Turbine

Abstract

Turbine (20) with an impeller (23) arranged in a housing (26), wherein the turbine (20) has an inlet region (21) and an outlet region (22) and is supplied with a working medium during operation, the working medium flowing into the inlet region (21), along a front face (23a) formed on the impeller (23) and subsequently out of the outlet region (22), wherein a pressure gradient exists on the front face (23a) between the inlet region (21) and the outlet region (22), wherein a pressure divider (9) is arranged on the rear face (23b) of the impeller (23) opposite the front face (23a), wherein the pressure divider (9) comprises a sliding ring (31) which interacts with the rear face (23b) of the impeller (23) and thus forms a steam-lubricated throttle, wherein a first flow path (51) passes through the throttle, the throttle lubricating the rear face (23b) divides into a first area (231) and a second area (232),wherein the first region (231) limits the inflow region (21) and wherein the second region (232) limits a pressure chamber (11), wherein in operation the inflow region (21) is subjected to a higher pressure than the pressure chamber (11), wherein the sliding ring (31) is axially movable, wherein a sealing ring (33) arranged in a groove (41) interacts with the sliding ring (31), wherein a second flow path (52) runs between the groove (41) and the sliding ring (31) from the inflow region (21) to the pressure chamber (11), wherein the second flow path (52) can be closed off by the sealing ring (33), wherein the sealing ring (33) is movable in a defined manner in the groove (41), characterized in that the groove (41) is formed in a clamping ring (40), wherein the clamping ring (40) is fixedly connected to the housing (26).

Description

[0001] The invention relates to a turbine with an impeller, wherein a pressure divider is arranged on the rear side of the impeller. The turbine can be used, for example, within a waste heat recovery system for utilizing waste heat from an internal combustion engine. State of the art

[0002] Turbines with a casing and an impeller arranged within the casing are known from the prior art, for example from patent EP 1 281 836 B1. The known turbine has an inlet region and an outlet region and is supplied with a working fluid during operation. The working fluid flows into the inlet region, along a front surface formed on the impeller, and then out of the outlet region, with a pressure gradient on the front surface between the inlet region and the outlet region. Furthermore, the known turbine has at least two control valves, which are controlled by a governor and regulate the pressure on a rear surface of the impeller. This allows the resulting aerodynamic flow acting on the impeller to be controlled in such a way that the resulting axial force, also called axial thrust, is minimized. Accordingly, a axial bearing of the shaft is subjected to less stress.The axial bearing can be made smaller.

[0003] Furthermore, patent application DE 10 2014 226 951 A1 discloses a turbine with a pressure divider. The pressure divider is arranged on the rear side of the impeller. It comprises a sliding ring that interacts with the rear side, thus forming a steam-lubricated throttle. A first flow path runs through the throttle, which hydraulically divides the rear side into a first and a second section. The first section defines the inlet area, and the second section defines a pressure chamber. The two sections are designed, i.e., separated by the throttle, to be subjected to different pressures. During turbine operation, the inlet area is subjected to a higher pressure than the pressure chamber. The sliding ring is axially movable or floating. A sealing ring arranged in a groove interacts with the sliding ring.A second flow path runs between the groove and the sliding ring, leading from the inlet area to the pressure chamber. This second flow path can be closed off by the sealing ring.

[0004] EP 2 060 804 A1 discloses an axial sliding bearing arrangement with a pressure ring designed for joint rotation with a rotating component.

[0005] US patent 9 188 133 B1 discloses a product that prevents oil from entering a turbocharger compressor.

[0006] Due to the pressure divider, the axial force acting on the impeller can be kept almost constant. However, the axial movement of the sliding ring places very high stress on the sealing ring of the known turbine during operation, thus increasing its risk of wear. Furthermore, the sealing function is therefore not robust across all operating pressures. Disclosure of the invention

[0007] In contrast, the turbine according to the invention has the advantage that the stresses acting on the sealing ring are minimized, while at the same time the sealing function is ensured across all operating pressures of the turbine. This increases the reliability and service life of the turbine. In addition, the sealing ring acts as a damping optimization element for the axial dynamics of the turbine.

