CA2668482A1 - Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another - Google Patents
Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another Download PDFInfo
- Publication number
- CA2668482A1 CA2668482A1 CA002668482A CA2668482A CA2668482A1 CA 2668482 A1 CA2668482 A1 CA 2668482A1 CA 002668482 A CA002668482 A CA 002668482A CA 2668482 A CA2668482 A CA 2668482A CA 2668482 A1 CA2668482 A1 CA 2668482A1
- Authority
- CA
- Canada
- Prior art keywords
- bearing
- hydraulic
- sealing ring
- arrangement according
- sealing
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Granted
Links
Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
- F16J—PISTONS; CYLINDERS; SEALINGS
- F16J15/00—Sealings
- F16J15/44—Free-space packings
- F16J15/443—Free-space packings provided with discharge channels
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
- F03B—MACHINES OR ENGINES FOR LIQUIDS
- F03B11/00—Parts or details not provided for in, or of interest apart from, the preceding groups, e.g. wear-protection couplings, between turbine and generator
- F03B11/006—Sealing arrangements
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
- F05B—INDEXING SCHEME RELATING TO WIND, SPRING, WEIGHT, INERTIA OR LIKE MOTORS, TO MACHINES OR ENGINES FOR LIQUIDS COVERED BY SUBCLASSES F03B, F03D AND F03G
- F05B2240/00—Components
- F05B2240/57—Seals
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
- F05B—INDEXING SCHEME RELATING TO WIND, SPRING, WEIGHT, INERTIA OR LIKE MOTORS, TO MACHINES OR ENGINES FOR LIQUIDS COVERED BY SUBCLASSES F03B, F03D AND F03G
- F05B2260/00—Function
- F05B2260/60—Fluid transfer
- F05B2260/602—Drainage
- F05B2260/603—Drainage of leakage having past a seal
-
- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
- Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
- Y02E10/00—Energy generation through renewable energy sources
- Y02E10/20—Hydro energy
Landscapes
- Engineering & Computer Science (AREA)
- General Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- Chemical & Material Sciences (AREA)
- Combustion & Propulsion (AREA)
- Magnetic Bearings And Hydrostatic Bearings (AREA)
- Hydraulic Turbines (AREA)
- Hydraulic Motors (AREA)
Abstract
Owing to the operating states in turbomachinery, effective and at the same time simple seals between parts of turbomachinery such as this which can move relative to one another, for example a rotor of a turbine and a turbine casing, have been impossible, or possible only in a highly restricted form. As a consequence, it was necessary to accept certain efficiency losses caused by gap-water and friction losses. In the case of known seals, in which a sealing element floats on two hydrostatic bearings with respect to parts which can move relative to one another, which hydrostatic bearings are virtually sealed and are nevertheless of very simple design, large sealing water pumps are required, however, in order to make it possible to provide the pressures which are required for operation of a seal such as this. The present invention overcomes this problem by providing a restriction device (37) in the hydraulic connection (30) between the first hydraulic bearing (22) and the second hydraulic bearing (23), which restriction device (37) makes it possible to set a pressure loss of at least 1%, preferably at least 5% and preferably at least 10% between the first hydraulic bearing (22) and the second hydraulic bearing (23), with respect to the pressure difference (po-pu) which is to be sealed and exists in the vicinity of the sealing element (20).
Description
Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another The present invention relates to an arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another, preferably between a housing and a rotor of a hydraulic turbomachine, with a sealing element, which with respect to the two parts is arranged supported on a first and a second hydraulic bearing in a floating manner, wherein the first and the second hydraulic bearing comprise at least two assigned bearing surfaces and the sealing element is supported by the hydraulic compressive forces acting on its bearing surfaces and wherein a hydraulic connection, preferably a connection bore, is provided in the sealing element, which hydraulic connection connects the first hydraulic bearing to the second hydraulic bearing in a hydraulic manner and with a supply line that opens into the first hydraulic bearing.
Due to the very unfavorable operating conditions in the area of the outer diameter of a rotor of a hydraulic machine, and here above all with larger machines with rotor diameters of up to several meters, it has not hitherto been possible to build a reliable seal between rotor and housing, which, in addition to a loss in efficiency through blade-tip leakage losses, can also lead to other major problems. This is due above all to the fact that very high circumferential speeds occur in the area of the outer diameter of the rotor, the rotor as well as the housing are subjected to strong vibrations and the rotor undergoes additional axial shifts due to the high pressures acting. These operating conditions have hitherto prevented the construction of a virtually or ideally completely tight seal.
Seals hitherto used, such as, e.g., labyrinth seals, are not seals in the real sense, but only devices for reducing the leakage water flow, wherein an undesirable leakage water flow occurs in any case, however. Others, such as, e.g., known ice ring seals according to EP 1 098 088 A, were in turn very complex and unreliable in a possible operation, since susceptible to breakdown.
US 5 052 694 A shows a sealing body that is supported on a hydrostatic bearing in one direction. This sealing body is rail-guided in a second direction by means I
of sealing elements. A throttle device is provided in the feed line to the hydrostatic bearing, with which throttle device fluctuations in the gap widths of the hydrostatic bearing are absorbed. Also AT 298 365 B shows a seal with a sealing body that is supported on a hydraulic bearing in one direction and that is again forcedly guided in the other direction by a piston and a conduit. The necessary pressing force for the sealing body is thereby effected by the hydraulic piston.The supply pressure for the hydraulic sealing and for the piston is set by separate, independent throttle devices in their respective supply line.
Likewise seals between two parts moved relative to one another by means of a hydrostatic bearing are known from the prior art. A sealing element is thereby supported on a hydrostatic bearing in one direction between the two parts moved relative to one another. In a second direction the sealing element is supported on a hydraulic bearing. The hydraulic bearing thereby comprises respectively two sealing bodies, such as, e.g., 0 rings, between which a pressure medium is fed.
Examples of seals of this type can be taken from US 4 118 040 A or GB 839 880.
In the axial rotor side space of a turbine, i.e., between hub and housing, a sometimes very high pressure would form through the leakage water flow and the seal between shaft and rotor side space (e.g., by means of gland seals), which high pressure essentially corresponds to the headwater pressure, and which would tend to displace the rotor in the axial direction, whereby large axial loads of the bearing of the rotor and relatively large axial shifts of the rotor would occur, which is avoided through a pressure reduction with the aid of a known leakage water line in the tailwater. A leakage water flow develops in both rotor side spaces, that is, between housing and ring or hub, whereby a certain proportion of the medium does not flow through the rotor, thus resulting in a drop in efficiency and a loss of power. Furthermore, a water disk rotating very quickly forms in both rotor side spaces, which disk of water counteracts the rotation of the shaft due to the developing friction and thus develops a braking effect, which in turn further reduces efficiency. For these reasons it is desirable to provide a virtually completely tight seal between rotor and housing which eliminates the leakage water flow.
Due to the very unfavorable operating conditions in the area of the outer diameter of a rotor of a hydraulic machine, and here above all with larger machines with rotor diameters of up to several meters, it has not hitherto been possible to build a reliable seal between rotor and housing, which, in addition to a loss in efficiency through blade-tip leakage losses, can also lead to other major problems. This is due above all to the fact that very high circumferential speeds occur in the area of the outer diameter of the rotor, the rotor as well as the housing are subjected to strong vibrations and the rotor undergoes additional axial shifts due to the high pressures acting. These operating conditions have hitherto prevented the construction of a virtually or ideally completely tight seal.
Seals hitherto used, such as, e.g., labyrinth seals, are not seals in the real sense, but only devices for reducing the leakage water flow, wherein an undesirable leakage water flow occurs in any case, however. Others, such as, e.g., known ice ring seals according to EP 1 098 088 A, were in turn very complex and unreliable in a possible operation, since susceptible to breakdown.
US 5 052 694 A shows a sealing body that is supported on a hydrostatic bearing in one direction. This sealing body is rail-guided in a second direction by means I
of sealing elements. A throttle device is provided in the feed line to the hydrostatic bearing, with which throttle device fluctuations in the gap widths of the hydrostatic bearing are absorbed. Also AT 298 365 B shows a seal with a sealing body that is supported on a hydraulic bearing in one direction and that is again forcedly guided in the other direction by a piston and a conduit. The necessary pressing force for the sealing body is thereby effected by the hydraulic piston.The supply pressure for the hydraulic sealing and for the piston is set by separate, independent throttle devices in their respective supply line.
Likewise seals between two parts moved relative to one another by means of a hydrostatic bearing are known from the prior art. A sealing element is thereby supported on a hydrostatic bearing in one direction between the two parts moved relative to one another. In a second direction the sealing element is supported on a hydraulic bearing. The hydraulic bearing thereby comprises respectively two sealing bodies, such as, e.g., 0 rings, between which a pressure medium is fed.
Examples of seals of this type can be taken from US 4 118 040 A or GB 839 880.
In the axial rotor side space of a turbine, i.e., between hub and housing, a sometimes very high pressure would form through the leakage water flow and the seal between shaft and rotor side space (e.g., by means of gland seals), which high pressure essentially corresponds to the headwater pressure, and which would tend to displace the rotor in the axial direction, whereby large axial loads of the bearing of the rotor and relatively large axial shifts of the rotor would occur, which is avoided through a pressure reduction with the aid of a known leakage water line in the tailwater. A leakage water flow develops in both rotor side spaces, that is, between housing and ring or hub, whereby a certain proportion of the medium does not flow through the rotor, thus resulting in a drop in efficiency and a loss of power. Furthermore, a water disk rotating very quickly forms in both rotor side spaces, which disk of water counteracts the rotation of the shaft due to the developing friction and thus develops a braking effect, which in turn further reduces efficiency. For these reasons it is desirable to provide a virtually completely tight seal between rotor and housing which eliminates the leakage water flow.
r e CA 02668482 2009-05-04 To this end it was proposed in WO 02/23038 Al and WO 2004/018870 Al to provide a floating sealing ring in the peripheral region of the rotor, which sealing ring is arranged on two hydrostatic bearings in a free floating manner with respect to the housing and the rotor, and thus prevents a penetration by leakage water into the rotor side space through the seal water exiting at the hydrostatic bearings, whereby the above-referenced problems can be solved. The sealing ring thereby has respectively at least one connection bore that connects the two hydrostatic bearings hydraulically to one another. The seal water is now fed via at least one supply line to a bearing and reaches the second bearing via the connection bore(s). It was always desirable thereby for the seal water to reach one bearing from the other as far as possible free of losses in order to be able to secure the pressures or pressure distributions in the bearings necessary for the floating of the sealing ring. To this end the connection bores were always embodied as large as possible and/or many connection bores were provided distributed over the circumference in order to keep as low as possible the unavoidable natural flow losses, which as is known depend essentially on the geometry of the flow rate through the connection bores.
