rod 4 is arranged in the piston rod recess 40, the elastic ring sits in these two ring receiving grooves 47a, 47b and is held therein.However, the elasticity of the elastic ring and the dimensions of the ring receiving grooves 47a, 47b are selected such that the forces generated in the radial direction are so small that they can be neglected, and thus the floating bearing of the piston 3 in the radial direction is not impaired.
[0066] During operation of the piston compressor 1, the cylinder pressure pz acting in cylinder Z1 acts on the piston 3, which also ensures that the piston 3 moves with the piston rod 4 during the return movement from top dead center TDC. During operation of the piston 3, the radial elastic connecting element 46 therefore has no function and does not impair the function of the piston 3.
[0067] However, when starting up the piston compressor 1, for example after maintenance, it cannot be expected that the cylinder Z1 is already pressurized. If the piston drive 12 is driven in this state, the reciprocating movement of the piston 3 at top dead center TDC can lead to a loss of contact between the piston rod 4 and the piston 3. This would result in the piston 3 periodically striking the cylinder cover until the pressure built up by the compression movement is sufficient to compensate for this loss of contact. The radial elastic connecting element 46 also serves to prevent such a loss of contact when starting up the piston compressor 1. The radial elastic connecting element 46 represents an elastic connection between the piston rod 4 and the piston 3, by means of which the piston 3 can be easily retracted with the piston rod 4.
[0068] In Fig.6 a convex end face 42 of the piston rod 4 is also shown.
[0069] The radially floating arrangement of the piston 3 on the piston rod 4 shown in Figs. 5 and 6 can also be provided in the embodiment according to Fig. 2 or Fig. 3. Instead of a radial elastic connecting element 46 between the piston rod 4 and the piston 3, an axial elastic connecting element 46 can also be provided between the piston rod 4 and the piston 3, as described with reference to Fig. 7. Such an axial elastic connection also allows the retraction of the piston 3 via the piston rod 4, but does not impair the radially floating arrangement of the piston 3 on the piston rod 4.
[0070] An axial elastic connecting element 46 in the form of an elastic body is provided in the axial end of the piston rod 4 facing the piston 3. The arrangement of the piston rod 4 in the piston rod recess 40 with the piston rod gap KS is not changed by this. Likewise, the base surface 42 of the piston rod recess 40 and the end surface 41 of the piston rod 4 abut one another. At least one of these surfaces can also be convexly curved, as described above. The axial elastic connecting element 46 is merely provided additionally. For this purpose, a connecting element recess 48 can be provided at the axial end of the piston rod 4, in which the axial elastic connecting element 46 is arranged.A threaded pin 49 protrudes from each of the axial elastic connecting elements 46 at each of the two axial ends, one of the threaded pins 49 being screwed into a threaded bore 50 on the piston body 20 and the opposite threaded pin 49 being screwed into a threaded bore 50 in the piston rod 4.
[0071] The radially floating arrangement of the piston 3 on the piston rod 4 shown in Fig.7 can also be provided in the embodiment according to Fig.2 or Fig.3.
[0072] Fig. 8 describes a further embodiment of a radially floating arrangement of the piston 3 on the piston rod 4. In this embodiment, too, the axial end of the piston rod 4 is arranged with a piston rod gap KS in the transverse direction in a piston rod recess 40 on the piston body 20. Likewise, the base surface 42 of the piston rod recess 40 and the end surface 41 of the piston rod 4 axially abut one another. At least one of these surfaces can also be convexly curved, as described above.
[0073] In the embodiment according to Fig.8, the piston rod 4 is additionally actively pressed axially against the base surface 42 of the piston rod recess 40 by a preloading unit 51, without impairing the floating bearing, in particular without impairing the piston rod gap KS.
[0074] In the embodiment shown, the preload unit 51 is designed with a clamping sleeve 52. The clamping sleeve 52 is fastened to the axial end of the piston body 20, for example by means of a screw connection as in Fig. 8, or glued, pressed in, or secured in some other way. A radially inwardly projecting shoulder 53 is provided on the axial end of the clamping 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 rod recess 40. The radial web 54 can be formed by a ring fastened to the piston rod 4, but can also be formed integrally with the piston rod 4. The radial web 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 web 54 and the shoulder 53, which preloads the radial web 54 and thus the piston rod 4 in the direction of the piston 3.This ensures a certain axial contact force between piston rod 4 and piston 3, which is particularly advantageous in the case of force reversal, since it prevents a loss of contact between piston rod 4 and piston 3 during the pressureless start-up phase of the piston compressor 1.