[0008] The turbine according to the invention comprises a housing in which an impeller is arranged. The turbine has an inlet region and an outlet region and is supplied with a working fluid during operation. The working fluid flows into the inlet region, along a front face formed on the impeller, and then out of the outlet region. A pressure gradient exists on the front face between the inlet region and the outlet region. A pressure divider is arranged on the rear face of the impeller opposite the front face. The pressure divider comprises a sliding ring that interacts with the rear face of the impeller, thus forming a steam-lubricated throttle. A first flow path runs through the throttle, which divides the rear face into a first region and a second region. The first region defines the inlet region, and the second region defines a pressure chamber.During operation, the inlet area is subjected to a higher pressure than the pressure chamber. The sliding ring is axially movable. A sealing ring arranged in a groove interacts with the sliding ring. A second flow path runs between the groove and the sliding ring, from the inlet area to the pressure chamber. This second flow path can be closed off by the sealing ring. The sealing ring has defined movement within the groove. The groove is formed within a clamping ring, which is firmly connected to the housing, for example, by being pressed into it. This allows for tighter dimensions and tolerances of the groove. The functionality of the groove for guiding the sealing ring can thus be made more robust. In an advantageous embodiment, the groove is formed within the housing. This saves on components, making the turbine more cost-effective.If the impeller moves due to fluctuating operating pressures, the sliding ring can follow the impeller's movement because the groove guides the sealing ring precisely, thus preventing the sealing ring from blocking its movement. At relatively low pressures in the inlet area, or low operating pressures, the axial movement of the sliding ring should cause the sealing ring to roll. For this to occur, the frictional force, and therefore the pressure on the sealing ring, must be high enough to prevent the sealing ring from slipping in the groove. Conversely, at high operating pressures, the axial movement of the sliding ring should cause the sealing ring to slip or slide. For this to happen, the frictional force, and therefore the pressure on the sealing ring, must be relatively low to allow this sliding motion.

[0009] The defined guidance of the sealing ring in the groove ensures slippage at high pressures and rolling at low pressures, as this defined guidance reduces the contact force between the sliding ring and the sealing ring with increasing inlet pressure. The groove is preferably concave for this purpose. However, the sealing function is always guaranteed, if required. Wear of the sealing ring is minimized because the maximum contact pressure is not applied to the sealing ring throughout its entire service life.

[0010] In advantageous embodiments, the pressure chamber is hydraulically connected to the discharge area. This means that at least part of the rear of the impeller, namely the second area, is subjected to the pressure level of the discharge area, i.e., the discharge pressure, or to a pressure level controlled by the discharge pressure. Since the pressure load on the front of the impeller also depends on the discharge pressure, the pressure loads on the front and rear of the impeller are similarly varied when the turbine is operated at different operating points. The resulting axial force acting on the impeller can thus be kept at a consistently low level. Preferably, however, the level of the axial force is set so that there is no change in the direction of the axial force at the different operating points, since the axial force then only has to be absorbed in one direction.

[0011] In advantageous embodiments, the groove is J-shaped. Preferably, the groove has a tapered section in the rounded area of ​​the J-shape. This allows the sealing ring to be guided very well within the J-shape. The distance between the center point of the sealing ring's cord thickness and the sliding ring can thus be varied, thereby controlling the contact pressure between the sealing ring and the sliding ring.

[0012] Advantageously, the tapered section has a ramp for inserting the sealing ring into the tapered area. This facilitates insertion and simultaneously reduces stress on the contact between the sealing ring and the sliding ring. Preferably, the ramp is rounded at its end to prevent damage to the sealing ring during insertion into the tapered section.

[0013] In advantageous embodiments, a linear ramp is formed on the ramp area to guide the sealing ring. The linear ramp is preferably inclined at 30° to 40° relative to the sliding ring. This is an optimized design to reduce the contact pressure between the sealing ring and the sliding ring with increasing inlet pressure. In advantageous embodiments, the sealing ring is lifted from the sliding ring to such an extent that the sealing function is eliminated and the second flow path is opened. The pressure of the inlet area can then be reduced; in this embodiment, the sealing ring also fulfills the function of a pressure relief valve.

