Sealing arrangement for a turbomachine, turbomachine, and method of operating a turbomachine

Abstract

A sealing arrangement (100) for a turbomachine is described. The sealing arrangement (100) includes a rotatable shaft (110) supportable in a bearing housing (150), the shaft having a bearing surface arrangeable in the bearing housing (150) to form a plain bearing, a shaft seal assembly (120) axially arrangeable between the plain bearing being a low-pressure side (122), and a rotating wheel of the turbomachine being a high-pressure side (124). The shaft seal assembly (120) includes a first seal (130) radially arrangeable between the bearing housing (150) and the shaft (110), the first seal (130) including at least one sealing ring (132) for being arranged between the bearing housing and the shaft. The shaft (110) has at least one groove (140) for accommodating the at least one sealing ring (132). The sealing ring (132) includes a chamfered portion and / or the groove (140) comprises an undercut portion. A first diameter d1 of the shaft (110) adjacent the at least one groove (140) on the low-pressure side (122) of the groove is larger than a second diameter d2 of the shaft (110) adjacent the at least one groove (140) on the high-pressure side (124) of the groove. The sealing arrangement (100) is configured for reducing aerodynamic forces acting on the sealing ring (132) during operation of the turbomachine so that the aerodynamic forces do not exceed clamping forces between the sealing ring (132) and the bearing housing (150).

Description

[0001] Sealing arrangement for a turbomachine, turbomachine, and method of operating a turbomachine

[0002] Aspects of the invention relate to turbomachines, particularly exhaust gas turbochargers. Aspects of the invention particularly relate to a sealing arrangement for a turbomachine having sealing rings to seal a high-pressure side of the turbomachine from a low-pressure side, such as a bearing housing of the turbomachine.

[0003] Technical background:

[0004] A turbomachine, such as an exhaust gas turbocharger, is often provided for an internal combustion engine, particularly to improve efficiency and power density of the internal combustion engine. A turbocharger generally includes a turbine in the exhaust pipe of the internal combustion engine and a compressor located before the internal combustion engine, which is connected to the turbine via a common shaft. With the boosting of the internal combustion engine by means of the exhaust gas turbocharger, the filling quantity in the cylinder and thus the fuel mixture is increased and thus a significant power boost for the engine is obtained. Alternatively, the energy combined in the exhaust of the internal combustion engine can be converted into electrical energy or mechanical energy by means of a power turbine. Instead of a compressor as in the case of an exhaust gas turbocharger, a generator or mechanical load may be coupled to the turbine shaft.

[0005] Turbomachines typically include sealing arrangements for sealing a chamber housing a wheel, such as a turbine or compressor wheel, from an interior space in a bearing housing of the turbomachine. Sealing arrangements including sealing rings, such as piston ring seals, are commonly used for turbomachines both on the compressor and turbine side. A piston ring seal can reliably control blow-by of exhaust gases or compressed air, and may additionally offer reduction of the oil leakage in case oil is reaching the sealing area during operation.

[0006] Generally, the clamping force of the sealing ring is chosen high enough so that gas forces of the blow-by gases are not sufficiently high to overcome the clamping force and axially shift the sealing ring to come in contact with the shaft, particularly a sidewall of a sealing ring groove provided in the shaft. In such designs, while generally stable during steady operation, the piston ring may have short contacts with the sealing ring groove when the rotor of the turbocharger moves axially due to changes in operation condition, e.g. during run up. A disadvantage of the design is that the required clamping force may require providing large sealing rings, require multiple sealing rings having different sizes, diameters and / or mechanical properties, and an increased wear during transient operating conditions that cause the shaft to move axially and cause the sealing ring groove to come in contact with the sealing ring. This may lead to a large, costly and / or complicated design of the turbomachine, and increased maintenance requirements.

[0007] The present invention solves the above-stated problems at least in part.

[0008] Summary of the invention

[0009] The invention is set out in the appended set of claims.

[0010] According to an aspect, a sealing arrangement for a turbomachine is described. The sealing arrangement includes a rotatable shaft supportable in a bearing housing, the shaft having a bearing surface arrangeable in the bearing housing to form a plain bearing, a shaft seal assembly axially arrangeable between the plain bearing being a low-pressure side, and a rotating wheel of the turbomachine being a high-pressure side. The shaft seal assembly includes a first seal radially arrangeable between the bearing housing and the shaft, the first seal including at least one sealing ring for being arranged between the bearing housing and the shaft. The shaft has at least one groove for accommodating the at least one sealing ring. The sealing ring includes a chamfered portion and / or the groove comprises an undercut portion. A first diameter of the shaft adjacent the at least one groove on the low-pressure side of the groove is larger than a second diameter of the shaft adjacent the at least one groove on the high-pressure side of the groove. The sealing arrangement may be configured for reducing aerodynamic forces acting on the sealing ring during operation of the turbomachine so that the aerodynamic forces do not exceed clamping forces between the sealing ring and the bearing housing. An aerodynamic force acting on the sealing ring during operation of the turbomachine may be lower than a clamping force between the sealing ring and the bearing housing.