[0075] The radially floating arrangement of the piston 3 on the piston rod 4 shown in Fig.8 can also be provided in the embodiment according to Fig.2 or Fig.3.
[0076] Due to the piston 3 being mounted radially floating on the piston rod 4 and the piston sleeve 21 being mounted floating on the piston 3, the piston body 20 could shift in the radial direction, which could lead to contact between the piston body 20 and the cylinder bore 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, for example, in Figures 2, 3, 4 and 5. The support ring 28 is arranged in a support ring groove 29 in the outer peripheral surface of the piston body 20, wherein the support ring groove 29 is, of course, axially spaced in the direction of the piston axis KA from the sleeve groove 24 for the at least one piston sleeve 21 closest to the piston rod-side end of the piston 3. The support ring 28 protrudes radially from the outer peripheral surface of the piston body 20.
[0077] The support ring 28 is intended to support any remaining transverse forces against the cylinder bore 10 during operation of the piston 3 and, above all, to ensure the central running of the piston body 20 along the cylinder axis (which coincides with the piston axis KA). The support ring 28 thus rests against the cylinder bore 10 during operation of the piston 3 and is therefore advantageously made of a tribologically favorable and wear-resistant material, in particular a polymer material. Since the support ring 28 is arranged on the piston rod end, and thus on the side facing away from the pressure during operation of the piston 3, and only has to absorb small transverse forces, wear on the support ring 28 is low. In addition, the support ring 28 can be pressure-balanced, so that no forces occur on the support ring 28 during operation of the piston 3 due to possible pressure differences acting on the support ring 28. This also contributes to keeping wear on the support ring 28 to a minimum.The outer diameter of the support ring 28 can correspond to a maximum of the inner diameter of the cylinder bore 10 and can be smaller or larger than the outer diameter of the at least one piston sleeve 21.
[0078] In order to prevent the piston body 20 from coming into contact with the cylinder bore 10 during operation of the piston 3, the largest outer diameter of the piston body 20 (at least of the part of the piston body 20 which runs in the cylinder bore 10 during operation) is in any case smaller than the outer diameter of the piston sleeve 21.
[0079] With reference to Fig. 4, further possible features of a piston 3 according to the invention are described. These additional features are independent of the design of the floating arrangement of the piston 3 on the piston rod 4.
[0080] In the design of the piston 3 according to Fig. 4, throttle grooves 30 in the form of circumferential grooves are provided on the outer circumferential surface 27. Such throttle grooves 30 are optional, and only one throttle groove 30 can also be provided. If present, the at least one throttle groove 30 serves to intercept and collect any particles located between the outer circumferential surface 27 of the piston sleeve 21 and the cylinder bore 10. This can prevent the formation of scratches on the cylinder bore 10 and / or on the outer circumferential surface 27 of the piston sleeve 21 or the lifting of a chip by such particles. In addition, the throttle groove 30 can be designed such that a throttling effect occurs, similar to a contactless labyrinth seal, which increases the contactless sealing performance on the piston sleeve 21 during operation of the piston 3.In the circumferential groove 30, turbulence occurs in the compression medium flowing along the outer circumferential surface 27 of the piston sleeve 21, leading to losses that, in turn, reduce the acting pressure. This throttling effect is particularly effective when the ratio between the sealing gap DS and the radial groove depth of the circumferential groove is at least 1.5, preferably at least 2.
[0081] In an embodiment with several piston sleeves 21 on the piston body as shown in Fig.3, a throttle groove 30 is formed automatically due to the axial separation of the sleeve grooves 24 between two adjacent piston sleeves 21, which has the same effect as a throttle groove in Fig.2. Nevertheless, in an embodiment according to Fig.3, at least one throttle groove 30 could additionally be provided in the outer circumferential surface 27 on at least one piston sleeve.