[0014] In advantageous embodiments, the turbine is designed as a radial turbine. This allows the rear of the impeller to be subjected to the pressures set by the pressure divider using comparatively simple designs, while the main flow direction of the working medium runs on the front of the impeller.

[0015] In preferred embodiments, the turbine according to the invention, designed as an expansion machine, is arranged in a waste heat recovery system, particularly of an internal combustion engine. The waste heat recovery system comprises, in the direction of flow of the working medium, a feed fluid pump, an evaporator, the turbine, and a condenser. The waste heat recovery system is preferably not operated at a single operating point, but at a wide range of operating points, since the internal combustion engine is also operated at different operating points. Thus, different pressures occur in the turbine's inlet area, causing the impeller to perform axial movements. Consequently, the sliding ring must also be able to accommodate these axial movements to maintain the robust functionality of the pressure divider across all operating points.The design of the sealing ring according to the invention allows the sliding ring to dynamically follow the impeller movement with as little friction and wear as possible, without losing its sealing effect.

[0016] Therefore, it is very advantageous for the overall efficiency and service life of the waste heat recovery system if the turbine can also operate dynamically and stably under different operating conditions. The turbine according to the invention is thus particularly well suited for this purpose. Drawings Fig. Figure 1 schematically shows a turbine in longitudinal section, as known from the prior art, with only the essential areas shown. Fig. Figure 2 shows a section of the turbine in the area of ​​a sealing ring in cross-section. Fig. Figure 3 shows section III of the Fig. 2. Fig. Figure 4 shows a cross-section of an embodiment of the turbine according to the invention in the area of ​​the sealing ring. Fig. Figure 5 shows a detailed cross-sectional design of a groove for receiving the sealing ring. Description

[0017] Fig. Figure 1 schematically shows the longitudinal section of the installation of a turbine 20 as known from the prior art, with only the essential areas depicted. The turbine 20 is designed as a radial turbine, with a radial inflow direction 21a of the working medium in an inflow region 21 and with an axial outflow direction 22a of the working medium in an outflow region 22. The turbine 20 of this embodiment can also be operated as a pump by reversing the flow directions 21a, 22a.

[0018] The turbine 20 comprises a housing 26 and, within the housing 26, an impeller 23 arranged on a shaft 24, the inlet section 21, the outlet section 22, a pressure divider 9, a pressure chamber 11, and a shaft seal 25. The pressure divider 9 is designed as a sliding ring 31, which interacts with a rear surface 23b of the impeller 23.

[0019] The inlet area 21 has an inlet pressure p 21 on, the outflow area 22 an outflow pressure p 22 The working medium flows along a front face 23a of the impeller 23 from the inlet area 21 to the outlet area 22 and is thereby depressurized, so that a pressure gradient exists between the inlet pressure p 21 and the outflow pressure p 22 arises.

[0020] On the rear side 23b of the impeller 23, opposite the front side 23a, the pressure chamber 11 is formed between the housing 26 and the impeller 23, which is subject to the compensation pressure p 11The compensation pressure p acts on impeller 23. 11 the axial components of the inlet pressure p 21 and the outflow pressure p 22 on the front side 23a opposite.

[0021] The shaft seal 25, arranged on the housing 26, has two sealing lips 25a and 25b, so that the pressure chamber 11 can be effectively sealed against an ambient space 50 in both directions. For this purpose, the shaft seal 25 interacts with the shaft 24 via both sealing lips 25a and 25b, thereby sealing the pressure chamber 11 under compensating pressure p. 11 against the surrounding space 50 under ambient pressure p 50 away.

[0022] Between the housing 26 and the impeller 23 or the rear side 23b, the pressure divider 9 is designed as a throttle or as a steam-lubricated throttle, with a gap between the sliding ring 31 and the rear side 23b. A shoulder 23c is arranged on the rear side 23b of the impeller 23. A sliding surface 23d is arranged on the end face of the shoulder 23c, facing the housing 26. A receiving groove 26b is formed in the housing 26 opposite the sliding surface 23d. The sliding ring 31 is arranged at least partially within the receiving groove 26b. An end face 31a formed on the sliding ring 31 interacts with the sliding surface 23d, preferably in the axial direction of the impeller 23, to form the throttle or the steam gap. A sliding ring spring 32 arranged in the receiving groove 26b clamps the sliding ring 31 against the impeller 23. A sealing ring 33 is arranged radially between the sliding ring 31 and the housing 26, sealing the sliding ring 31 to the housing 26.