The disadvantage of the arrangements of sealing rings in the turbine disclosed in these two applications lies on the one hand in that the sealing ring in practice has unsatisfactory functionality, in particular with a sealing ring according to Fig. 3 of WO 02/23038 Al. Namely if the sealing ring bears radially firmly against the housing, it is hard to lift the sealing ring in the radial direction, since the entire seal water that is pressed in via the supply line is namely conducted via the connection bore to the second bearing and there leads to a strong lift in the axial direction. The sealing ring would therefore rub against the housing, which leads to damage and as a result can lead to destruction. However, in the event that the ring is installed with a certain radial play, although it would center itself during operation and lift radially and axially, it would not adopt a preferred position due to the lack of equilibrium of forces, it would be unstable and it would be just as impossible to regulate radially. Namely, if an attempt is made to change the radial position by changing the volume flow, only the axial position would change, since a changed volume flow in turn would be transferred directly to the axial bearing via the connection bore. A sealing ring of this type would therefore not be very practicable in practice. On the other hand, with the other variants of the sealing ring seal water is required with a pressure that is considerably higher than the headwater pressure itself in order to lift the sealing ring and to keep it in a stable condition. In the operation of the sealing ring an additional unit of corresponding power is consequently necessary to increase the pressure of the seal water, such as, e.g., a seal-water pump, which naturally increases the expenditure and again in part eliminates the gain in efficiency through the sealing ring.
The object of the present invention is now to eliminate the disadvantages still existing with the known sealing ring arrangements, in particular to disclose a compact arrangement, with which additional units for increasing the seal water pressure can be omitted or much smaller additional units are required and a stable operation of the sealing ring is nevertheless rendered possible.
This object is attained according to the invention in that a throttle device is provided between the first hydraulic bearing and the second hydraulic bearing in the hydraulic connection, with which throttle device a pressure loss, based on the pressure difference to be sealed bearing against the sealing element, of at least 1 %, preferably at least 5%, preferentially at least 10%, can be set between the first hydraulic bearing and the second hydraulic bearing. A small compact construction can be achieved through an "internal" restriction of this type between the first and second hydraulic bearing. Furthermore, a broader feed groove results through the lower feed pressure, which advantageously enlarges the axial movement clearance of the sealing element. However, in particular the arrangement according to the invention can be operated with a greatly reduced feed pressure of the seal water, which means that ideally no additional unit is necessary to increase the pressure. The seal water can be taken directly from the headwater, for example, and set to the desired seal water pressure via pressure regulating devices. This results in a much lower energy loss than with the known generic sealing rings of the prior art. Furthermore, the arrangement according to the invention also renders possible a stable operation of the sealing element, since through the internal restriction between the two hydraulic bearings it is also possible to control the radial bearing. Preferably the hydraulic bearings are embodied as hydrostatic bearings.
The throttle device is embodied particularly simply as a hydraulic connection, pipe insert in the hydraulic connection, as a collar or perforated plate or as a sudden expansion of the cross section of the hydraulic connection. With a throttle device of this type the pressure loss occurring at the throttle device can be precisely determined and adjusted, whereby the behavior of the sealing element can be predetermined.
The sealing element is preferably arranged in a rotationally fixed manner by means of an anti-rotation element with respect to one of the two parts of the hydraulic turbomachine, wherein the anti-rotation element advantageously exerts essentially only a circumferential force on the sealing element.
In particular with sealing rings with a large diameter it is advantageous to embody it divided into at least two parts, which facilitates the production and the transport of the sealing ring.
It can thereby be provided for an automatic peripheral length compensation that the sealing ring is embodied in a slotted manner at a number of dividing points, and at the remaining dividing points the parts of the sealing ring are connected to one another rigidly, wherein a gap is provided at the slotted dividing point in the circumferential direction. In order to prevent seal water from flowing through the developing gap in the axial and/or radial direction through the sealing element, a means for preventing a flow through can be provided. In order to prevent a deformation of the slotted sealing ring, a means for preventing a deformation of the sealing ring can be provided at the slotted dividing point, which means does not obstruct the mobility of the sealing ring in the circumferential direction.
Likewise it can be advantageous for a uniform distribution of forces over the circumference of the sealing ring to embody at least one bearing recess adjoining a dividing point with an at least partially enlarged surface.
When the sealing element in operation is held in an equilibrium of forces and is torque free, the sealing element can be constructed in a torsionally flexible manner, which advantageously reduces the dimensions as well as the weight of a sealing element of this type.
The present invention is described below based on the exemplary non-limiting Figs. 1 through 5 showing special embodiment variants. They show Fig. 1 a cross section of a typical Francis turbine Fig. 2 a detailed view of the sealing area between rotor and housing with a sealing element according to the invention Fig. 3 a perspective representation of a sealing element according to the invention Fig. 4 a further detail of a sealing element according to the invention and Fig. 5 a further advantageous embodiment of the sealing element.
Before the actual description of the invention, some terms are defined and explained in more detail below:
A bearing recess or bearing surface can have an annular or cylindrical shape, can have any widths and depths or heights and can be continuous in the circumferential direction or also interrupted in sections at one or more places.
Naturally, a bearing recess can have any cross-sectional shape, e.g., also a triangular flute and does not necessarily need to be embodied as a rectangular groove, as in the concretely described exemplary embodiment. A hydraulic bearing, such as, e.g., a hydrostatic bearing, always comprises bearing surfaces facing towards one another. A feeder for bearing medium such as, e.g., seal water opens into at least one of the bearing surfaces of a hydraulic bearing.
At least one bearing recess, such as, e.g., a groove, flute or the like is preferably arranged in at least one of the bearing surfaces. However, a hydraulic bearing can also have two or more bearing recesses. When now several bearing recesses are arranged over the circumference, because one bearing recess, e.g., as described above is interrupted in some sections, to simplify matters in the following description only one hydraulic bearing and one bearing recess is discussed, however.
A hydraulic connection, such as, e.g., a connection bore, means for the purposes of this application at least one cavity with at least two open ends, wherein this cavity can be flowed through by a medium in any desired manner from one end to the other end.
When a supply line is mentioned it should be noted that a number of such identical or similar supply lines, that is a row of supply lines, can be arranged in the circumferential direction. The same naturally also applies to a hydraulic connection. In order to prevent the description from becoming too complicated, however, usually only one supply line or one hydraulic connection is mentioned, wherein this thus as necessary naturally also covers a row of supply lines or hydraulic connections distributed over the circumference.
To simplify matters the seal according to the invention is described only based on a turbine, in the specific case a Francis turbine, wherein however this seal of course can also be used equivalently with all other hydraulic turbomachines with parts moveable with respect to one another, such as a rotor which runs in a machine housing, such as, e.g., with pumps or pump turbines.
Fig. 1 now shows a turbine 1, in this case a Francis turbine, with a rotor 2, which runs in a turbine housing 12. The rotor 2 has a number of turbine blades 3 that are delimited by the ring 10 and the hub 11. The rotor 2 is attached in a rotationally fixed manner with respect to the shaft 8 by means of a hub cap 9 and possibly by means of further mounting means, such as, e.g., studs or screws, at one end of the shaft 8. The shaft 8 is hinge-mounted by means of shaft bearings (not shown) and drives in a known manner, for example, a generator (not shown either) for generating electric energy, which generator is preferably arranged at the other end of the shaft 8.
The inflow of the liquid medium, usually water, from a headwater, such as, e.g., a water reservoir lying higher, occurs in most cases via a volute casing (not shown here) that is sufficiently known. A diffuser 4, comprising a number of guide vanes 5, which in this example can be rotated by means of an adjusting device 6 are between the volute casing and the rotor 2. The adjustable guide vanes 5 serve to regulate the power of the turbine 1 by changing both the volume flow through the turbines 1 as well as the rotor inlet swirl. In addition, stay vanes could also be arranged in a known manner between the volute casing and the guide vanes 5.
The outflow of the water takes place, as shown in Fig. 1, via a suction pipe directly adjoining the turbine 1, which suction pipe opens into a tailwater (not shown). This results in a main water flow F labeled by the arrow from the volute casing via the diffuser 4 and the rotor 2 to the suction pipe 13.
In addition to the main water flow F, with conventional seals 14, such as, e.g., a labyrinth seal shown in Fig. 1, a leakage water flow also forms in a known manner through the rotor side spaces between the turbine housing 12 and ring or hub 11. Losses occur due to these leakage water flows, on the one hand because the energy of a certain quantity of water cannot be processed in the rotor 2 and on the other hand because considerable friction forces can be generated by the rotating rotor 2 in the rotor side spaces filled with water, which friction forces brake the rotor 2. The leakage water of the radial rotor side space (i.e., between turbine housing 12 and hub 11) is often guided away, e.g. by means of a leakage water line 7 via a throttle device and guided into the suction pipe 13. In addition, as indicated in Fig. 1, pressure relief holes are often provided in the hub 11, via which the radial rotor side space is connected to the main water flow F. However, these measures of course cannot prevent the occurrence of a leakage water flow. Through the sealing according to the invention, as described below, these leakage water flows are now eliminated, so that the entire water flowing in flows through the rotor 2 and the flow energy thereof, essentially without blade-tip leakage losses, can be fully utilized.
Furthermore, the friction in the rotor side spaces is reduced or even minimized, since with a seal of this type rotating water disks no longer form in the rotor side spaces, instead these spaces, which the exception of the bearing water, are filled with air. Furthermore, the axial thrust acting on the shaft 8 and on the shaft bearing is thereby also greatly reduced without having to guide away the leakage flow via a leakage line 7.
Fig. 2 now shows a detailed view of an exemplary inventive seal of a peripheral part of a rotor 2 of a turbine 1 between the turbine housing 12 and hub 11 by means of a sealing element embodied as a sealing ring 20, which is arranged supported in a floating manner on two hydraulic bearings, in this case two hydrostatic bearings 22, 23. Supported in a floating thereby means that the sealing element is essentially freely moveable in all directions and the movements of the sealing element are not counteracted by any other forces, such as, e.g., holding forces, friction forces, forces caused by further sealing bodies between the sealing element and parts of the turbine 1 or forces of further counteracting hydraulic bearings, in the same order of magnitude as the hydraulic compressive forces acting on the sealing element. To this end the turbine housing 12 has a shoulder 35 on which a radial bearing surface 26 is arranged. Likewise an axial bearing surface 27 is arranged on the rotor 2 on the hub 11. These bearing surfaces 26, 27 can be separate components that are subsequently applied at the necessary location, e.g., by welding, screwing, adhering, etc., or naturally can also be directly worked into the corresponding component, e.g., the hub 11, e.g., a face ground section on the hub 11 or the shoulder 35 coated with a wear-resistant coating. A radial 34 or axial bearing surface 33 on the sealing ring 20 is respectively assigned to the radial bearing surface 26 on the turbine housing 12 or the axial bearing surface 27 on the hub 11. The two associated bearing surfaces 26, 34 and 27, 33 respectively form a part of a radial and axial hydrostatic bearing 22, 23.