[0011] According to an aspect, a turbomachine is described. The turbomachine may be an exhaustgas turbocharger or a power turbine. The turbomachine includes at least one impeller arranged on a shaft, and a bearing housing in which the shaft is rotatably mounted. A shaft seal assembly according to aspects and / or embodiments is arranged between the bearing housing and the shaft.

[0012] According to an aspect, a method of operating a turbomachine is described. The method includes sealing a cavity of a bearing housing from a wheel backspace of a rotating wheel of the turbomachine with a shaft seal assembly. The shaft seal assembly includes a first seal radially arranged between the bearing housing and a shaft. T ne first seal includes at least one sealing ring arranged between the bearing housing and the shaft. The shaft has at least one groove for accommodating the at least one sealing ring. The sealing ring includes a chamfered portion and / or the groove comprises an undercut portion. A first diameter of the shaft adjacent the at least one groove on the low-pressure side of the groove is larger than a second diameter of the shaft adjacent the at least one groove on the high-pressure side of the groove. The shaft seal assembly may be configured for reducing aerodynamic forces acting on the sealing ring during operation of the turbomachine so that the aerodynamic forces do not exceed clamping forces between the sealing ring and the bearing housing. An aerodynamic force acting on the sealing ring during operation of the turbomachine may be lower than a clamping force between the sealing ring and the bearing housing. An aerodynamic force acting on the sealing ring is limited to not exceed a static friction force caused by a clamping force pushing the sealing ring against the bearing housing.

[0013] According to an aspect, a sealing ring is described. In the art, the sealing ring may be referred to as a piston ring. A sealing ring is typically provided within a groove of a rotating shaft. The sealing ring may be clamped against a bearing housing and remain stationary while the shaft rotates. Accordingly, a sealing ring being provided within a groove of the shaft should not necessarily be understood as the sealing ring being in contact with the shaft. For example, the sealing ring may be provided so that it does not contact the groove, does not contact a bottom surface of the groove, or only contacts one of the sidewalls of the groove.

[0014] The clamping force pushing the sealing ring against the bearing housing causes a static friction It is generally intended that the static friction secures the sealing ring against axial movement. Accordingly, aerodynamic and / or hydrostatic forces caused by e.g. a pressure difference between a high-pressure side and a low-pressure side should generally not overcome the static friction during operation of the turbomachine.

[0015] According to an aspect, the sealing ring may include a chamfered portion. A chamfered portion may be understood as a sealing ring having a cross-section that differs from a rectangular cross-section. For example, a chamfered sealing ring may have a portion of a cross-section that is trapezoidal, an axial width of the cross-section of the sealing ring becoming smaller in a radially inward direction, i.e. towards the groove. For example, the sealing ring may be a sealing ring with a double inside chamfer, or a double trapezoid ring. For example, a sealing ring having a chamfered portion may include at least one chamfer reducing, e.g. gradually and / or continuously, an axial thickness of the sealing ring radially towards the inside of the sealing ring. According to an aspect, the groove may include an undercut portion. An undercut portion may be understood as the groove having a cross section that widens radially inwards and / or towards the inside of the groove, i.e. towards an axis of the shaft.

[0016] A chamfered portion of the sealing ring and / or an undercut portion of the groove may allow gas from high pressure side to enter a portion the groove. Accordingly, the gas may exert a pressure onto the sealing ring from more than one side, which may beneficially counteract a force pushing the sealing ring in an axial direction, i.e. axially towards the low-pressure side.

[0017] Beneficially, embodiments of the present invention allow slimmer and / or smaller sealing rings to be used, since the required clamping force is reduced. Beneficially, in case two or more sealing rings are used, both sealing rings may have the same size and / or mechanical properties, which may simplify the design of the turbomachine. Beneficially, e.g. due to reducing the required clamping forces, a wear of the piston ring and / or shaft is reduced in case of contact between the shaft and the sealing ring e.g. due to axial movement of the shaft, since the mechanical forces required to axially shift the sealing ring are lower.