[0023] The pressure divider 9 hydraulically connects the pressure chamber 11 to the inlet area 21 via a first flow path that runs through the throttle. The pressure divider 9 is not located at the outermost circumference of the impeller 23, but rather between the outermost circumference and the diameter of the shaft 24, so that a first area 231 with the inlet pressure p is located on the rear side 23b. 21 of the inflow area 21 and a second area 232 with the lower pressure of the pressure chamber 11, namely with the compensation pressure p 11 The sealing ring 33 blocks a second flow path from the inlet area 21 to the pressure chamber 11.

[0024] The shaft 24 has a blind bore 24a extending axially. Furthermore, the shaft 24 has a connecting bore 24b radially to the blind bore 24a and a groove 24c on its surface. The impeller 23 has an outlet channel 13 which hydraulically connects the groove 24c to the pressure chamber 11.

[0025] The blind bore 24a also opens into the outlet area 22. This creates a hydraulic connection in the flow direction of the working medium from the pressure chamber 11 via the outlet channel 13, the groove 24c, the connecting bore 24b and the blind bore 24a to the outlet area 22. In the exemplary embodiment of the Fig. 1. The outlet channel 13, the connecting bore 24b and the blind bore 24a have no throttling function; thus, the pressure chamber 11 has the same pressure level as the outflow area 22: p 11 = p 22 .

[0026] The following pressure conditions therefore result: The pressure profile on the front side 23a of the impeller 23 shows a continuous drop in pressure from the inlet pressure p. 21 to the outflow pressure p 22 with decreasing radius r.

[0027] The pressure curve on the back side 23b shows the constant inlet pressure p. 21 In the first area 231, a steady drop in pressure across pressure divider 9 from the inlet pressure p 21 for the compensation pressure p 11 , a constant compensation pressure p 11 at the second area 232 and a pressure jump at the shaft seal 25 from the compensation pressure p 11 relative to ambient pressure p 50 The pressure divider 9 is therefore designed as a steam-lubricated throttle between the inlet area 21 and the pressure chamber 11. Depending on the design of the turbine 20 and the operating point at which it is operated, the compensation pressure p can 11smaller, but also the same size or larger than the ambient pressure p 50 be. In the present case, the compensation pressure is p. 11 still equal to the outflow pressure p 22 .

[0028] The steam-lubricated throttle can also be considered a throttle with an extremely small flow cross-section. The function of the steam-lubricated pressure divider is therefore essentially to ensure a contactless, defined, and as small a throttling point as possible during operation, in order to guarantee a pressure-dividing function.

[0029] At low rotational speeds of the impeller 23, the sliding ring 31, with its end face 31a, comes into contact with the sliding surface 23d of the impeller 23 due to the spring force of the sliding ring spring 32. The hydraulic connection from the inlet area 21 to the pressure chamber 11 is interrupted. At high rotational speeds of the impeller 23, a lubricating film or a vapor cushion of the working medium forms between the sliding surface 23d and the end face 31a, causing the two surfaces to separate and thus forming a throttling point. Due to the vapor cushion, only a very small amount of leakage occurs from the inlet area 21 into the pressure chamber 11.

[0030] The operating principle of the well-known Turbine 20 is as follows: During operation of turbine 20, there is a pressure gradient on the front face 23a of the impeller 23. If turbine 20 is operated as a radial turbine, the pressure drops from the radially outer inlet region 21 to the radially inner outlet region 22 from inlet pressure p 21 to outlet pressure p 22 . If the turbine 20 is operated as a pump, the flow direction of the working medium is reversed and accordingly the two areas inlet area 21 and outlet area 22 are arranged in opposite directions; nevertheless, there is also a pressure gradient on the front 23a for these versions.