The orientations "axial" or "radial" thereby always refer to the directions of action of the hydraulic bearings based on the rotation axis of the rotor 2 of the turbine 1 and are mainly inserted in the description to make it easier to distinguish between the two hydraulic bearings, such as the two hydrostatic bearings 22, 23.
In the embodiment according to Fig. 2 a bearing medium, such as e.g., water (seal water), is fed into the radial hydrostatic bearing 22 via the turbine housing 12 by means of a supply line 21. The supply line 21 is formed here from bores in the turbine housing 12, which if necessary is connected via further lines indirectly or directly to a supply source (not shown) such as a pump and/or the headwater, possibly via auxiliary devices, such as filters, cyclone separators, throttle devices, flow control valves, etc. With a throttle device or a flow control valve in the supply line 21 the fed quantity of seal water Q or the seal water pressure p1 can be adjusted very easily.
Naturally a plurality of supply lines 21 can be arranged distributed, preferably symmetrically, over the circumference, wherein an arrangement favorable for a sufficient and uniform supply, e.g., three supply lines 21, which are arranged respectively offset at an angle of 1200, can be provided. Naturally any other useful arrangement is also conceivable.
The radial bearing 22 now has a bearing recess in the form of a groove 24, in which the supply line 21 opens and which, as known (ideally) forms an area of constant pressure. This groove 24 of the radial hydrostatic bearing 22 is connected via a hydraulic connection, here a connection bore 30, to a groove of the axial bearing 23 arranged in the sealing ring 20. The two grooves 24, are likewise part of the associated hydrostatic bearings 22, 23.
It should also be noted thereby in particular that the bearing recesses, here grooves 24, 25, can also equally be arranged in the axial or radial bearing surface 26, 27 of the turbine housing 12 or the rotor 2, as here in the hub 11. It would likewise be possible to provide bearing recesses in the sealing body 20 as well as in the turbine housing 12 or at any desired location of the rotor 2 in the area of the associated hydraulic bearing. It is likewise conceivable for a hydraulic bearing to have no bearing recess at all.
The bearing recesses, here the grooves 24, 25, naturally do not have to be continuously connected over the circumference, instead it is also possible to arrange several pockets distributed over the circumference and forming the grooves 24, 25, as shown in Fig. 3 and described in detail below.
Thus both hydrostatic bearings 22, 23 are supplied with bearing medium, the seal water, from a single supply line 21 (or from a number of supply lines distributed over the circumference). The seal water is thereby fed in the present example in the radial bearing 22 and flows via the connection bores 30 into the axial bearing 23. Of course the supply line 21 could also open into the axial bearing 23 however.
A throttle device 37 is now provided in the connection bore 30, which throttle device causes a predetermined pressure drop between the first hydrostatic bearing 22 and the second hydrostatic bearing 23. According to basic laws of hydrodynamics the restriction effect of a throttle device 27 is essentially proportional to a loss coefficient i;, to the density p of the flowing medium and to the square of the flow velocity through the throttle device 37. The throttle device 37 can be embodied in different ways, e.g., as a collar, as a perforated plate, as a baffle through the design of the geometry, such as a corresponding selection of the diameter or length, of the connection bore 30, or, as in the example according to Fig. 2, the known Carnot shock loss can be utilized with a sudden expansion of the cross section of a flow channel. As is adequately known, the Carnot shock loss pv is thereby essentially proportional to the square of the flow velocity u through the flow channel and the square of the ratio of the cross-sectional surfaces after A2 and before Al of the expansion of the flow channel (loss coefficient Q and can be easily calculated with the known density of the flow medium p and adjusted as desired through the selection of the geometry of the flow channel, such as e.g., the connection bore 30:
In the current example according to Fig. 2 an area 31 with large cross section is provided in the area of the groove 24 in the entry area of the connection bore 30, from which area 31 the connection bore 30 starts, wherein the transition from entry area 31 to connection bore 30 can be rounded in order to reduce possible wear in this area and thus to keep the selected restriction effect constant over longer operating times of the sealing element 20. The opposite end 32 of the connection bore 30 has a sudden cross-sectional expansion at which the Carnot shock loss occurs. The speed through the connection bore 30 can thereby be adjusted with a certain pressure of the seal water p1 or at a certain volume flow Q (which can be controllable e.g., through a flow control valve in the supply line 21) through the cross-sectional surface of the connection bore 30 and/or preferably through the number of connection bores 30 arranged distributed over the circumference.
The throttle device 37 is now selected or designed such that a pressure loss based on the pressure difference po-pu to be sealed bearing against the sealing ring 20, of at least 1%, preferably at least 5%, preferentially at least 10%
occurs between the first hydrostatic bearing 22 and the second hydrostatic bearing 23.
The natural unavoidable pressure loss caused through a flow through a bore as far as possible free of loss, as desired in previous sealing rings, is typically lower by approx. two to three decimal powers, thus is negligible with respect to the adjustable "inner restriction." Thus the desired pressure loss according to the invention through the throttle device 37 is considerably higher than the unavoidable natural restricting effect (essentially through pipe friction;
possible deflections in the connection bore and as a result of a cross-sectional expansion).
The constricted center part of the connection bore 30 can also be achieved through a tubular insert 38, which is arranged in a larger through bore, as shown in Fig. 5. An arrangement of this type is in particular advantageous when the sealing ring 20 is produced from a soft material, e.g., of plastic, and the pipe insert 38 is then produced from a hard, rugged material in order to resist the flow at high flow velocity through the narrow connection bores 30 over the long term and thus to keep the adjusted intemal restriction essentially constant for as long as possible.
In order to be able to describe the function of the sealing ring 20, in Fig. 2 in addition the pressure distributions resulting in the axial and radial bearing 23, 22 are shown diagrammatically. The bearing medium is fed as described above via the supply line 21 with a volume flow Q or with a pressure p1 into the radial bearing 22. The volume flow Q of the bearing medium is divided into three flows in the radial bearing 22. One flow flows downwards and finally opens into the axial rotor side space with pressure pu. A second flow flows upwards and opens into the bearing space 36 with the pressure po prevailing at the rotor entry, which pressure naturally also acts on the sealing ring 20. A third flow, largest by far, flows via the connection bore 30 into the axial bearing 23. The division of the volume flow Q thereby results naturally through the radial bearing gap which is essentially adjusted through the pressure p1 prevailing in the hydrostatic bearing 22 and the pressure po at the rotor entry.
Due to basic known laws of fluid mechanics and depending on the geometry of the radial bearing 22 in the radial bearing 22, the volume flow Q causes the pressure distribution shown with a maximum pressure p1 in the groove 24 into which the supply line 21 opens. This pressure distribution causes the sealing ring 20 to lift in the radial direction. Radial lifting means in this context on the one hand that the sealing ring 20 correspondingly centers itself, since the pressure acts over the entire inner circumference, and on the other hand also expands.
The headwater pressure po but also according to the theory of elasticity the elastic restoring forces of the sealing ring 20 counteract this expansion. The maximum pressure p1 must therefore be large enough to be able to cause the lifting of the sealing ring 20 on the desired bearing gap, e.g., typically 50 -pm.
In the axial hydrostatic bearing 23 a pressure p2 is established, which according to the adjustable pressure loss between the two bearings 22, 23 is smaller than the pressure p1 in the radial hydrostatic bearing 22. In turn a pressure distribution in the axial bearing 23 can be adjusted via the width and position of the groove 25, in order to also lift the sealing ring 20 in the axial direction.
Through the provision of a restriction in the hydraulic connection, here connection bore 30, between the radial hydrostatic bearing 22 and the axial hydrostatic bearing 23 the pressure p1 in the radial bearing 22 can also be influenced via the pressure in the supply line 21 or the fed volume flow Q, whereby the radial bearing 22 can also be controlled.
Since the sealing ring 20 in real applications can have large diameters up to several meters, but small cross sections and light materials are desirable, the sealing ring 20 is very torsionally flexible, that is, very sensitive to moments, such as tilting moments or pulling moments. It is therefore important to keep the sealing ring 20 torque free during operation, since otherwise a ring very resistant to deformation would be required. The adjustment of a moment equilibrium is thereby possible through the selection of the geometry of the sealing ring 20, the hydrostatic bearings 22, 23 and the pressure distributions and the pressure level in the bearings 22, 23. As is easily realized, this can be achieved in that the grooves 24, 25 are arranged such that the balance of torques due to the pressure distributions acting on the sealing ring 20 is equalized around the surface cross section of the sealing ring cross section. In order to achieve this, in addition to the entire geometry of the sealing ring 20, such as the position and dimensions of the grooves 24, 25, the bearing gap widths, the outer dimensions etc., among other things the recess 29 is also used.
The sealing ring 20 consequently floats practically without friction and in a stable manner on two hydraulic bearings and is kept in balance only by the hydraulic compressive forces acting on the sealing ring 20. In balance thereby means that the sealing ring 20 is in an equilibrium of forces as well as free of torque.
If the sealing ring 20 is not arranged in a torsionally fixed manner, it additionally rotates at a lower rotational speed together with the rotor 2. This thereby also results in a dynamic gain in stability, since the limiting circumferential velocity or the flutter limit is thus raised. Furthermore, the friction losses are also lower due to the lower relative velocities.
With a design of a sealing element according to the invention, this can consequently be held in an equilibrium of forces and absence of torque can be ensured at the sealing element, whereby a tilting or pulling that is dangerous for the safe operation of the torsionally flexible sealing element can be prevented.
In the case of a rotationally fixed sealing ring 20, the rotational fixing can be carried out virtually arbitrarily, with the proviso that the axial and radial mobility of the sealing ring 20 in operation may not be hindered or may be hindered only negligibly. For example, a simple stationary limit stop can be provided on the turbine housing 12, against which a correspondingly shaped limit stop on the sealing ring 20 bears in order to prevent the rotation of the sealing ring 20.
Since the sealing ring 20 as described above floats in a virtually lossiess and frictionless manner on the pressure cushions forming in the hydraulic bearings, only very low circumferential forces develop on the sealing ring 20, typically in the range of several hundred N or less. With bearing forces in the hydraulic bearings 22, 23 of typically several hundred kN and greater (that is, greater by a factor of - 1000), these circumferential forces (and friction forces or retaining forces possibly developing therefrom at the anti-rotation element) can be disregarded, however. An anti-rotation element of this type consequently would not influence the axial and radial mobility of the sealing ring.