[0018] Brief description of the Figures:

[0019] The details will be described in the following with reference to the figures, wherein

[0020] Fig. 1 is a schematic cutaway side view of a sealing arrangement according to embodiments on a turbine side of a turbomachine;

[0021] Fig. 2 is a schematic cutaway side view of a sealing arrangement according to embodiments on a compressor side of a turbomachine

[0022] Fig. 3A shows a sealing ring having a chamfered portion in a groove;

[0023] Fig. 3B shows a sealing ring in a groove having an undercut portion extending to a sidewall of the groove;

[0024] Fig. 3C shows a sealing ring in a groove having an undercut portion forming an edge;

[0025] Fig. 4A is a graph showing gas forces acting on a sealing ring in a conventional sealing arrangement;

[0026] Fig. 4B is a graph showing gas forces acting on a sealing ring in a sealing arrangement according to embodiments. Detailed description of the Figures and of embodiments:

[0027] Reference will now be made in detail to the various embodiments, one or more examples of which are illustrated in each figure. Each example is provided by way of explanation and is not meant as a limitation. For example, features illustrated or described as part of one embodiment can be used on or in conjunction with any other embodiment to yield yet a further embodiment. It is intended that the present disclosure includes such modifications and variations.

[0028] Within the following description of the drawings, the same reference numbers refer to the same or to similar components. Generally, only the differences with respect to the individual embodiments are described. Unless specified otherwise, the description of a part or aspect in one embodiment applies to a corresponding part or aspect in another embodiment as well.

[0029] Reference is made to European patent application EP 2 375 000 A2 which shows a turbomachine and a sealing arrangement thereof. Aspects of the document, particularly aspects relating to the sealing arrangement shown in Fig. 2 of the document, may help in understanding the present invention, and are incorporated herein by reference.

[0030] Reference is made to European patent application EP 2 743 460 A1 which shows a turbomachine and a sealing arrangement thereof. Aspects of the document, particularly aspects relating to the sealing arrangement shown in Fig. 4 of the document, may help in understanding the present invention, and are incorporated herein by reference.

[0031] Referring now to Fig. 1 , a sealing arrangement 100 according to embodiments is described. The sealing arrangement 100 may be a sealing arrangement of and / or for a turbomachine, such as a turbocharger. In the example shown in Fig. 1, the sealing arrangement is provided on a turbine side of the turbomachine. To help in understanding, in the following, the sealing arrangement 100 will be explained with reference to the turbomachine.

[0032] The turbomachine includes a rotatable shaft 110 supported in a bearing housing 150. The shaft 110 is supported in a bearing, such as a plain bearing (not shown). The bearing is formed between a bearing surface of the shaft and a bearing surface of the bearing housing 150. In the example shown in Fig. 1 , the bearing is provided on the low-pressure side 122. The sealing arrangement may include further components of the turbomachine, such as a bearing housing cavity formed in-between the bearing and the shaft seal arrangement 120.

[0033] The sealing arrangement 100 includes a shaft seal assembly 120. The shaft seal assembly separates a high-pressure side 124 of the turbomachine from a low-pressure side 122. The high-pressure side may be defined by a presence of a wheel (not shown) of the turbomacmne, such as a turbine impeller, a chamber housing the wheel and / or a wheel backspace being fluidly connected to the shaft seal assembly 120 on the high-pressure side. The low-pressure side may be defined by an enclosed volume of the turbomachine, such as the bearing housing cavity.

[0034] During operation of the turbomachine, a pressure difference may build up between the high- pressure side 124 and the low pressure side 122, e.g. through action of the wheel, the introduction of pressurized gas into the chamber housing the wheel, or even a sealing gas being introduced on the high-pressure side 122 close to the sealing arrangement 120. The pressure difference, i.e. an overpressure of the high-pressure side 124 in relation to the low- pressure side 122, may be in a range of 2 bar or less, 3 bar or less, 4 bar or less, 5 bar or less, 7 bar or less, or even 10 bar or less, depending on the operating conditions and / or specifications of the turbomachine. It is understood that a turbomachine rated for a defined pressure difference should include a sealing arrangement 100 configured for withstanding at least the rated pressure difference, even if this rated pressure difference only occurs under some operating conditions of the turbomachine.

[0035] The shaft seal assembly 120 includes a first seal 130 radially arranged between the bearing housing 150 and the shaft 110. A first sealing ring 132 is arranged and / or accommodated in a groove 140 of the shaft 110.

[0036] As shown in Fig. 1 , the shaft seal assembly 120 may include a second seal, the second seal including a second sealing ring 134 accommodated in a second groove 142 of the shaft 110. The second seal may be optional. Additional seals, such as a third or a fourth seal may be provided. Accordingly, the shaft seal assembly may include at least two axially spaced sealing rings arranged between the bearing housing 150 and the shaft 110. The sealing rings 132, 134 may be individually provided in separate grooves 140, 142.