[0031] The pressure differential on the front side 23a generates an axial force on the impeller 23, i.e., in the case of a radial turbine, a force oriented opposite to the outflow direction 22a. According to the invention, the pressure on the back side 23b of the impeller 23 is now designed by means of the pressure divider 9 such that this pressure generates a counterforce approximately equal to the axial force acting on the front side 23a, so that the resulting hydraulic force acting on the impeller 23 in the axial direction is approximately zero. Consequently, simple and therefore cost-effective and space-saving axial bearings can be used for the impeller 23 or the shaft 24. Advantageously, the pressure on the back side 23b is designed such that the lowest possible resulting hydraulic force is present in a fixed direction across all operating conditions, so that the corresponding axial bearing only has to absorb forces in one direction.

[0032] The pressure applied to the back side 23b is divided into the first area 231 with inlet pressure p. 21 and the second area 232, the compensation pressure p 11 or in the execution of the Fig. 1. Outflow pressure p 22 exhibits.

[0033] When the turbine 20 is put into operation, a local gas / vapor pressure builds up between the impeller 23 and the sliding ring 31 in the pressure divider 9, the resulting force of which causes the sliding ring 31 to float. During operation of the turbine 20, the floating sliding ring 31 therefore performs axial movements, so that the sealing ring 33 is also subjected to corresponding stress. According to the invention, the sealing ring 33 is arranged and guided in such a way that its wear is minimized and the vibration dynamics of the sliding ring 31 and impeller 23 remain stable. [Figure 1] Fig. 2 a corresponding section of the turbine 20 in the area of ​​the sealing ring 33 in section.

[0034] In the execution of the Fig. 2 The turbine 20 has a clamping ring 40 which is firmly connected to the housing 26, for example, by being pressed into it. The sealing ring 33 is arranged in a groove 41 of the clamping ring 40 and therefore interacts with the clamping ring 40 and the sliding ring 31. The sliding ring 31 is preloaded against the impeller 23 by the sliding ring spring 32, but not in a sealing manner. Instead, the pressure divider 9, in the form of a hydrodynamic gap or throttle, is formed between the sliding ring 31 and the rear surface 23b of the impeller 23 during operation of the turbine 20.

[0035] In advantageous embodiments, a stop ring 49 is arranged on the clamping ring 40, which limits a maximum axial displacement of the sliding ring 31, thus representing a quasi axial bearing of the sliding ring 31.

[0036] The sealing ring 33, preferably made of elastomer material, has a sealing and damping function between the sliding ring 31 and the clamping ring 40 for the axial movement of the sliding ring 31. The pressure of the sealing ring 33 must be sufficiently high to enable the sealing function, but must also not be too high so as not to impede the required axial movement of the sliding ring 31.

[0037] Fig. Figure 3 shows section III of the Fig. 2. In Fig. Figure 3 shows the flow conditions in the area of ​​the sliding ring 31. The first flow path 51 runs from the inlet area 21 via the pressure divider 9 or through the throttle into the pressure chamber 11, and the second flow path 52 runs from the inlet area 21 to the sealing ring 33, which seals the second flow path 52 to the pressure chamber 11. If the second flow path 52 were not sealed, it would continue into the pressure chamber 11.

[0038] The sealing ring 33 is thus exposed on one side to the inlet pressure p 21 of the inflow area 21, and on its other side with the compensation pressure p 11of the pressure chamber 11. Pressure changes, especially in the inlet area 21, result in axial movement of the sliding ring 31, as illustrated by arrow 60. If this axial movement of the sliding ring 31 is not possible due to the rolling of the sealing ring 33, because the sealing ring 33 is pressed into the groove 41 on one side by pressure build-up in the second flow path 52, a sliding movement occurs between the sealing ring 33 and the sliding ring 31. The force required for this sliding movement depends on the coefficient of friction between the two friction partners, sealing ring 33 and sliding ring 31, and the corresponding normal force or contact force, i.e., the static force or frictional force between the friction partners. The greater the pressure with which the sealing ring 33 is pressed into the groove 41, the greater the normal force and thus also the static force or frictional force.