Through the high stability of a hydraulic bearing and the free mobility of the sealing ring 20, the sealing ring 20 is able to equa(ize oscillations of the rotor 2 and/or of the turbine housing 12, as well as axial displacements of the rotor 2, without losing the sealing action and without coming into contact with the rotor 2 and/or the turbine housing 12. The sealing ring 20 sustains virtually no wear thereby, since it floats in a virtually frictionless manner on liquid films, whereby the service life of a sealing ring 20 of this type is very high. Because the sealing ring 20 can be preferably constructed as a very slim, tight ring that has hardly any inertia forces, this action is further intensified.
The sealing ring 20 can be constructed to be very small in comparison to the dimensions of the rotor 2, edge lengths of several centimeters, e.g., 5 cm or cm, with outer diameters of several meters are completely adequate, and it can be manufactured of any desired material, such as steel, bearing bronze, plastic (e.g., PE). Furthermore, the bearing surfaces 26, 34 and 27, 33 can also be covered with a suitable layer, such as Teflon, bearing bronze, etc., in order to improve the properties of the seal still further. Typically the sealing ring 20 is manufactured from a softer material, such as, e.g., Teflon, bronze, etc., than the housing 12 or the rotor 2 of the hydraulic machine. As a rule it thus becomes lighter on the one hand and on the other hand in an extreme case the sealing ring 20 and not the rotor 2 or the housing 12 is damaged or even destroyed.
Fig. 3 shows a perspective representation of a sealing ring 20 according to the invention. A plurality of grooves 25 are inserted on a front face of the sealing ring 20, which grooves are part of the axial hydrostatic bearing 23, wherein at least one connection bore 30 opens in each of these grooves 25. A row of grooves 24 is inserted on the inside of the sealing ring 20, which grooves are part of the radial hydrostatic bearing 22 and wherein in turn at least one connection bore opens in each of these grooves 24. The number, size and distribution of the grooves 24, 25 and the connection bores 30 can thereby be selected accordingly to achieve an equilibrium of forces and moment equilibrium and to adjust the desired pressure distributions and the bearing gaps in the assigned hydrostatic bearings 22, 23.
However, the sealing ring 20, as shown in Figs. 3 and 4, can also be embodied divided in the circumferential direction at one or more dividing points 41, 44 which among other things facilitates the production or the transport of the sealing ring 20. In the exemplary embodiment shown, the sealing ring 20 is divided at two dividing points 41, 44 wherein the two halves are rigidly connected to one another at a first dividing point 41 via a connection means, such as, e.g., via dowel pins and a screw. At a second division point 44 the sealing ring 20 remains slotted in the assembled condition so that a gap 50 is formed in the circumferential direction, as shown in Fig. 4. This gap 50 is used essentially to render possible an automatic peripheral length compensation, which compensates for possible wear of the sealing ring 20 at the bearing surface 34, e.g., by fine solids (such as, e.g., sand) in the fed seal water, which have an abrasive action. The automatic peripheral length compensation results in that the bearing gaps (that is, the thickness of the liquid films in the bearings) are adjusted through the supply of a specific feed quantity with a specific pressure.
The slotted sealing ring 20 is compressed in the radial direction through the acting compressive forces that press the sealing ring 20 against the two hydrostatic bearings 22, 23 (the gap 50 at the dividing point 44 thus becomes smaller) so that the desired bearing gap is reestablished in the hydrostatic bearing. When the entire gap 50 is "used up," the two parts of the sealing ring thus strike against one another and a further wear can thus no longer be compensated, the sealing ring 20 can be replaced.
Since a slotted sealing ring 20 of this type, however, would naturally deform very considerably due to the high pressures acting thereon, in particular would bend and twist, corresponding measures have to be taken in order to also maintain the shape of the sealing ring 20 during operation. Moreover, seal water could flow through a slot or gap 50 of this type, which of course would negatively affect the function of the sealing ring 20. To this end on the one hand a shoulder 45 is now provided at the slotted dividing point 44 at a front face of the sealing ring arranged in the circumferential direction, against which shoulder a recess 47 on the opposite front face of the sealing ring 20 arranged in the circumferential direction bears. By pressing the recess 47 against the shoulder 45 a seal is realized which prevents seal water from flowing through the slot in the radial direction. Furthermore, the sealing ring 20 thereby cannot bend in the radial direction. On the other hand, a plate can be arranged on the outer front face of the sealing ring 20 facing away from a hydrostatic bearing, which plate extends over the gap 50. A seal would thus be created again which prevents an axial flow through the slot by seal water. Furthermore, an axial bending and a twisting of the sealing ring 20 would be prevented thereby. The plate could thereby be attached to the sealing ring 20 by means of suitable connection means, such as, e.g., by means of dowel pins and screws, wherein these connection means naturally may be arranged only on one half of the sealing ring 20 in order not to impede the mobility of the two sealing ring halves in the circumferential direction (which is important for the peripheral length compensation). Of course, arbitrary other suitable arrangements can also be provided in order to realize a corresponding seal and to prevent a deformation of the sealing ring 20. For example, instead of the plate, analogously to the radial direction as described above, the sealing ring 20of course could also have a shoulder on one half in the axial direction, which shoulder interacts with a corresponding recess on the other half.
Through the slot 44 the radial as well as the axial groove 24, 25 can be interrupted or the symmetrical arrangement of the groves 24, 25 around the circumference can be disturbed. Lower bearing forces could thus form in the area of this dividing point, since in this area the surfaces of the grooves 24, 25 on which the bearing pressure p1, p4 acts, would be smaller, which would lead to an uneven pressure load on the sealing ring 20. This could deform the deformable, elastic, slim sealing ring 20, which can lead to problems in practice with the low bearing gaps in the hundred pm range, such as e.g., semifluid friction conditions and, in the worst case, to contact between the sealing ring 20 and rotor 2 or turbine housing 12. In order to prevent this, the grooves 24, 25 adjoining the gap 50 can be embodied larger, or in this area another groove division could be provided in the circumferential direction and thus a different groove surface again. For example, the grooves 25 adjoining the gap 50 on the front face of the sealing ring 20 in the end facing towards the gap 50 could have an area 46 with a larger groove surface. Likewise the grooves 24 adjoining the gap 50 on the inside of the sealing ring 20 in the end facing towards the gap 50 can have an enlarged groove surface 42. With these enlarged groove surfaces 42, 46 a corresponding equalization and a reestablishment or retention of the of forces and moment equilibrium can be carried out.
If the sealing ring 20 is embodied to rotate in the same direction, it is advantageous to balance it, in particular with a slotted or divided embodiment.
The sealing ring 20 can, of course, have an arbitrary cross section, such as, e.g., an L-shaped cross section, wherein for reasons of manufacturing engineering a square or rectangular shape is preferred.
Until now, flat bearing surfaces 26, 34 and 27, 33 have always been assumed.
However, it is naturally also conceivable to embody the bearing surfaces 26, and 27, 33 not flat, but, e.g., concave or convex or stepped, wherein the fundamental principle of the seal according to the invention is not changed.
With bearing surfaces of this type 26, 34 and 27, 33 that are not flat only the pressure distributions would change somewhat, but this is clearly evident to one skilled in the art. It is likewise conceivable that the bearing surfaces 26, 34 and 27, based on the rotation axis of the turbine 1 are not axially parallel and axially perpendicular, but are arranged at an angle thereto.
To improve the bearing effect and the stability, one or more of the bearing surfaces 26, 34 and 27, 33 could also additionally be provided with adequately known hydrodynamic oil pockets to form an additional hydrodynamic bearing.
A seal according to the invention with a sealing ring 20 can naturally be provided at any suitable point and is not limited to the exemplary embodiments according to Fig. 2. For example the sealing ring 20 could also be arranged between the front face of the rotor 2 or hub 11 and the turbine housing 12. Likewise it is conceivable to provide a seal of this type at a suitable point between the ring 10 and the machine housing 12.
Likewise it would be conceivable with a rotationally fixed sealing ring 20, as known per se from the prior art, in the bearing surface 26, which is not subject to any relative motion regarding the assigned bearing surface 34 on the sealing ring 20, on both sides laterally next to the supply line 21 or at the side of a groove 24, to provide respectively one auxiliary sealing body between the seating ring 20 and the facing bearing surface 26 on the housing 12. With an auxiliary sealing body of this type, seal water would be prevented from exiting at this hydraulic bearing, whereby the losses through the seal according to the invention can be reduced still further. However, above all therewith excessive wear of the hydraulic bearing by exiting contaminated seal water could be prevented or considerably reduced. Of course, the auxiliary sealing bodies would thereby have to be used such that they develop the lowest possible action of force with a still adequate sealing action, since the sealing ring 20 is to remain supported in a floating manner. Since the sealing ring 20 is supported in a floating manner on the pressure cushions in a stable manner, this can be easily achieved via the level of the sealing gap to be adjusted. Through the possible low force on the auxiliary bodies, typically lower by a factor of 1000 or more than the hydraulic bearing forces, this in turn can be referred to as a floating bearing and it can be ensured that the axial and radial mobility of the sealing ring 20 are influenced only slightly and insignificantly.
The seal described above represents a largely tight seal. The entire water quantity flowing in flows through the rotor and can be converted into rotational energy. The leakage water losses are thereby reduced exclusively to the exiting bearing medium, thus based on the low gap heights in the hydraulic bearings, e.g., typically in the 10 -60 pm range with hydrostatic bearings 22, 23, are very low and can be recovered in part by the introduction of the leakage water into the main water flow F again.
In all phases of operation it should be avoided as far as possible that the sealing ring 20 comes into contact with the rotor 2 or the housing 12 or semifluid friction conditions form in the hydraulic bearings 22, 23, since then the sealing ring can be very easily damaged or even destroyed. At the start of the turbine 1 the sealing ring 20 should therefore already be lifted, i.e., the desired bearing gaps should already have been achieved. This can be easily achieved in that the supply of the hydraulic bearings 22, 23 is switched on first and only then the turbine 1 is switched on. In the case of a breakdown of the supply of the hydrostatic bearings 22, 23, for example, an emergency supply such as am air chamber, could be provided in order to avoid damage to the sealing ring 20 of the hydraulic machine, which would entail complex repair work.
Upon first putting into service the sealing ring 20, however, it can be desirable to adjust a controlled semifluid friction condition in the hydraulic bearings 22, 23, so that a bearing image can be ground in the bearing surfaces 26, 34 and 27, 33, whereby certain manufacturing inaccuracies can be compensated. However, since the bearing gaps are in the hundred pm range or lower, corresponding care is thereby naturally necessary.