[0082] In the embodiment according to Fig. 4, a split ring 43, preferably made of a polymer such as PTFE, is provided in the axial gap AS between the piston body 20 and the axial end face of the piston sleeve 21 facing away from the connecting rod 4. The split ring 43 is designed such that the piston sleeve 21 is not axially clamped in the sleeve groove 24 or is mounted with a slight preload, which is < 10% of the surface pressure imposed by the final compression pressure, so that the floating bearing of the piston sleeve 21 in the sleeve groove 24 is not lost. This can be achieved by designing the split ring 43 elastically, whereby the effective elastic forces in the direction of the piston axis KA are so small that they can be neglected. In the case of a rigid design of the split ring 43, the axial length of the split ring 43 may correspond to a maximum of the axial length of the axial gap AS, preferably less than the axial length of the axial gap AS.Secondly, the split ring 43 is intended to ensure a connection between the sleeve gap HS and the outer peripheral surface of the piston 3 or the outer peripheral surface 27 of the piston sleeve 21, so that the cylinder pressure pz can reach the sleeve gap HS. For this purpose, the split ring 43 can be designed to create such a connection, or this connection can be achieved by having the axial length of the split ring 43 smaller than the axial length of the axial gap AS.
[0083] Such a split ring 43 can also be provided in an embodiment according to Fig. 2. In an embodiment according to Fig. 3, a split ring 43 can be provided at least in the axial gap AS between a piston sleeve 21 and the piston body 20, preferably in all piston sleeves 21.
[0084] In the embodiment according to Fig. 4, an elastic sealing element 44 is also provided between the piston rod-side end face 45 of the piston sleeve 21 and the piston rod-side axial end of the sleeve groove 24. In the embodiment according to Fig. 4, the sealing element 44 is arranged in a groove at the axial end of the sleeve groove 24, but could also simply be inserted between the piston sleeve 21 and the axial end of the sleeve groove 24.
[0085] The sealing element 44 improves the sealing of the sleeve gap HS, in which the cylinder pressure pz is applied. The piston sleeve 21 is pressed by the cylinder pressure pz against the piston rod-side axial end of the sleeve groove 24, as already explained above. This also seals the sleeve gap HS, so that no leakage path for the compression medium can develop via the sleeve gap HS. The sealing element 44 can improve this seal or make it more reliable. In particular, the sealing element 44 can also compensate for possible inclinations of the piston sleeve 21 relative to the piston axis KA, and in the event of such an inclination, a leakage path can be prevented from opening at the piston rod-side axial end of the piston sleeve 21.
[0086] Alternatively, an elastic sealing element 44 could also be arranged between the cylinder-side axial end of the sleeve groove 24 and the piston sleeve 21. In this embodiment, the split ring 43 would be omitted or replaced by the elastic sealing element 44. The elastic sealing element 44 at the opposite axial end (as shown in Fig. 4) could also be omitted in this embodiment or deliberately designed to be gas-permeable. This embodiment would have the effect that the pressure applied to the inner diameter of the piston sleeve 21 is lower than the pressure applied to the outer diameter of the piston sleeve 21 and thus the piston sleeve 21 would be slightly compressed by the resulting differential pressure. This would indeed increase the sealing gap DS during operation and thus also the leakage compared to a embodiment as shown in Fig.4, but would have the advantage that this design would ensure that in borderline cases no pressure build-up at the inner diameter of the piston sleeve 21 could occur that was higher than at the outer diameter of the piston sleeve 21, which could lead to a radial deformation of the piston sleeve 21 to the outside and possible contact of the piston sleeve 21 with the cylinder bore 10.
[0087] A sealing element 44 as described above can also be provided in an embodiment according to Fig. 2. In an embodiment according to Fig. 3, a sealing element 44 can be provided in at least one piston sleeve 21, preferably in all piston sleeves 21. The sealing element 44 can be provided in addition to a split ring 43, as in Fig. 4, but can also be provided without a split ring 43.