[0039] In operation, two types of movement of the sealing ring 33 are distinguished, whereby these two types of movement lead to the best possible sealing function of the sealing ring 33 with minimal wear, whereby a design of the groove 41 as for example according to Fig. 4 is required: - If the axial movement of the sliding ring 31 is to cause the sealing ring 33 to roll, then the frictional force and thus also the pressure on the sealing ring 33 must be large enough to prevent the sealing ring 33 from slipping in the groove 41. This is the case when the inlet pressure p 21 is comparatively low. - If the axial movement of the sliding ring 31 is to cause the sealing ring 33 to slip or slide, then the frictional force and thus also the pressure on the sealing ring must be comparatively small to allow slippage. This is the case when the inlet pressure p 21 is comparatively high.

[0040] Fig. Figure 4 shows a cross-sectional view of the inventive design of the J-shaped, concave groove 41 of the clamping ring 40, wherein the groove 41 could alternatively also be formed in the housing 26. The sealing ring 33 is subjected to pressures from three chambers around its circumference: - A first space 71 is formed between groove 41, sealing ring 33 and sliding ring 31 towards the inlet area 21 and accordingly with the inlet pressure p 21 imposed. - A second chamber 72 is formed between sealing ring 31, clamping ring 40 and sliding ring 31 towards the pressure chamber 11 and is accordingly filled with the compensation pressure p 11 imposed. - A third space 73 represents a kind of dead volume and is formed between clamping ring 40 and sealing ring 33. This third space 73 is created when the inlet pressure p 21so high that the sealing ring 33 is subsequently pushed into a tapered section 42 of the groove 41. If the third chamber 73 is hydraulically separated from the pressure chamber 11, a dynamic pressure p prevails in it. 73 , otherwise the compensatory pressure p prevails within him 11 of pressure room 11.

[0041] Fig. Figure 4 further shows the reaction forces of the sealing ring 33 to its contact partners at a comparatively high inflow pressure p. 21 , if the sealing ring 33 is at least partially pressed into the tapered area 42: - An internal sealing force F i 81 acts between the sealing ring 33 and the sliding ring 31. - An external sealing force F a 82 acts between the sealing ring 33 and the clamping ring 40. - Another sealing force F p 83 acts between the sealing ring 33 and the tapered area 42.

[0042] In order to seal the clamping ring 40 to the sliding ring 31, or to close the second flow path 52 at the end, the internal sealing force F must i 81 and the external sealing force F a 82 must be large enough so that the sealing ring 33 can perform its sealing function.

[0043] An increase in the inlet pressure p 21 In the first chamber 71, this results in the sealing ring 33 being rolled or pushed towards the tapered area 42, but also in it being compressed, i.e., its diameter 85 being reduced in the direction of movement 60. This in turn would result in free expansion perpendicular to this direction, so that in the tensioned state the internal sealing force F i 81 and the external sealing force F a 82 would be increased.

[0044] Due to the design of the groove 41 according to the invention, the internal sealing force F is increased when the sealing ring 33 is inserted into the tapered area 42. i81, however, is reduced again because the center point M of the cross-section of the sealing ring 33 moves away from the sliding ring 31. Thus, in all operating states of the turbine 20, the sliding ring 31 can follow the movement of the impeller 23 without being restrained by the frictional force against the sealing ring 33. On the contrary: with increasing rolling inwards towards the tapered area 42, the internal sealing force F is reduced. i 81 reduced, so that the movement state of the sealing ring 33 can transition into slipping. The associated increase in the external sealing force F a 82 has a very small, negligible influence on efficiency and wear behavior.

[0045] By inserting the sealing ring 33 into the tapered area 42, more precisely with the build-up of the further sealing force F p83 or with the contact between sealing ring 33 and ramp area 43, the third chamber 73 separates from the second chamber 72, and the dynamic pressure p builds up. 73 which increases with increasing insertion. The inflow pressure p then increases 21 If the sealing ring 33 is depressed due to the back pressure p 73 pushed out of the tapered area 42, the sealing ring 33 is essentially automatically reset.