In the description above to simplify matters, water is described as the bearing medium. Of course, the bearing medium, above all in the case of pumps, naturally can also be any other suitable medium, such as, e.g., oil.
The disadvantage of the arrangements of sealing rings in the turbine disclosed in these two applications lies on the one hand in that the sealing ring in practice has unsatisfactory functionality, in particular with a sealing ring according to Fig. 3 of WO 02/23038 Al. Namely if the sealing ring bears radially firmly against the housing, it is hard to lift the sealing ring in the radial direction, since the entire seal water that is pressed in via the supply line is namely conducted via the connection bore to the second bearing and there leads to a strong lift in the axial direction. The sealing ring would therefore rub against the housing, which leads to damage and as a result can lead to destruction. However, in the event that the ring is installed with a certain radial play, although it would center itself during operation and lift radially and axially, it would not adopt a preferred position due to the lack of equilibrium of forces, it would be unstable and it would be just as impossible to regulate radially. Namely, if an attempt is made to change the radial position by changing the volume flow, only the axial position would change, since a changed volume flow in turn would be transferred directly to the axial bearing via the connection bore. A sealing ring of this type would therefore not be very practicable in practice. On the other hand, with the other variants of the sealing ring seal water is required with a pressure that is considerably higher than the headwater pressure itself in order to lift the sealing ring and to keep it in a stable condition. In the operation of the sealing ring an additional unit of corresponding power is consequently necessary to increase the pressure of the seal water, such as, e.g., a seal-water pump, which naturally increases the expenditure and again in part eliminates the gain in efficiency through the sealing ring.
The object of the present invention is now to eliminate the disadvantages still existing with the known sealing ring arrangements, in particular to disclose a compact arrangement, with which additional units for increasing the seal water pressure can be omitted or much smaller additional units are required and a stable operation of the sealing ring is nevertheless rendered possible.
This object is attained according to the invention in that a throttle device is provided between the first hydraulic bearing and the second hydraulic bearing in the hydraulic connection, with which throttle device a pressure loss, based on the pressure difference to be sealed bearing against the sealing element, of at least 1 %, preferably at least 5%, preferentially at least 10%, can be set between the first hydraulic bearing and the second hydraulic bearing. A small compact construction can be achieved through an "internal" restriction of this type between the first and second hydraulic bearing. Furthermore, a broader feed groove results through the lower feed pressure, which advantageously enlarges the axial movement clearance of the sealing element. However, in particular the arrangement according to the invention can be operated with a greatly reduced feed pressure of the seal water, which means that ideally no additional unit is necessary to increase the pressure. The seal water can be taken directly from the headwater, for example, and set to the desired seal water pressure via pressure regulating devices. This results in a much lower energy loss than with the known generic sealing rings of the prior art. Furthermore, the arrangement according to the invention also renders possible a stable operation of the sealing element, since through the internal restriction between the two hydraulic bearings it is also possible to control the radial bearing. Preferably the hydraulic bearings are embodied as hydrostatic bearings.
The throttle device is embodied particularly simply as a hydraulic connection, pipe insert in the hydraulic connection, as a collar or perforated plate or as a sudden expansion of the cross section of the hydraulic connection. With a throttle device of this type the pressure loss occurring at the throttle device can be precisely determined and adjusted, whereby the behavior of the sealing element can be predetermined.
The sealing element is preferably arranged in a rotationally fixed manner by means of an anti-rotation element with respect to one of the two parts of the hydraulic turbomachine, wherein the anti-rotation element advantageously exerts essentially only a circumferential force on the sealing element.
In particular with sealing rings with a large diameter it is advantageous to embody it divided into at least two parts, which facilitates the production and the transport of the sealing ring.
It can thereby be provided for an automatic peripheral length compensation that the sealing ring is embodied in a slotted manner at a number of dividing points, and at the remaining dividing points the parts of the sealing ring are connected to one another rigidly, wherein a gap is provided at the slotted dividing point in the circumferential direction. In order to prevent seal water from flowing through the developing gap in the axial and/or radial direction through the sealing element, a means for preventing a flow through can be provided. In order to prevent a deformation of the slotted sealing ring, a means for preventing a deformation of the sealing ring can be provided at the slotted dividing point, which means does not obstruct the mobility of the sealing ring in the circumferential direction.
Likewise it can be advantageous for a uniform distribution of forces over the circumference of the sealing ring to embody at least one bearing recess adjoining a dividing point with an at least partially enlarged surface.
When the sealing element in operation is held in an equilibrium of forces and is torque free, the sealing element can be constructed in a torsionally flexible manner, which advantageously reduces the dimensions as well as the weight of a sealing element of this type.
The present invention is described below based on the exemplary non-limiting Figs. 1 through 5 showing special embodiment variants. They show Fig. 1 a cross section of a typical Francis turbine Fig. 2 a detailed view of the sealing area between rotor and housing with a sealing element according to the invention Fig. 3 a perspective representation of a sealing element according to the invention Fig. 4 a further detail of a sealing element according to the invention and Fig. 5 a further advantageous embodiment of the sealing element.
Before the actual description of the invention, some terms are defined and explained in more detail below:
A bearing recess or bearing surface can have an annular or cylindrical shape, can have any widths and depths or heights and can be continuous in the circumferential direction or also interrupted in sections at one or more places.
Naturally, a bearing recess can have any cross-sectional shape, e.g., also a triangular flute and does not necessarily need to be embodied as a rectangular groove, as in the concretely described exemplary embodiment. A hydraulic bearing, such as, e.g., a hydrostatic bearing, always comprises bearing surfaces facing towards one another. A feeder for bearing medium such as, e.g., seal water opens into at least one of the bearing surfaces of a hydraulic bearing.
At least one bearing recess, such as, e.g., a groove, flute or the like is preferably arranged in at least one of the bearing surfaces. However, a hydraulic bearing can also have two or more bearing recesses. When now several bearing recesses are arranged over the circumference, because one bearing recess, e.g., as described above is interrupted in some sections, to simplify matters in the following description only one hydraulic bearing and one bearing recess is discussed, however.
A hydraulic connection, such as, e.g., a connection bore, means for the purposes of this application at least one cavity with at least two open ends, wherein this cavity can be flowed through by a medium in any desired manner from one end to the other end.
When a supply line is mentioned it should be noted that a number of such identical or similar supply lines, that is a row of supply lines, can be arranged in the circumferential direction. The same naturally also applies to a hydraulic connection. In order to prevent the description from becoming too complicated, however, usually only one supply line or one hydraulic connection is mentioned, wherein this thus as necessary naturally also covers a row of supply lines or hydraulic connections distributed over the circumference.
To simplify matters the seal according to the invention is described only based on a turbine, in the specific case a Francis turbine, wherein however this seal of course can also be used equivalently with all other hydraulic turbomachines with parts moveable with respect to one another, such as a rotor which runs in a machine housing, such as, e.g., with pumps or pump turbines.
Fig. 1 now shows a turbine 1, in this case a Francis turbine, with a rotor 2, which runs in a turbine housing 12. The rotor 2 has a number of turbine blades 3 that are delimited by the ring 10 and the hub 11. The rotor 2 is attached in a rotationally fixed manner with respect to the shaft 8 by means of a hub cap 9 and possibly by means of further mounting means, such as, e.g., studs or screws, at one end of the shaft 8. The shaft 8 is hinge-mounted by means of shaft bearings (not shown) and drives in a known manner, for example, a generator (not shown either) for generating electric energy, which generator is preferably arranged at the other end of the shaft 8.
The inflow of the liquid medium, usually water, from a headwater, such as, e.g., a water reservoir lying higher, occurs in most cases via a volute casing (not shown here) that is sufficiently known. A diffuser 4, comprising a number of guide vanes 5, which in this example can be rotated by means of an adjusting device 6 are between the volute casing and the rotor 2. The adjustable guide vanes 5 serve to regulate the power of the turbine 1 by changing both the volume flow through the turbines 1 as well as the rotor inlet swirl. In addition, stay vanes could also be arranged in a known manner between the volute casing and the guide vanes 5.
The outflow of the water takes place, as shown in Fig. 1, via a suction pipe directly adjoining the turbine 1, which suction pipe opens into a tailwater (not shown). This results in a main water flow F labeled by the arrow from the volute casing via the diffuser 4 and the rotor 2 to the suction pipe 13.
In addition to the main water flow F, with conventional seals 14, such as, e.g., a labyrinth seal shown in Fig. 1, a leakage water flow also forms in a known manner through the rotor side spaces between the turbine housing 12 and ring or hub 11. Losses occur due to these leakage water flows, on the one hand because the energy of a certain quantity of water cannot be processed in the rotor 2 and on the other hand because considerable friction forces can be generated by the rotating rotor 2 in the rotor side spaces filled with water, which friction forces brake the rotor 2. The leakage water of the radial rotor side space (i.e., between turbine housing 12 and hub 11) is often guided away, e.g. by means of a leakage water line 7 via a throttle device and guided into the suction pipe 13. In addition, as indicated in Fig. 1, pressure relief holes are often provided in the hub 11, via which the radial rotor side space is connected to the main water flow F. However, these measures of course cannot prevent the occurrence of a leakage water flow. Through the sealing according to the invention, as described below, these leakage water flows are now eliminated, so that the entire water flowing in flows through the rotor 2 and the flow energy thereof, essentially without blade-tip leakage losses, can be fully utilized.
Furthermore, the friction in the rotor side spaces is reduced or even minimized, since with a seal of this type rotating water disks no longer form in the rotor side spaces, instead these spaces, which the exception of the bearing water, are filled with air. Furthermore, the axial thrust acting on the shaft 8 and on the shaft bearing is thereby also greatly reduced without having to guide away the leakage flow via a leakage line 7.
Fig. 2 now shows a detailed view of an exemplary inventive seal of a peripheral part of a rotor 2 of a turbine 1 between the turbine housing 12 and hub 11 by means of a sealing element embodied as a sealing ring 20, which is arranged supported in a floating manner on two hydraulic bearings, in this case two hydrostatic bearings 22, 23. Supported in a floating thereby means that the sealing element is essentially freely moveable in all directions and the movements of the sealing element are not counteracted by any other forces, such as, e.g., holding forces, friction forces, forces caused by further sealing bodies between the sealing element and parts of the turbine 1 or forces of further counteracting hydraulic bearings, in the same order of magnitude as the hydraulic compressive forces acting on the sealing element. To this end the turbine housing 12 has a shoulder 35 on which a radial bearing surface 26 is arranged. Likewise an axial bearing surface 27 is arranged on the rotor 2 on the hub 11. These bearing surfaces 26, 27 can be separate components that are subsequently applied at the necessary location, e.g., by welding, screwing, adhering, etc., or naturally can also be directly worked into the corresponding component, e.g., the hub 11, e.g., a face ground section on the hub 11 or the shoulder 35 coated with a wear-resistant coating. A radial 34 or axial bearing surface 33 on the sealing ring 20 is respectively assigned to the radial bearing surface 26 on the turbine housing 12 or the axial bearing surface 27 on the hub 11. The two associated bearing surfaces 26, 34 and 27, 33 respectively form a part of a radial and axial hydrostatic bearing 22, 23.