[0088] The materials of the cylinder bore 10 and the piston sleeve 21 are selected so that the thermal deformations caused by different thermal expansion coefficients, together with the respective deformation due to the applied pressure, enable the maintenance of a functional sealing gap DS. This means that the sealing gap DS cannot become too large or too small due to thermal expansion and load-induced deformation. It is advantageous if the thermal expansion coefficients of the materials used differ by a maximum of 1 x 10' 61 / K. Preferably, but not necessarily, the thermal expansion coefficient of the material of the cylinder bore 10 is greater than the thermal expansion coefficient of the material of the piston sleeve 21. This can prevent the piston sleeve 21 from expanding more than the cylinder bore 10, thus reliably preventing the thermally induced closing of the sealing gap DS between the piston sleeve 21 and the cylinder bore 10. Possible material combinations to ensure this are ceramic for the cylinder bore 10 and Invar (iron-nickel alloy with a very low thermal expansion coefficient) for the piston sleeve 21 or steel for the cylinder bore 10 and ceramic for the piston sleeve 21. Likewise, with a suitable design, a piston sleeve 21 made of steel can be combined with a ceramic cylinder bore 10. A design of both contact partners made of ceramic materials is also possible. Fig.Figure 9 shows a cylinder Z1 of a piston compressor 1, in which a piston 3 according to the invention with a piston rod 4 is moved back and forth. In this embodiment, the cylinder chamber K1 in the cylinder Z1 is delimited by a cylinder head 60, in which the suction valve SV1 and the pressure valve DV1 are also arranged. The piston 3 with piston rod 4 can be designed as described above with reference to Figures 2 to 8. In this embodiment, the cylinder running surface 10 is formed on a cylinder liner 61, which is arranged in the cylinder Z1, which corresponds to the usual design.
[0089] In the embodiment shown in Fig. 9, cylinder Z1 is cooled. In the cylinder housing of cylinder Z1 and around cylinder surface 10, or rather around cylinder liner 41, temperature control lines 62 are provided, through which a temperature-controlled temperature control medium is passed. By means of the temperature control lines 62 and the temperature control medium, cylinder surface 10 can be temperature-controlled to a specific temperature, whereby the thermal expansion of cylinder surface 10 and thus also the sealing gap DS between cylinder surface 10 and the at least one piston sleeve 21 can be controlled. For this purpose, a temperature sensor 63 can also be provided in cylinder Z1, which detects the temperature in the area of cylinder surface 10. Temperature sensor 63 can be integrated into a closed control loop in order to regulate the temperature in the area of cylinder surface 10 by adjusting the temperature of the temperature control medium.The temperature control medium can be both cooled and heated in a controlled manner. Cooling will be necessary, in particular, during operation of the piston compressor 1 when the cylinder Z1 heats up due to the compression of the compression medium. When starting the piston compressor 1 or when operating in low ambient temperatures, it may also be necessary to heat the temperature control medium. The temperature of the temperature control medium can also or additionally be controlled depending on the current sealing gap DS in order to keep the sealing gap DS within a desired range via the thermal expansion of the cylinder bore 10 and the piston sleeve 21. For this purpose, the sealing gap DS can also be measured, for example using an ultrasonic measurement.
Claims
Patent claims 1 . Contactless sealing piston (3) with piston rod (4) for a piston compressor (1), characterized in that the piston (3) comprises a piston body (20) and at least one piston sleeve (21), wherein at least one sleeve groove (24) is provided on the piston body (20), in which the at least one piston sleeve (21) is arranged floatingly in the radial direction transverse to a piston axis (KA), in that a sleeve gap (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) in the direction of the piston axis (KA) is smaller than the axial groove length (NL) of the at least one sleeve groove (24) to form an axial gap (AS), in that a piston rod recess (40) is formed on a piston rod-side axial end of the piston body (20), in which 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 base surface (42) of the piston rod recess (40), and the piston body (20) is arranged on the piston rod (4) in a floating manner in the radial direction, in that a piston rod gap (KS) is provided between the piston rod recess (40) and the piston rod (4) in the radial direction, and in that at the piston rod-side axial end of the piston body (20), on an outer peripheral surface of the piston body (20), a support ring groove (29) is provided, in which a support ring (28) is arranged, which protrudes radially from the outer peripheral surface of the piston body (20).