[0037] As shown in Fig. 1, the groove 140 may include undercut portions 141. Likewise, the groove 142 may include undercut portions 143, and features described with reference to the groove 140 may apply to the groove 142. The undercut portions 141 allow gases entering the groove 140 from the high-pressure side 124 to exert a pressure on the sealing ring 132 onto a portion of the sidewall on the low-pressure side 122 of the sealing ring 132. A pressure of the gases may drop to the pressure of the low-pressure side at the portion of the sidewall of the sealing ring 132 where the sidewall comes closest to the sidewall of the groove 140 and / or where the sidewall comes into contact with the groove 140. Accordingly, the undercut portions 141 beneficially causes some or even most of the aerodynamic forces exerting pressure onto tne sealing ring 132 in an axial direction to cancel each other out, and only a portion of axial forces pushing the sealing ring 132 axially towards the low-pressure side remains. In particular, the aerodynamic and / or hydrostatic forces acting on the sealing ring during operation of the turbomachine do not exceed clamping forces, and / or a static friction resulting from the clamping forces, between the sealing ring and the bearing housing. In particular, the sealing ring, e.g. dimensions and / or material properties of the sealing ring, may define the clamping forces, and a sealing ring may be chosen according to the expected aerodynamic forces and / or the axial force resulting from the aerodynamic forces and the properties of the groove and / or the sealing ring. Beneficially, the sealing arrangement 100 and / or shaft seal assembly 120 described herein may, e.g. due to a combination of providing the undercut portions 141 and / or chamfered portion, and the first diameter d-, being larger than the second diameter d2, cause aerodynamic forces acting on the sealing ring to be reduced with respect to known sealing arrangement and / or shaft seal assemblies, particularly sealing arrangements and / or shaft seal assemblies in which the first diameter d! is equal to the second diameter d2.

[0038] According to embodiments, as shown in Fig. 1 , the undercut 141 of the groove 140 may extend from a bottom surface of the groove to a sidewall within the groove. An exemplary arrangement 320 of a sealing ring provided in a groove having an undercut extending from a bottom of the groove to a sidewall within the groove is shown in further detail in Fig. 3B. The sidewall may form a portion of the groove that extends radially inwards from an outer surface of the shaft into the groove. Accordingly, the undercut 141 may transition into the sidewall portion of the groove 140 within the groove. The sidewall portion of the groove 140 may extend essentially parallel to a sidewall of the sealing ring 132.

[0039] Alternatively, according to some embodiments, the undercut may extend from a bottom surface of the groove to the top of the groove, i.e. the outer surface of the shaft 110. In the alternative embodiment, the groove may be devoid of a sidewall portion of the groove 140 and / or the undercut may form a (sharp) edge. An exemplary arrangement 330 of a sealing ring provided in a groove having an undercut extending from a bottom of the groove and forming an edge is shown in Fig. 3C.

[0040] Additionally, or alternatively to the groove having undercut portions 141 , in some embodiments, the sealing ring 132 may include one or more chamfered portions. An exemplary arrangement 310 of a sealing ring having chamfered portions and provided in a groove is shown in Fig. 3A. Similar to the groove having undercut portions 141, a gas may flow within the groove past the chamfered portion and result in a reduced total axial pressure acting on tne sealing ring 132, such as on a projected surface of the sealing ring.

[0041] It should be noted that Fig. 3A, 3B and 3C are given as examples to help in understanding the invention, and may not fall under the scope of the claims.

[0042] As shown in Fig. 1 , a first diameter d-j of the shaft 110 adjacent the groove 140 on the low- pressure side 122 of the groove 140 is larger than a second diameter d2of the shaft adjacent the groove 140 on the high-pressure side 124 of the groove. As shown in Fig. 1 , a diameter d3of the remaining shaft portion of the sealing arrangement may be lower than the diameter drFor example, as shown in Fig. 1 , the diameter d2and the diameter d3may be the same diameter. As shown in Fig. 1 , the diameter d-i being larger on the low-pressure side may cause the undercut 141 of the groove 140 on the low-pressure side 122 to extend radially further outward than the undercut 141 on the high-pressure side 124. Accordingly, a portion of the sidewall of the groove 140 coming closes to the sidewall of the sealing ring 132 and / or contacting the sealing ring 132 may be located radially further outward.

[0043] As shown in Fig. 1 , the bearing housing 150 may have an essentially constant diameter in the region spanning the sealing rings 132, 134. In particular, the sealing rings 132, 134 may be seated in an essentially cylindrical portion of the bearing housing 150. Beneficially, an essentially cylindrical portion of the bearing housing may be easier to manufacture and / or maintain than e.g. a stepped portion configured for having sealing rings with different outer diameters inserted therein.