The orientations "axial" or "radial" thereby always refer to the directions of action of the hydraulic bearings based on the rotation axis of the rotor 2 of the turbine 1 and are mainly inserted in the description to make it easier to distinguish between the two hydraulic bearings, such as the two hydrostatic bearings 22, 23.
In the embodiment according to Fig. 2 a bearing medium, such as e.g., water (seal water), is fed into the radial hydrostatic bearing 22 via the turbine housing 12 by means of a supply line 21. The supply line 21 is formed here from bores in the turbine housing 12, which if necessary is connected via further lines indirectly or directly to a supply source (not shown) such as a pump and/or the headwater, possibly via auxiliary devices, such as filters, cyclone separators, throttle devices, flow control valves, etc. With a throttle device or a flow control valve in the supply line 21 the fed quantity of seal water Q or the seal water pressure p1 can be adjusted very easily.
Naturally a plurality of supply lines 21 can be arranged distributed, preferably symmetrically, over the circumference, wherein an arrangement favorable for a sufficient and uniform supply, e.g., three supply lines 21, which are arranged respectively offset at an angle of 1200, can be provided. Naturally any other useful arrangement is also conceivable.
The radial bearing 22 now has a bearing recess in the form of a groove 24, in which the supply line 21 opens and which, as known (ideally) forms an area of constant pressure. This groove 24 of the radial hydrostatic bearing 22 is connected via a hydraulic connection, here a connection bore 30, to a groove of the axial bearing 23 arranged in the sealing ring 20. The two grooves 24, are likewise part of the associated hydrostatic bearings 22, 23.
It should also be noted thereby in particular that the bearing recesses, here grooves 24, 25, can also equally be arranged in the axial or radial bearing surface 26, 27 of the turbine housing 12 or the rotor 2, as here in the hub 11. It would likewise be possible to provide bearing recesses in the sealing body 20 as well as in the turbine housing 12 or at any desired location of the rotor 2 in the area of the associated hydraulic bearing. It is likewise conceivable for a hydraulic bearing to have no bearing recess at all.
The bearing recesses, here the grooves 24, 25, naturally do not have to be continuously connected over the circumference, instead it is also possible to arrange several pockets distributed over the circumference and forming the grooves 24, 25, as shown in Fig. 3 and described in detail below.
Thus both hydrostatic bearings 22, 23 are supplied with bearing medium, the seal water, from a single supply line 21 (or from a number of supply lines distributed over the circumference). The seal water is thereby fed in the present example in the radial bearing 22 and flows via the connection bores 30 into the axial bearing 23. Of course the supply line 21 could also open into the axial bearing 23 however.
A throttle device 37 is now provided in the connection bore 30, which throttle device causes a predetermined pressure drop between the first hydrostatic bearing 22 and the second hydrostatic bearing 23. According to basic laws of hydrodynamics the restriction effect of a throttle device 27 is essentially proportional to a loss coefficient i;, to the density p of the flowing medium and to the square of the flow velocity through the throttle device 37. The throttle device 37 can be embodied in different ways, e.g., as a collar, as a perforated plate, as a baffle through the design of the geometry, such as a corresponding selection of the diameter or length, of the connection bore 30, or, as in the example according to Fig. 2, the known Carnot shock loss can be utilized with a sudden expansion of the cross section of a flow channel. As is adequately known, the Carnot shock loss pv is thereby essentially proportional to the square of the flow velocity u through the flow channel and the square of the ratio of the cross-sectional surfaces after A2 and before Al of the expansion of the flow channel (loss coefficient Q and can be easily calculated with the known density of the flow medium p and adjusted as desired through the selection of the geometry of the flow channel, such as e.g., the connection bore 30:
In the current example according to Fig. 2 an area 31 with large cross section is provided in the area of the groove 24 in the entry area of the connection bore 30, from which area 31 the connection bore 30 starts, wherein the transition from entry area 31 to connection bore 30 can be rounded in order to reduce possible wear in this area and thus to keep the selected restriction effect constant over longer operating times of the sealing element 20. The opposite end 32 of the connection bore 30 has a sudden cross-sectional expansion at which the Carnot shock loss occurs. The speed through the connection bore 30 can thereby be adjusted with a certain pressure of the seal water p1 or at a certain volume flow Q (which can be controllable e.g., through a flow control valve in the supply line 21) through the cross-sectional surface of the connection bore 30 and/or preferably through the number of connection bores 30 arranged distributed over the circumference.
The throttle device 37 is now selected or designed such that a pressure loss based on the pressure difference po-pu to be sealed bearing against the sealing ring 20, of at least 1%, preferably at least 5%, preferentially at least 10%
occurs between the first hydrostatic bearing 22 and the second hydrostatic bearing 23.
The natural unavoidable pressure loss caused through a flow through a bore as far as possible free of loss, as desired in previous sealing rings, is typically lower by approx. two to three decimal powers, thus is negligible with respect to the adjustable "inner restriction." Thus the desired pressure loss according to the invention through the throttle device 37 is considerably higher than the unavoidable natural restricting effect (essentially through pipe friction;
possible deflections in the connection bore and as a result of a cross-sectional expansion).
The constricted center part of the connection bore 30 can also be achieved through a tubular insert 38, which is arranged in a larger through bore, as shown in Fig. 5. An arrangement of this type is in particular advantageous when the sealing ring 20 is produced from a soft material, e.g., of plastic, and the pipe insert 38 is then produced from a hard, rugged material in order to resist the flow at high flow velocity through the narrow connection bores 30 over the long term and thus to keep the adjusted intemal restriction essentially constant for as long as possible.
In order to be able to describe the function of the sealing ring 20, in Fig. 2 in addition the pressure distributions resulting in the axial and radial bearing 23, 22 are shown diagrammatically. The bearing medium is fed as described above via the supply line 21 with a volume flow Q or with a pressure p1 into the radial bearing 22. The volume flow Q of the bearing medium is divided into three flows in the radial bearing 22. One flow flows downwards and finally opens into the axial rotor side space with pressure pu. A second flow flows upwards and opens into the bearing space 36 with the pressure po prevailing at the rotor entry, which pressure naturally also acts on the sealing ring 20. A third flow, largest by far, flows via the connection bore 30 into the axial bearing 23. The division of the volume flow Q thereby results naturally through the radial bearing gap which is essentially adjusted through the pressure p1 prevailing in the hydrostatic bearing 22 and the pressure po at the rotor entry.
Due to basic known laws of fluid mechanics and depending on the geometry of the radial bearing 22 in the radial bearing 22, the volume flow Q causes the pressure distribution shown with a maximum pressure p1 in the groove 24 into which the supply line 21 opens. This pressure distribution causes the sealing ring 20 to lift in the radial direction. Radial lifting means in this context on the one hand that the sealing ring 20 correspondingly centers itself, since the pressure acts over the entire inner circumference, and on the other hand also expands.
The headwater pressure po but also according to the theory of elasticity the elastic restoring forces of the sealing ring 20 counteract this expansion. The maximum pressure p1 must therefore be large enough to be able to cause the lifting of the sealing ring 20 on the desired bearing gap, e.g., typically 50 -pm.
In the axial hydrostatic bearing 23 a pressure p2 is established, which according to the adjustable pressure loss between the two bearings 22, 23 is smaller than the pressure p1 in the radial hydrostatic bearing 22. In turn a pressure distribution in the axial bearing 23 can be adjusted via the width and position of the groove 25, in order to also lift the sealing ring 20 in the axial direction.
Through the provision of a restriction in the hydraulic connection, here connection bore 30, between the radial hydrostatic bearing 22 and the axial hydrostatic bearing 23 the pressure p1 in the radial bearing 22 can also be influenced via the pressure in the supply line 21 or the fed volume flow Q, whereby the radial bearing 22 can also be controlled.
Since the sealing ring 20 in real applications can have large diameters up to several meters, but small cross sections and light materials are desirable, the sealing ring 20 is very torsionally flexible, that is, very sensitive to moments, such as tilting moments or pulling moments. It is therefore important to keep the sealing ring 20 torque free during operation, since otherwise a ring very resistant to deformation would be required. The adjustment of a moment equilibrium is thereby possible through the selection of the geometry of the sealing ring 20, the hydrostatic bearings 22, 23 and the pressure distributions and the pressure level in the bearings 22, 23. As is easily realized, this can be achieved in that the grooves 24, 25 are arranged such that the balance of torques due to the pressure distributions acting on the sealing ring 20 is equalized around the surface cross section of the sealing ring cross section. In order to achieve this, in addition to the entire geometry of the sealing ring 20, such as the position and dimensions of the grooves 24, 25, the bearing gap widths, the outer dimensions etc., among other things the recess 29 is also used.
The sealing ring 20 consequently floats practically without friction and in a stable manner on two hydraulic bearings and is kept in balance only by the hydraulic compressive forces acting on the sealing ring 20. In balance thereby means that the sealing ring 20 is in an equilibrium of forces as well as free of torque.
If the sealing ring 20 is not arranged in a torsionally fixed manner, it additionally rotates at a lower rotational speed together with the rotor 2. This thereby also results in a dynamic gain in stability, since the limiting circumferential velocity or the flutter limit is thus raised. Furthermore, the friction losses are also lower due to the lower relative velocities.
With a design of a sealing element according to the invention, this can consequently be held in an equilibrium of forces and absence of torque can be ensured at the sealing element, whereby a tilting or pulling that is dangerous for the safe operation of the torsionally flexible sealing element can be prevented.
In the case of a rotationally fixed sealing ring 20, the rotational fixing can be carried out virtually arbitrarily, with the proviso that the axial and radial mobility of the sealing ring 20 in operation may not be hindered or may be hindered only negligibly. For example, a simple stationary limit stop can be provided on the turbine housing 12, against which a correspondingly shaped limit stop on the sealing ring 20 bears in order to prevent the rotation of the sealing ring 20.
Since the sealing ring 20 as described above floats in a virtually lossiess and frictionless manner on the pressure cushions forming in the hydraulic bearings, only very low circumferential forces develop on the sealing ring 20, typically in the range of several hundred N or less. With bearing forces in the hydraulic bearings 22, 23 of typically several hundred kN and greater (that is, greater by a factor of - 1000), these circumferential forces (and friction forces or retaining forces possibly developing therefrom at the anti-rotation element) can be disregarded, however. An anti-rotation element of this type consequently would not influence the axial and radial mobility of the sealing ring.