2. Contactless sealing piston (3) with piston rod (4) according to claim 1, characterized in that a plurality of sleeve grooves (24) axially spaced apart in the direction of the piston axis (KA) are provided on the piston body (20), wherein in each sleeve groove (24) a piston sleeve (21) is arranged floatingly mounted in the radial direction.
3. Contactless sealing piston (3) with piston rod (4) according to claim 1 or 2, characterized in that at least one throttle groove (30) is provided in an outer peripheral surface (27) of at least one piston sleeve (21).
4. Contactless sealing piston (3) with piston rod (4) according to claim 1, characterized in that a plurality of throttle grooves (30) spaced axially apart from one another are provided in an outer peripheral surface (27) of the at least one piston sleeve.
5. Contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 4, characterized in that the support ring groove (29) is axially between a piston rod-side axial end of the piston body (20) and an axial end of the sleeve groove (24) closest to the piston rod-side axial end of the piston body (20).
6. Contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 5, characterized in that the axial end face (41) of the piston rod (4) and / or the base surface (42) of the piston rod recess (40) is convexly curved.
7. Contactless sealing piston (3) with piston rod (4) according to 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 recess (40) via a radial or axial elastic connecting element (46).
8. Contactless sealing piston (3) with piston rod (4) according to claim 7, characterized in that an elastic ring is provided as a radial elastic connecting element (46) between the piston rod (4) and the piston rod recess (40).
9. Contactless 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 recess (40) and on the piston rod (4) and the elastic ring is arranged in the ring receiving grooves (47a, 47b).
10. Contactless 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 recess (48) is provided on the axial end of the piston rod (4) facing the piston (3), in which connecting element recess the axial elastic connecting element (46) is arranged, wherein an axially projecting threaded pin (49) is provided on each of the two axial ends of the axial elastic connecting element (46), and wherein one threaded pin (49) is screwed into a threaded bore (50) on the piston body (20) and the opposite threaded pin (49) is screwed into a threaded bore (50) on the piston rod (4).
11. Contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 6, characterized in that at the piston rod side axial end of the piston body (20) a clamping sleeve (52) is fastened to the piston body (20), wherein at the axial end of the clamping sleeve (52) facing away from the piston (3) a shoulder (53) projects radially inwards, wherein on the piston rod (4) axially between the shoulder (53) and the piston rod recess (40) a radial web (54) is provided and wherein between the radial web (54) and the shoulder (53) a spring element (55) is arranged, which presses the piston rod (4) in the direction of the piston (3).
12. Contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 11, characterized in that an elastic sealing element (44) is arranged between a piston rod-side end face (45) of the piston sleeve (21) and a piston rod-side axial end of the sleeve groove (24).
13. Contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 12, characterized in that a split ring (43) is arranged in the axial gap (AS) between the piston body (20) and an axial end face of the piston sleeve (21) facing away from the piston rod (4), which split ring establishes a connection of the sleeve gap (HS) with an outer peripheral surface of the piston (3).
14. Cylinder for a piston compressor (1), wherein a contactless sealing piston (3) with piston rod (4) according to one of claims 1 to 13 is reciprocally arranged in the cylinder (Z1) and the piston (3) at least partially delimits a compression chamber (K1) formed in the cylinder (Z1), wherein a sealing gap (DS) is provided between the outer circumferential surface (27) of the at least one piston sleeve (21) or the outer circumferential surfaces (27) of the piston sleeves (21) and a cylinder running surface (10) of the cylinder (Z1).
15. Cylinder according to claim 14, characterized in that in the cylinder (Z1) around the cylinder running surface (10) a tempering line (62) is arranged, through which a tempered tempering medium is passed.
16. Cylinder according to claim 14 or 15, characterized in that the material provided is ceramic for the cylinder running surface (10) and Invar for the piston sleeve (21) or steel for the cylinder running surface (10) and ceramic for the piston sleeve (21) or ceramic for the cylinder running surface (10) and steel for the piston sleeve (21) or ceramic for the cylinder running surface (10) and ceramic for the piston sleeve (21).
17. Piston compressor with at least one cylinder (Z1) according to one of claims 14 to 16, characterized in that the piston rod (4) is connected to a piston drive (12) of the piston compressor (1) at the axial end facing away from the piston (3).