[0044] According to embodiments, a first radial clearance of the at least one sealing ring 132 provided in the groove 140 and extending from the groove 140 past the outer diameter of the shaft 110 may be smaller than a second radial clearance of the at least one sealing ring 132 provided in the groove 140 and extending from the groove 140 past the outer diameter of the shaft 110, the second clearance being axially opposite the first clearance. The radial clearance may define a length by which the sealing ring 132 extends radially outwards from the groove 140. The first radial clearance may be the clearance on the low-pressure side 122 of the sealing ring 132, and the second radial clearance may be the clearance on the high-pressure side 124 of the sealing ring 132.

[0045] According to embodiments, the first radial clearance 170 formed radially between the shaft 110 and the bearing housing 150 adjacent the at least one groove 140 on the low-pressure side 122 of the groove 140 may be smaller than the second radial clearance 172 formed radially between the shaft 110 and the bearing housing 150 adjacent the at least one groove 140 on the high-pressure side 124 of the groove 140. For example, a ratio or tne first radial clearance 170 and the second radial clearance 172 may be 1:1.5 or larger, particularly 1 :2 or larger.

[0046] The diameter di being larger than the diameter d2, and / or the first radial clearance 170 being smaller than the second radial clearance 172 may, individually or in combination, beneficially increase the area of the low-pressure side sidewall of the sealing ring 132 coming in contact with the high-pressure gas, which may beneficially result in a decrease in the total axial force being generated by the pressure difference between the high-pressure side 124 and the low- pressure side 122. Accordingly, the clamping force of the sealing ring 132 in a bearing arrangement 100 according to embodiments may be lower, which may allow providing e.g. a smaller-sized sealing ring 132 and / or simplify and / or reduce the costs of the design of the sealing arrangement 100.

[0047] As shown in Fig. 1 , in embodiments having more than one sealing ring 132, 134, the shaft portion having the first diameter d-, and the second diameter d2may be located adjacent the groove 140 closes to the low-pressure side 122. It was observed that the sealing ring 132 closest to the low-pressure side 122 often experiences the largest pressure differences due to continuity constraints. Accordingly, the benefits of the invention may be particularly prominent for the sealing ring 132 closest to the low-pressure side 122. Additionally, the axial length of the shaft portion having the larger diameter d-, , i.e. a shaft portion in-between the sealing rings 132, 134 may have the lower diameter d2and / or the larger radial clearance 172. Likewise, the remainder of the shaft of the sealing arrangement may have the lower diameter d2and / or d3. Beneficially, a higher-precision machining required for providing the smaller radial clearance 170 may be limited to only a comparably small portion of the shaft 110 and / or the bearing housing 150.

[0048] According to embodiments, as shown in Fig. 1 , the shaft 110 may include a sealing bush 112. The sealing bush 112 may include the grooves 140, 142. The sealing bush 112 is a separate part provided on the shaft and configured for rotating together with the shaft 110. For example, the sealing bush 112 may be shrunk onto a seat on the shaft 110 and an edge formed on the shaft 110 may serves as an axial stop for the sealing bush.

[0049] According to embodiments, the bearing housing 150 may include a sealing insert. The sealing insert may form a counterpart to the sealing bush 112 provided on the shaft, and may be a separate part from e.g. an outer bearing housing 150. For example, the sealing insert may be installed in the bearing housing during assembly of the bearing housing 150. According to embodiments, as shown in Fig. 1, the shaft 110 and / or tne bearing nousing 15U may have a contour which forms one or more oil discharge chambers 160, Fig. 1 showing a single oil discharge chamber 160, however, more than one oil discharge chamber, such as two or even three oil discharge chambers may be provided. The oil discharge chamber may contribute to the oil tightness of the sealing arrangement 100, particularly for preventing an oil present on the low-pressure side 122 to reach the shaft seal assembly 120. Beneficially, the sealing arrangement 100 may synergistically provide an improved seal having improved oil and gas tightness.

[0050] Referring now to Fig. 2, a sealing arrangement 200 according to embodiments is described. The sealing arrangement 200 shares many of the features of the sealing arrangement 100 described with reference to Fig. 1 , and only the differences will be explained.

[0051] The sealing arrangement 200 may be a sealing arrangement of and / or for a turbomachine, such as a turbocharger. In the example shown in Fig. 2, the sealing arrangement 200 is provided on a compressor side of the turbomachine. Accordingly, some properties and / or requirements of the sealing arrangement 200 may differ from the sealing arrangement 100 provided on the turbine side of the turbomachine.

[0052] For example, the sealing arrangement 200 differs from the sealing arrangement 100 by being devoid of oil discharge chambers provided in the shaft seal assembly 220. In some embodiments, to provide oil tightness, a sealing disc (not shown) may be provided on the low- pressure side 122 to prevent an oil from reaching the sealing ring 132.