Through the high stability of a hydraulic bearing and the free mobility of the sealing ring 20, the sealing ring 20 is able to equa(ize oscillations of the rotor 2 and/or of the turbine housing 12, as well as axial displacements of the rotor 2, without losing the sealing action and without coming into contact with the rotor 2 and/or the turbine housing 12. The sealing ring 20 sustains virtually no wear thereby, since it floats in a virtually frictionless manner on liquid films, whereby the service life of a sealing ring 20 of this type is very high. Because the sealing ring 20 can be preferably constructed as a very slim, tight ring that has hardly any inertia forces, this action is further intensified.
The sealing ring 20 can be constructed to be very small in comparison to the dimensions of the rotor 2, edge lengths of several centimeters, e.g., 5 cm or cm, with outer diameters of several meters are completely adequate, and it can be manufactured of any desired material, such as steel, bearing bronze, plastic (e.g., PE). Furthermore, the bearing surfaces 26, 34 and 27, 33 can also be covered with a suitable layer, such as Teflon, bearing bronze, etc., in order to improve the properties of the seal still further. Typically the sealing ring 20 is manufactured from a softer material, such as, e.g., Teflon, bronze, etc., than the housing 12 or the rotor 2 of the hydraulic machine. As a rule it thus becomes lighter on the one hand and on the other hand in an extreme case the sealing ring 20 and not the rotor 2 or the housing 12 is damaged or even destroyed.
Fig. 3 shows a perspective representation of a sealing ring 20 according to the invention. A plurality of grooves 25 are inserted on a front face of the sealing ring 20, which grooves are part of the axial hydrostatic bearing 23, wherein at least one connection bore 30 opens in each of these grooves 25. A row of grooves 24 is inserted on the inside of the sealing ring 20, which grooves are part of the radial hydrostatic bearing 22 and wherein in turn at least one connection bore opens in each of these grooves 24. The number, size and distribution of the grooves 24, 25 and the connection bores 30 can thereby be selected accordingly to achieve an equilibrium of forces and moment equilibrium and to adjust the desired pressure distributions and the bearing gaps in the assigned hydrostatic bearings 22, 23.
However, the sealing ring 20, as shown in Figs. 3 and 4, can also be embodied divided in the circumferential direction at one or more dividing points 41, 44 which among other things facilitates the production or the transport of the sealing ring 20. In the exemplary embodiment shown, the sealing ring 20 is divided at two dividing points 41, 44 wherein the two halves are rigidly connected to one another at a first dividing point 41 via a connection means, such as, e.g., via dowel pins and a screw. At a second division point 44 the sealing ring 20 remains slotted in the assembled condition so that a gap 50 is formed in the circumferential direction, as shown in Fig. 4. This gap 50 is used essentially to render possible an automatic peripheral length compensation, which compensates for possible wear of the sealing ring 20 at the bearing surface 34, e.g., by fine solids (such as, e.g., sand) in the fed seal water, which have an abrasive action. The automatic peripheral length compensation results in that the bearing gaps (that is, the thickness of the liquid films in the bearings) are adjusted through the supply of a specific feed quantity with a specific pressure.
The slotted sealing ring 20 is compressed in the radial direction through the acting compressive forces that press the sealing ring 20 against the two hydrostatic bearings 22, 23 (the gap 50 at the dividing point 44 thus becomes smaller) so that the desired bearing gap is reestablished in the hydrostatic bearing. When the entire gap 50 is "used up," the two parts of the sealing ring thus strike against one another and a further wear can thus no longer be compensated, the sealing ring 20 can be replaced.
Since a slotted sealing ring 20 of this type, however, would naturally deform very considerably due to the high pressures acting thereon, in particular would bend and twist, corresponding measures have to be taken in order to also maintain the shape of the sealing ring 20 during operation. Moreover, seal water could flow through a slot or gap 50 of this type, which of course would negatively affect the function of the sealing ring 20. To this end on the one hand a shoulder 45 is now provided at the slotted dividing point 44 at a front face of the sealing ring arranged in the circumferential direction, against which shoulder a recess 47 on the opposite front face of the sealing ring 20 arranged in the circumferential direction bears. By pressing the recess 47 against the shoulder 45 a seal is realized which prevents seal water from flowing through the slot in the radial direction. Furthermore, the sealing ring 20 thereby cannot bend in the radial direction. On the other hand, a plate can be arranged on the outer front face of the sealing ring 20 facing away from a hydrostatic bearing, which plate extends over the gap 50. A seal would thus be created again which prevents an axial flow through the slot by seal water. Furthermore, an axial bending and a twisting of the sealing ring 20 would be prevented thereby. The plate could thereby be attached to the sealing ring 20 by means of suitable connection means, such as, e.g., by means of dowel pins and screws, wherein these connection means naturally may be arranged only on one half of the sealing ring 20 in order not to impede the mobility of the two sealing ring halves in the circumferential direction (which is important for the peripheral length compensation). Of course, arbitrary other suitable arrangements can also be provided in order to realize a corresponding seal and to prevent a deformation of the sealing ring 20. For example, instead of the plate, analogously to the radial direction as described above, the sealing ring 20of course could also have a shoulder on one half in the axial direction, which shoulder interacts with a corresponding recess on the other half.
Through the slot 44 the radial as well as the axial groove 24, 25 can be interrupted or the symmetrical arrangement of the groves 24, 25 around the circumference can be disturbed. Lower bearing forces could thus form in the area of this dividing point, since in this area the surfaces of the grooves 24, 25 on which the bearing pressure p1, p4 acts, would be smaller, which would lead to an uneven pressure load on the sealing ring 20. This could deform the deformable, elastic, slim sealing ring 20, which can lead to problems in practice with the low bearing gaps in the hundred pm range, such as e.g., semifluid friction conditions and, in the worst case, to contact between the sealing ring 20 and rotor 2 or turbine housing 12. In order to prevent this, the grooves 24, 25 adjoining the gap 50 can be embodied larger, or in this area another groove division could be provided in the circumferential direction and thus a different groove surface again. For example, the grooves 25 adjoining the gap 50 on the front face of the sealing ring 20 in the end facing towards the gap 50 could have an area 46 with a larger groove surface. Likewise the grooves 24 adjoining the gap 50 on the inside of the sealing ring 20 in the end facing towards the gap 50 can have an enlarged groove surface 42. With these enlarged groove surfaces 42, 46 a corresponding equalization and a reestablishment or retention of the of forces and moment equilibrium can be carried out.
If the sealing ring 20 is embodied to rotate in the same direction, it is advantageous to balance it, in particular with a slotted or divided embodiment.
The sealing ring 20 can, of course, have an arbitrary cross section, such as, e.g., an L-shaped cross section, wherein for reasons of manufacturing engineering a square or rectangular shape is preferred.
Until now, flat bearing surfaces 26, 34 and 27, 33 have always been assumed.
However, it is naturally also conceivable to embody the bearing surfaces 26, and 27, 33 not flat, but, e.g., concave or convex or stepped, wherein the fundamental principle of the seal according to the invention is not changed.
With bearing surfaces of this type 26, 34 and 27, 33 that are not flat only the pressure distributions would change somewhat, but this is clearly evident to one skilled in the art. It is likewise conceivable that the bearing surfaces 26, 34 and 27, based on the rotation axis of the turbine 1 are not axially parallel and axially perpendicular, but are arranged at an angle thereto.
To improve the bearing effect and the stability, one or more of the bearing surfaces 26, 34 and 27, 33 could also additionally be provided with adequately known hydrodynamic oil pockets to form an additional hydrodynamic bearing.
A seal according to the invention with a sealing ring 20 can naturally be provided at any suitable point and is not limited to the exemplary embodiments according to Fig. 2. For example the sealing ring 20 could also be arranged between the front face of the rotor 2 or hub 11 and the turbine housing 12. Likewise it is conceivable to provide a seal of this type at a suitable point between the ring 10 and the machine housing 12.
Likewise it would be conceivable with a rotationally fixed sealing ring 20, as known per se from the prior art, in the bearing surface 26, which is not subject to any relative motion regarding the assigned bearing surface 34 on the sealing ring 20, on both sides laterally next to the supply line 21 or at the side of a groove 24, to provide respectively one auxiliary sealing body between the seating ring 20 and the facing bearing surface 26 on the housing 12. With an auxiliary sealing body of this type, seal water would be prevented from exiting at this hydraulic bearing, whereby the losses through the seal according to the invention can be reduced still further. However, above all therewith excessive wear of the hydraulic bearing by exiting contaminated seal water could be prevented or considerably reduced. Of course, the auxiliary sealing bodies would thereby have to be used such that they develop the lowest possible action of force with a still adequate sealing action, since the sealing ring 20 is to remain supported in a floating manner. Since the sealing ring 20 is supported in a floating manner on the pressure cushions in a stable manner, this can be easily achieved via the level of the sealing gap to be adjusted. Through the possible low force on the auxiliary bodies, typically lower by a factor of 1000 or more than the hydraulic bearing forces, this in turn can be referred to as a floating bearing and it can be ensured that the axial and radial mobility of the sealing ring 20 are influenced only slightly and insignificantly.
The seal described above represents a largely tight seal. The entire water quantity flowing in flows through the rotor and can be converted into rotational energy. The leakage water losses are thereby reduced exclusively to the exiting bearing medium, thus based on the low gap heights in the hydraulic bearings, e.g., typically in the 10 -60 pm range with hydrostatic bearings 22, 23, are very low and can be recovered in part by the introduction of the leakage water into the main water flow F again.
In all phases of operation it should be avoided as far as possible that the sealing ring 20 comes into contact with the rotor 2 or the housing 12 or semifluid friction conditions form in the hydraulic bearings 22, 23, since then the sealing ring can be very easily damaged or even destroyed. At the start of the turbine 1 the sealing ring 20 should therefore already be lifted, i.e., the desired bearing gaps should already have been achieved. This can be easily achieved in that the supply of the hydraulic bearings 22, 23 is switched on first and only then the turbine 1 is switched on. In the case of a breakdown of the supply of the hydrostatic bearings 22, 23, for example, an emergency supply such as am air chamber, could be provided in order to avoid damage to the sealing ring 20 of the hydraulic machine, which would entail complex repair work.
Upon first putting into service the sealing ring 20, however, it can be desirable to adjust a controlled semifluid friction condition in the hydraulic bearings 22, 23, so that a bearing image can be ground in the bearing surfaces 26, 34 and 27, 33, whereby certain manufacturing inaccuracies can be compensated. However, since the bearing gaps are in the hundred pm range or lower, corresponding care is thereby naturally necessary.