[0046] By designing the groove 41 as in the execution of the Fig. As described in section 4, this results in the following advantages for the turbine 20: - At low inlet pressure p 21 The sealing ring 33 rolls in the groove 41 between the clamping ring 40 and the sliding ring 31. - A sufficiently minimal pressure for the sealing function between sealing ring 33 and sliding ring 31 is always maintained, unless the sealing ring 33 also has the functionality of a pressure relief valve. - On the other hand, the pressure between sealing ring 33 and sliding ring 31 never increases to such an extent that it impedes the movement of the sliding ring 33. As a result, the pressure divider 9 has a constant throttling function, since the sliding ring 31 can follow the impeller movement. - The deformation of the sealing ring 33 is minimized, resulting in an almost symmetrical stress state, which in turn is very advantageous for the sealing function and service life. - Due to the pressure cushion caused by the dynamic pressure p 73 The return movement of the sealing ring 33 out of the groove 41 is pressure-assisted and therefore very efficient and low-wear.

[0047] Fig. Figure 5 shows a detailed cross-sectional view of the groove 41 in a preferred embodiment. In this embodiment, the sealing ring 33 is completely pushed into the tapered section 42, so that the volume of the third chamber 73 is reduced almost to zero; preferably, however, a residual volume of the third chamber 73 with a comparatively high dynamic pressure p is maintained. 73 present. In this position, the sealing ring 33 no longer seals the second chamber 72 against the first chamber 71, since the sealing ring 33 no longer has contact with the clamping ring 31. The illustration of the sealing ring 33 in the Fig. Figure 5 is merely schematic; the sealing ring 33 is shown in an exaggeratedly compressed form.

[0048] In the position shown for the sealing ring 33, very high inlet pressure p 21 in the first chamber 71, and thus also in the inflow area 21, the inflow pressure p 21This can be reduced by opening the second flow path 52. Then, as a result of the leakage through the second flow path 52, the inlet pressure p decreases. 21 In the first chamber 71, the sealing ring 33 comes into contact with the hydraulic force due to the dynamic pressure p. 73 - emerges again from the tapered section 42 and seals again towards the sliding ring 31. This process is therefore reversible, so that the sealing ring 33 has the additional function of a pressure relief valve.

[0049] Due to the favorable design of the groove 41, pressing the sealing ring 33 into the tapered area 42 results in only a relatively small deformation of the sealing ring 33. The resealing of the second flow path 52 or the sliding out of the tapered area 42 thus occurs relatively unimpeded, so that the movements of the sealing ring 33 are very robustly repeatable.

[0050] Fig.Figure 5 further shows particularly advantageous geometries of the J-shaped groove 41 for a turbine 20 with an inlet pressure p 21 of up to 40 bar and a compensation pressure p 11 of approximately 1 bar. The sliding ring 31 has an outer diameter D 31 90 of approximately 42 mm, and the groove 41 has a diameter D 41 91 of approximately 45.5 mm. This results in a ring diameter of the sealing ring 33 - i.e., a diameter of its cross-section or the cord thickness - of more than 1.75 mm, since the sealing ring 33 must be pressed in with an interference fit between the groove 41 and the sliding ring 31 in order to fulfill the sealing function.

[0051] The tapering area 42 extends over a height h 42 93 of approximately 1.7 mm, so that when the second flow path 52 is open, an annular gap with a gap height of approximately 0.05 mm is formed. On the side opposite the tapered section 41, the groove 41 has a lower height h 4192 of approximately 1.25 mm. The groove 41 has a total width b 41 94 of approximately 3 mm, with the tapering area 42 having a width b 42 95 of approximately 0.7 mm.

[0052] The inner contour of the tapered section 42 is designed to optimize the insertion of the sealing ring 33: In the ramp section 43, the inner contour has a linear ramp 44, i.e., a straight contour section, so that the sealing ring 33 can be pushed along this surface towards the third chamber 73 without requiring excessive force. Adjoining this, the inner contour has a circular section 45, i.e., a circular contour section, against which the sealing ring 33 can conform in extreme cases. The circular section 45 preferably has a smaller diameter than the sealing ring 33 itself, so that the third chamber 73 can be formed.