[0053] The solutions presented herein may be particularly suitable for the compressor side of a turbomachine, such as a turbocharger, since the pressure difference between the high- pressure side 124 and the low-pressure side 122 may be particularly high on the compressor side of the turbomachine. Additionally, or alternatively, the solutions presented herein may be particularly suitable for a high-pressure turbocharger, such as a second-stage turbocharger in a two-stage system.

[0054] Referring now to Fig. 4A and Fig. 4B (collectively Fig. 4), the benefits of the present invention are demonstrated schematically as an analysis of the pressure profile generated with analytical equations. Fig. 4 shows a graph of the gas forces acting on a sealing ring in a sealing arrangement. Fig. 4 shows a cross-section of a sealing ring, such as the sealing ring 130, in a shaft seal assembly 120. A pressure HP on the high-pressure side is 3 bar. A pressure LP on the low-pressure side is 1 bar. The sealing ring is clamped to a bearing housing surface at a radial top and / or outward surface, e.g. as shown in Fig. 1 and Fig. 2. Accordingly, the pressure gradient is present from HP to LP along the clamping surface. I ne sealing ring exerts a clamping force Fradof 734 N onto the bearing housing, which results in a static friction force Ffnc of 132 N. Accordingly, the sealing ring remains stationary unless an axial force Faxexceeds the static friction force Ffric.

[0055] As shown in Fig. 4, the pressure HP is exerted onto the complete side surface on the high- pressure side and the radial bottom surface of the sealing ring. On the low-pressure side of the sealing ring, the pressure HP drops to the pressure LP at the position at which the sidewall of the sealing ring comes closest to and / or in contact with a sidewall or edge of the groove. Accordingly, only a part of the low-pressure side sidewall of the sealing ring is subjected to the pressure HP, while the remainder of the sidewall of the sealing ring is subjected to a pressure lower than HP and / or the pressure LP.

[0056] Fig. 4A shows a conventional shaft seal assembly 120 in which the shaft diameter is essentially the same at the high-pressure side and the low-pressure side. Fig. 4B shows a shaft seal assembly 120 according to embodiments, in which the diameter di is larger than the diameter d2. Accordingly, as shown in Fig. 4B, the location at which the pressure drops from the pressure HP to the pressure LP is moved radially outwards, and the portion of the sealing ring sidewall on the low-pressure side onto which the pressure HP is exerted is comparably larger. Accordingly, the force acting axially towards the high-pressure side of the sealing ring of Fig. 4A is 361 N, while the force acting axially towards the high-pressure side of the sealing ring according to embodiments of Fig. 4B is 406 N.

[0057] As shown in Fig. 4, an axial force pushing the sealing ring towards the low-pressure side may be calculated and / or approximated by the difference of forces exerted onto the high-pressure side sidewall of the sealing ring by the pressure HP, and the forces exerted onto the low- pressure side sidewall of the sealing ring by the pressures HP and LP. In Fig. 4A, the total axial force Faxis 106 N, while in Fig. 4B, the total axial force Faxis 61 N. Accordingly, the factor LF (Fax / Ffric) of Fig. 4A is 1.24, while the factor LF of Fig. 4B is 2.15.

[0058] It should be noted that Fig. 4 demonstrates the effect of only one changed parameter, i.e. the change in geometry of the shaft 110 and / or the groove 140. Since the axial force Faxin a shaft seal assembly 120 according to embodiments is significantly lower, the sealing ring according to embodiments may beneficially e.g. be dimensioned smaller, slimmer, have a lower diameter, different material properties or the like, while providing a friction force Ffricsufficiently higher than the axial force Fax. For example, as demonstrated in Fig. 4, in some embodiments, both sealing rings may be chosen to have the same diameter, wnicn may beneficially reduce tne cost and / or complexity of the turbomachine.

[0059] According to embodiments, the use of a shaft seal assembly according to embodiments, such as the shaft seal assembly 120 shown in Fig. 1 and / or Fig. 2, in a turbomachine is described. The turbomachine may include at least one impeller arranged on the shaft 110, such as a turbine wheel or a compressor wheel. The turbomachine includes a shaft seal assembly 120 and / or a sealing arrangement according to aspects and / or embodiments described herein. The turbomachine may be an exhaust gas turbocharger. In particular, the turbomachine may be a turbocharger for industrial applications, such as marine applications, such as for a marine diesel engine. The turbomachine may be particularly suitable for engines having a power rating of 1000 kW or above, particularly of 10 MW or above, or even 20 MW or above.

[0060] According to embodiments, a method of operating a turbomachine is described. The method includes sealing a cavity of a bearing housing, such as the bearing housing 150 described with reference to Fig. 1 and / or Fig. 2, from a wheel backspace of a rotating wheel of the turbomachine.