In the description above to simplify matters, water is described as the bearing medium. Of course, the bearing medium, above all in the case of pumps, naturally can also be any other suitable medium, such as, e.g., oil.
Claims (15)
1. Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another, preferably between a housing (12) and a rotor (2) of a hydraulic turbomachine (1), with a sealing element (20), which with respect to the two parts is arranged supported on a first and a second hydraulic bearing (22, 23) in a floating manner, wherein the first and the second hydraulic bearing (22, 23) comprise at least two assigned bearing surfaces (26, 34 and 27, 33) and the sealing element (20) is supported by the hydraulic compressive forces acting on its bearing surfaces (33, 34) in two different directions and wherein a hydraulic connection (30), preferably a connection bore, is provided in the sealing element (20), which hydraulic connection connects the first hydraulic bearing (22) to the second hydraulic bearing (23) in a hydraulic manner and with a supply line (21) that opens into the first hydraulic bearing (22) and supplies bearing medium to the first hydraulic bearing (22), which flows via the hydraulic bearing (30) to the second hydraulic bearing (23), characterized in that a throttle device (37) is provided between the first hydraulic bearing (22) and the second hydraulic bearing (23) in the hydraulic connection (30) for setting a pressure loss, whereas a pressure loss, based on the pressure difference (po-pu) to be sealed bearing against the sealing element (20), of at least 1%, preferably at least 5%, preferentially at least 10%, can be set between the first hydraulic bearing (22) and the second hydraulic bearing (23) with the throttle device (37).
2. Arrangement according to claim 1, characterized in that the hydraulic connection (30) itself is provided as the throttle device (37).
3. Arrangement according to claim 1, characterized in that a pipe insert (38) in the hydraulic connection (30) is provided as a throttle device (37).
4. Arrangement according to one of claims 1 through 3, characterized in that the throttle device (37) is embodied as a collar, perforated plate or baffle.
5. Arrangement according to one of claims 1 through 4, characterized in that the throttle device (37) is embodied as a sudden expansion of the cross section of the hydraulic connection (30).
6. Arrangement according to one of claims 1 through 5, characterized in that a sealing ring is provided as a sealing element (20).
7. Arrangement according to claim 6, characterized in that the sealing ring (20) is embodied divided into at least two parts.
8. Arrangement according to claim 7, characterized in that the sealing ring (20) is embodied in a slotted manner at a number of the dividing points (44), and at the remaining dividing points (41) the parts of the sealing ring (20) are connected to one another rigidly.
9. Arrangement according to claim 8, characterized in that a gap (50) is provided at at least one slotted dividing point (44) to render possible a peripheral length compensation.
10. Arrangement according to claim 8 or 9, characterized in that a means to prevent seal water from flowing through the dividing point (44) is provided at the slotted dividing point (44), which means does not obstruct the mobility of the sealing ring (20) in the circumferential direction.
11. Arrangement according to one of claims 8 through 10, characterized in that a means for preventing a deformation of the sealing ring (20) is provided at the slotted dividing point (44), which means does not obstruct the mobility of the sealing ring (20) in the circumferential direction.
12. Arrangement according to one of claims 7 through 11, characterized in that at least one bearing recess (24, 25) adjoining a dividing point (41, 44) is embodied in the area of the dividing point (41, 44) with a different bearing recess surface from other bearing recesses or the remaining bearing recess.
13. Arrangement according to one of claims 1 through 12, characterized in that the sealing element (20) is arranged in a rotationally fixed manner by means of an anti-rotation element with respect to one of the two parts of the hydraulic turbomachine (1).
14. Arrangement according to claim 13, characterized in that the anti-rotation element exerts essentially only a circumferential force on the sealing element (20).
15. Arrangement according to one of claims 1 through 14, characterized in that at least one hydraulic bearing (22, 23) is a hydrostatic bearing.
Applications Claiming Priority (3)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| AT0183706A AT504394B1 (en) | 2006-11-03 | 2006-11-03 | ARRANGEMENT FOR SEALING BETWEEN TWO RELATIVELY MOVABLE PARTS OF A HYDRAULIC FLOW MACHINE |
| ATA1837/2006 | 2006-11-03 | ||
| PCT/AT2007/000490 WO2008052231A1 (en) | 2006-11-03 | 2007-10-22 | Arrangement for sealing two parts of hydraulic turbomachinery which can move relative to one another |
Publications (2)
| Publication Number | Publication Date |
|---|---|
| CA2668482A1 true CA2668482A1 (en) | 2008-05-08 |
| CA2668482C CA2668482C (en) | 2015-04-28 |
Family
ID=39125082
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA2668482A Expired - Fee Related CA2668482C (en) | 2006-11-03 | 2007-10-22 | Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another |
Country Status (5)
| Country | Link |
|---|---|
| EP (1) | EP2084395B1 (en) |
| AT (2) | AT504394B1 (en) |
| CA (1) | CA2668482C (en) |
| DE (1) | DE502007005085D1 (en) |
| WO (1) | WO2008052231A1 (en) |
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN101825179A (en) * | 2010-04-30 | 2010-09-08 | 天津市天发重型水电设备制造有限公司 | Spindle-combined sealing device of mixed-flow water turbine |
| US10428948B2 (en) | 2015-03-16 | 2019-10-01 | Nok Corporation | Seal ring |
| DE102020122601A1 (en) | 2020-08-28 | 2022-03-03 | Rolls-Royce Deutschland Ltd & Co Kg | Seal system, transmission with a seal system and gas turbine engine with a seal system |
Families Citing this family (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| KR20200020941A (en) * | 2017-07-10 | 2020-02-26 | 지멘스 에너지, 인코포레이티드 | Generator sealing assembly |
| DE102020203767B4 (en) * | 2020-03-24 | 2022-05-05 | Eagleburgmann Germany Gmbh & Co. Kg | Self-priming mechanical seal assembly |
Family Cites Families (7)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| AT298365B (en) * | 1970-06-22 | 1972-05-10 | Andritz Ag Maschf | Device for the mutual sealing of two rooms |
| CH598514A5 (en) * | 1975-08-29 | 1978-04-28 | Escher Wyss Ag | |
| US5052694A (en) * | 1986-07-08 | 1991-10-01 | Eg&G Sealol, Inc. | Hydrostatic face seal and bearing |
| US5558341A (en) * | 1995-01-11 | 1996-09-24 | Stein Seal Company | Seal for sealing an incompressible fluid between a relatively stationary seal and a movable member |
| EP0961059B1 (en) * | 1997-11-21 | 2006-02-01 | Nippon Pillar Packing Co., Ltd. | Static pressure noncontact gas seal |
| AT411092B (en) * | 2000-09-15 | 2003-09-25 | Gittler Philipp Dipl Ing Dr Te | SEALING THE WHEEL OF HYDRAULIC TURBO MACHINES |
| AT413049B (en) * | 2002-07-31 | 2005-10-15 | Philipp Dipl Ing Dr Te Gittler | SEAL BETWEEN TWO RELATIVELY MOVABLE PARTS OF A HYDRAULIC MACHINE |
-
2006
- 2006-11-03 AT AT0183706A patent/AT504394B1/en not_active IP Right Cessation
-
2007
- 2007-10-22 EP EP07815157A patent/EP2084395B1/en not_active Not-in-force
- 2007-10-22 WO PCT/AT2007/000490 patent/WO2008052231A1/en active Application Filing
- 2007-10-22 CA CA2668482A patent/CA2668482C/en not_active Expired - Fee Related
- 2007-10-22 DE DE502007005085T patent/DE502007005085D1/en active Active
- 2007-10-22 AT AT07815157T patent/ATE481568T1/en active
Cited By (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN101825179A (en) * | 2010-04-30 | 2010-09-08 | 天津市天发重型水电设备制造有限公司 | Spindle-combined sealing device of mixed-flow water turbine |
| US10428948B2 (en) | 2015-03-16 | 2019-10-01 | Nok Corporation | Seal ring |
| DE102020122601A1 (en) | 2020-08-28 | 2022-03-03 | Rolls-Royce Deutschland Ltd & Co Kg | Seal system, transmission with a seal system and gas turbine engine with a seal system |
Also Published As
| Publication number | Publication date |
|---|---|
| EP2084395A1 (en) | 2009-08-05 |
| WO2008052231A1 (en) | 2008-05-08 |
| CA2668482C (en) | 2015-04-28 |
| ATE481568T1 (en) | 2010-10-15 |
| AT504394A1 (en) | 2008-05-15 |
| DE502007005085D1 (en) | 2010-10-28 |
| AT504394B1 (en) | 2008-10-15 |
| EP2084395B1 (en) | 2010-09-15 |
Similar Documents
| Publication | Publication Date | Title |
|---|---|---|
| US7775763B1 (en) | Centrifugal pump with rotor thrust balancing seal | |
| US8646979B2 (en) | Hybrid hydro (air) static multi-recess journal bearing | |
| CA2285319C (en) | Thrust bearing | |
| CA2668482C (en) | Arrangement for sealing between two parts of a hydraulic turbomachine moveable relative to one another | |
| CA1327481C (en) | Centrifugal pump with hydraulic thrust balance and tandem axial seals | |
| JP4675643B2 (en) | Journal bearing | |
| US4546836A (en) | Downhole motor fluid flow restrictor | |
| US7731476B2 (en) | Method and device for reducing axial thrust and radial oscillations and rotary machines using same | |
| EP0521007A1 (en) | Control system for regulating the axial loading of a rotor of a fluid machine | |
| JP5650372B2 (en) | Rotating machine including passive axial balancing system | |
| JP5840390B2 (en) | Thrust control system for steam turbine | |
| KR100753180B1 (en) | Sealing in a hydraulic turbine unit | |
| CN107795577B (en) | Radial sliding bearing | |
| US20050087933A1 (en) | Seal for use between two mobile parts of a hydraulic machine | |
| JP5840389B2 (en) | Thrust control system for steam turbine | |
| CZ20021454A3 (en) | Device for compensation of axial shift in turbine machines | |
| JP6389785B2 (en) | Downhole compressor | |
| JP6189224B2 (en) | Journal bearing and steam turbine | |
| CN102308098A (en) | Turbomachine having a compensating piston | |
| JP5812973B2 (en) | Journal bearing and steam turbine | |
| CN209800343U (en) | axial thrust self-balancing device of centrifugal pump | |
| CN217582560U (en) | Efficient balancing device for multistage pump of offshore floating platform | |
| CN1179818A (en) | Improved hydrostatic seal | |
| US20200392965A1 (en) | Nuclear coolant pump seal and methods of sealing | |
| WO2014040641A1 (en) | Bearing system for rotor in rotating machines |
Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| EEER | Examination request | ||
| MKLA | Lapsed |
Effective date: 20171023 |