[0053] The ramp section 43 is ideally provided with a radius 46, for example 0.1 mm, to prevent damage to the sealing ring 33 when it is inserted into the tapered section 42. Preferably, the ramp section 43 or the linear ramp 44 rises at an angle 96° of 30° to 40°, which optimizes the reduction of the internal sealing force F. i 81 at increasing inlet pressure p 21 guaranteed.

[0054] Upon contact between the sealing ring 33 and the tapered section 42, the sealing ring 33 rolls onto the inclined plane of the linear ramp 44. The linear ramp 44 acts as a ramp and enables a directed deformation of the sealing ring 33. The radius 46 of the ramp section 43 is intentionally smaller than the sealing ring radius – or half the cord thickness – and allows the sealing ring 33 to shift into the tapered section 42, thereby building up pressure on the back side of the sealing ring 33, i.e., generating the dynamic pressure p. 73in the third room 73.

[0055] Preferably, the turbine 20 according to the invention is arranged in a waste heat recovery system of an internal combustion engine. The waste heat recovery system has a circuit carrying a working fluid, which, in the direction of flow of the working fluid, includes a feed fluid pump, an evaporator, the turbine 20, and a condenser. The working fluid can be fed into the circuit as needed from a collection tank via a branch line and a valve arrangement. Alternatively, the collection tank can also be integrated into the circuit.

[0056] The evaporator is connected to an exhaust pipe of the internal combustion engine, thus utilizing the heat energy of the exhaust gas of the internal combustion engine.

[0057] Liquid working fluid is pumped by the feed fluid pump, possibly from the collection tank, into the evaporator and vaporized there by the heat energy of the exhaust gas from the internal combustion engine. The vaporized working fluid is then expanded in the expansion machine or turbine 20, releasing mechanical energy, for example, to a generator or a gearbox. The working fluid is then liquefied again in the condenser and returned to the collection tank or fed back to the feed fluid pump.

Claims

[1] Turbine (20) with an impeller (23) arranged in a casing (26), wherein the turbine (20) has an inlet region (21) and an outlet region (22) and is supplied with a working medium during operation, the working medium flowing into the inlet region (21), along a front face (23a) formed on the impeller (23) and subsequently out of the outlet region (22), wherein a pressure gradient exists on the front face (23a) between the inlet region (21) and the outlet region (22), wherein a pressure divider (9) is arranged on the rear face (23b) of the impeller (23) opposite the front face (23a), wherein the pressure divider (9) comprises a sliding ring (31) which interacts with the rear face (23b) of the impeller (23) and thus forms a steam-lubricated throttle, wherein a first flow path (51) passes through the throttle, the throttle lubricating the rear face (23b) divides into a first area (231) and a second area (232),wherein the first region (231) limits the inflow region (21) and wherein the second region (232) limits a pressure chamber (11), wherein in operation the inflow region (21) is subjected to a higher pressure than the pressure chamber (11), wherein the sliding ring (31) is axially movable, wherein a sealing ring (33) arranged in a groove (41) interacts with the sliding ring (31), wherein a second flow path (52) runs between the groove (41) and the sliding ring (31) from the inflow region (21) to the pressure chamber (11), wherein the second flow path (52) can be closed off by the sealing ring (33), and wherein the sealing ring (33) is movable in a defined manner in the groove (41). characterized by , that the groove (41) is formed in a clamping ring (40), wherein the clamping ring (40) is firmly connected to the housing (26). [2] Turbine (20) according to claim 1, characterized by , that the pressure chamber (11) is hydraulically connected to the outflow area (22). [3] Turbine (20) according to claim 1 or 2, characterized by , that the groove (41) is J-shaped, wherein the groove (41) has a tapered area (42). [4] Turbine (20) according to claim 3, characterized by , that the tapered area (42) has a ramp area (43) for inserting the sealing ring (33) into the tapered area (42). [5] Turbine (20) according to claim 4, characterized by , that a linear ramp (44) for guiding the sealing ring (33) is formed on the ramp area (43), wherein the linear ramp (44) is inclined by 30° to 40° relative to the sliding ring (31). [6] Turbine (20) according to any one of the preceding claims, characterized by , that the turbine (20) is designed as a radial turbine.

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