[0061] The shaft seal assembly may be a shaft seal assembly according to aspects and / or embodiments described herein, such as the shaft seal assembly 120 described with reference to Fig. 1 and / or Fig. 2. In particular, the shaft seal assembly may include a first seal 130 radially arranged between the bearing housing 150 and a shaft 110, the first seal 130 including at least one sealing ring 132 arranged between the bearing housing 150 and the shaft (110). The shaft 110 has at least one groove 140 for accommodating the at least one sealing ring 132. The sealing ring includes a chamfered portion and / or the groove includes an undercut portion so that aerodynamic forces acting on the sealing ring do not exceed clamping forces between the sealing ring and the bearing housing. A first diameter d! of the shaft 110 adjacent the at least one groove 140 on the low-pressure side 122 of the groove 140 is larger than a second diameter d2of the shaft 110 adjacent the at least one groove 140 on the high-pressure side 124 of the groove 140.

[0062] According to embodiments, the method may include operating the turbomachine at a pressure difference between the high-pressure side and the low-pressure side of 2 bar or less, 3 bar or less, 4 bar or less, 5 bar or less, 7 bar or less, or even 10 bar or less.

[0063] The method may include directing an oil away from the shaft seal arrangement on the low- pressure side, e.g. by providing a sealing disc and / or one or more oil discharge chambers.

[0064] The method may include providing two sealing rings having essentially the same dimensions. The method may include lowering and / or defining an axial force being exerted onto at least one sealing ring by providing a contact point of the groove and the sealing ring radially further outward on the low-pressure side of the sealing ring than on the high-pressure side of the sealing ring. Beneficially, the solutions provided herein provide a more efficient sealing arrangement in which clamping forces of the sealing rings may be reduced while the sealing ring remains securely provided in a bearing housing. Beneficially, sealing rings having lower clamping forces may allow a less complex design of the sealing arrangement. Beneficially, sealing rings having lower clamping forces may result in less wear in some conditions in which the shaft comes in contact with the sealing ring.

[0065] List of reference numerals

[0066] 100, 200 sealing arrangement

[0067] 110 shaft

[0068] 112 sealing bush

[0069] 120, 220 shaft seal assembly

[0070] 122 low-pressure side

[0071] 124 high-pressure side

[0072] 130 first seal

[0073] 132 first sealing ring

[0074] 134 second sealing ring

[0075] 140 groove

[0076] 141 undercut

[0077] 142 groove

[0078] 143 undercut

[0079] 150 bearing housing

[0080] 160 oil discharge chamber

[0081] 170 first radial clearance

[0082] 172 second radial clearance

[0083] 310 sealing ring arrangement

[0084] 320 sealing ring arrangement

[0085] 330 sealing ring arrangement

Claims

Claims:

1. Sealing arrangement (100) for a turbomachine, the sealing arrangement comprising a rotatable shaft (110) supportable in a bearing housing (150), the shaft (110) comprising a bearing surface arrangeable in the bearing housing (150) to form a plain bearing, and a shaft seal assembly (120) axially arrangeable between the plain bearing being a low-pressure side (122), and a rotating wheel of the turbomachine being a high-pressure side (124), the shaft seal assembly (120) comprising: a first seal (130) radially arrangeable between the bearing housing (150) and the shaft (110), the first seal (130) comprising at least one sealing ring (132) for being arranged between the bearing housing (150) and the shaft, wherein the shaft (110) has at least one groove (140) for accommodating the at least one sealing ring (132), wherein the sealing ring (132) comprises a chamfered portion (141) and / or the groove (140) comprises an undercut portion, and wherein a first diameter (d^ of the shaft (110) adjacent the at least one groove (140) on the low- pressure side (122) of the groove (140) is larger than a second diameter (d2) of the shaft (110) adjacent the at least one groove (140) on the high-pressure side (124) of the groove (140).

2. The sealing arrangement (100) according to claim 1 , wherein a first radial clearance (170) of the at least one sealing ring (132) provided in the groove (140) and extending from the groove (140) past the outer diameter of the shaft (110) is smaller than a second radial clearance (172) of the at least one sealing ring (132) provided in the groove (140) and extending from the groove (140) past the outer diameter of the shaft (110), the second radial clearance (172) being axially opposite the first radial clearance (170).

3. The sealing arrangement (100) according to any one of the preceding claims, comprising at least two axially spaced sealing rings (132, 134) arrangeable between the bearing housing (150) and the shaft (110), the sealing rings (132, 134) being individually provided in separate grooves (140, 142), wherein a shaft portion having the first diameter (di) and a shaft portion having the second diameter (d2) are located adjacent the groove (140) closest to the low pressure side.

4. The sealing arrangement (100) according to any one of the preceding claims, wherein the shaft (110) comprises a shaft sealing bush (112), the shaft sealing bush (112) comprising the at least one groove (140).

5. The sealing arrangement (100) according to any one of the preceding claims, wherein the shaft (110) has a contour which, together with the bearing housing (150), forms at least one oil discharge chamber (160), the oil discharge chamber (160) being arranged along the shaft (110) in-between the plain bearing and the at least one sealing ring (132).

6. The sealing arrangement (100) according to any one of the preceding claims, wherein the shaft seal assembly (120) is configured for a pressure difference during operation of the turbomachine between the high-pressure side (124) and the low-pressure side (122) of between 0 to 7 bar, more particularly 0 to 4 bar.

7. The sealing arrangement (100) according to any one of the preceding claims, wherein the groove comprises the undercut portion comprising at least one undercut widening the groove towards the inside of the groove, and / or wherein the sealing ring comprising a chamfered portion comprises at least one chamfer reducing an axial thickness of the sealing ring radially towards the inside of the sealing ring.

8. The sealing arrangement (100) according to claim 7, wherein the at least one undercut extends from the bottom surface of the groove (140) to a sidewall within the groove (140), the sidewall extending radially inwards from an outer surface of the shaft into the groove.

9. Turbomachine, particularly exhaust-gas turbocharger or power turbine, comprising at least one impeller arranged on a shaft (110), and a bearing housing (150) in which the shaft (110) is rotatably mounted, wherein a shaft seal assembly (120) is arranged between the bearing housing (150) and the shaft (110), the shaft seal assembly comprising:a first seal (130) radially arrangeable between the bearing nousing (150) and tne snatt (110), the first seal (130) comprising at least one sealing ring (132) for being arranged between the bearing housing (150) and the shaft, wherein the shaft (110) has at least one groove (140) for accommodating the at least one sealing ring (132), wherein the sealing ring (132) comprises a chamfered portion (141) and / or the groove (140) comprises an undercut portion, and wherein a first diameter (d^ of the shaft (110) adjacent the at least one groove (140) on the low- pressure side (122) of the groove (140) is larger than a second diameter (d2) of the shaft (110) adjacent the at least one groove (140) on the high-pressure side (124) of the groove (140), and wherein the sealing arrangement (100) is configured for reducing aerodynamic forces acting on the sealing ring (132) during operation of the turbomachine so that the aerodynamic forces do not exceed clamping forces between the sealing ring and the bearing housing.

10. The turbomachine according to claim 9, wherein a first radial clearance (170) formed radially between the shaft (110) and the bearing housing (150) adjacent the at least one groove (140) on the low-pressure side (122) of the groove (140) is smaller than a second radial clearance (172) formed radially between the shaft (110) and the bearing housing (150) adjacent the at least one groove (140) on the high-pressure side (124) of the groove (140).

11. The turbomachine according to claim 10, wherein a ratio of the first radial clearance (170) and the second radial clearance (172) is 1 :1.5 or larger, particularly 1 :2 or larger.

12. The turbomachine according to any one of the claims 9 to 11 , comprising at least two axially spaced sealing rings (132, 134) arranged between the bearing housing (150) and the shaft (110), the sealing rings (132, 134) being individually provided in separate grooves (140, 142), wherein a shaft portion having the first diameter (d^ and a shaft portion having the second diameter (d2) are located adjacent the groove (140) closest to the low pressure side, wherein the bearing housing (150) has an essentially constant diameter in the region spanning the at least two sealing rings (132, 134).

13. The turbomachine according to any one of claims 9 to 12, wherein the bearing housing (150) comprises a sealing insert, the first seal (130) being provided between the sealing insert and the shaft seal assembly (120).

14. Use of the sealing arrangement (100) according to any one of claims 1 to 8 in a turbomachine, particularly the turbomachine according to any one of claim 9 to 13.

15. Method of operating a turbomachine, comprising: sealing a cavity of a bearing housing (150) from a wheel backspace of a rotating wheel of the turbomachine with a shaft seal assembly, the shaft seal assembly comprising: a first seal (130) radially arranged between the bearing housing (150) and a shaft (110), the first seal (130) comprising at least one sealing ring (132) arranged between the bearing housing (150) and the shaft (110), wherein the shaft (110) has at least one groove (140) for accommodating the at least one sealing ring (132), wherein the sealing ring comprises a chamfered portion and / or the groove comprises an undercut portion, wherein a first diameter (d^ of the shaft (110) adjacent the at least one groove (140) on the low- pressure side (122) of the groove (140) is larger than a second diameter (d2) of the shaft (110) adjacent the at least one groove (140) on the high-pressure side (124) of the groove (140), and wherein an aerodynamic force acting on the sealing ring (132) is limited to not exceed a static friction force caused by a clamping force pushing the sealing ring against the bearing housing (150).

Images