WO2009109355A1

WO2009109355A1 - Rotary union

Info

Publication number: WO2009109355A1
Application number: PCT/EP2009/001490
Authority: WO
Inventors: Daniel Birlinger
Original assignee: Hunger Maschinen Gmbh
Priority date: 2008-03-05
Filing date: 2009-03-03
Publication date: 2009-09-11
Also published as: DE102008012676A1

Abstract

A rotary union has a stator (12) and a rotor (14), which is mounted in the stator (12) using a rotationally-symmetric guide face (30) and can be rotated about an axis (16). The stator (12) has at least one first fluid line (52, 54, 62, 64, 90) and the rotor (14) has at least one second fluid line (22, 24, 26), which each open into the guide face (30) and are connected to one another upon rotation of the rotor (14) in the stator (12). Furthermore, at least one annular seal (66) is provided axially adjacent to the mouth, in which an annular seal body (86), which is disposed in the stator (12) or the rotor (14), is pressed at predetermined contact pressure against the rotor (14) or stator (12). Means are provided in order to press on the seal body (86) alternately using at least two different contact pressures.

Description

Rotary union

The invention relates to a rotary feedthrough with a stator and a rotatably mounted in the stator by means of a rotationally symmetrical guide surface, rotatable about an axis rotor, wherein the stator at least a first fluid conduit and the rotor has at least one second fluid conduit, each opening in the guide surface and during rotation the rotor in the stator are in communication with each other, and further wherein axially adjacent to the mouth at least one annular seal is provided, in which an arranged in the stator or the rotor, annular sealing body is pressed with a predetermined contact pressure against the rotor or the stator. The invention further relates to a method for operating a rotary feedthrough.

Finally, the invention relates to a use of a rotary feedthrough.

When a fluid under pressure is to be transferred to a device rotating relative to the unit in a housing-fixed unit, a rotary feedthrough is used. Rotary unions are well known, for example from DE 40 19 987 C2.

In a typical application in the context of the present invention, boring bars in derricks are successively screwed together in drilling down a borehole to form an increasingly longer drill string. For end screwing the boring bars an aggregate is used, which is arranged around the end of a first boring bar to be screwed in and fixed in space. In the unit is a device that is firmly connected to the end to be screwed and rotates when screwed with this. The rotary union serves to supply a working fluid (or more) of, for example, 150 bar pressure and a control fluid (or more) of lower pressure to the boring tool for the boring bars.

The screwing in takes place at a relatively low speed in the range between 15 and 25 rpm, depending on the steepness of the respective thread. When the end to be screwed is fully threaded into the opposite end of the next drill rod, the pressure is removed and drilling can continue. The device remains on the drill rod, so it rotates with this idle. When drilling a much higher speed is set, for example, from 90 to 100 U / min.

In this and many other applications, there is a problem that the seals provided in the rotary feedthrough, when in a working phase fluid under high pressure flows through the rotary union, with high Contact pressure must be operated to prevent the high-pressure fluid escapes. As long as the rotary feedthrough is running at low speed and for limited periods of time, this poses no problem. If the speed is increased in a subsequent idling phase, then increased gasket wear sets in, even if there is no working pressure during the idling phase and the seal only pressed by a bias. The seals heat up very strongly, for example to over 100 ^° C., which leads to a chemical change and, as a consequence, to a uselessness of the fluid.

Rotary feedthroughs are also known, for example from the already mentioned DE 40 19 987 C2, which use a hydrostatic seal, for example a gap seal. Here there is no seal friction in the idling phase and in the working phase. In the working phase, however, the leakage rate of such seals is considerable.

The invention is based on the object, a rotary feedthrough, a use and a method of the type mentioned in such a way that the above-mentioned disadvantages are avoided. In particular, in the working phase, the leakage rate of the rotary feedthrough should be low and in the idle phase, the seal friction low.

In a rotary feedthrough of the type mentioned above, this object is achieved in that means are provided to selectively press the sealing body with at least two different contact pressures.

In a method of the type mentioned above, the object is achieved in that the rotary feedthrough is operated with high contact pressure and small peripheral speed of the guide surface when the fluid lines are subjected to a fluid under working pressure, and with low contact pressure and high peripheral speed is operated when the fluid lines are depressurized.

Finally, we solved the problem by using a rotary feedthrough of the type mentioned above for screwing drill rods of derricks, for example, for oil or gas wells, especially in the offshore area.

The object underlying the invention is completely solved in this way.

According to the invention, namely, the contact pressure of the seals in the rotary feedthrough is set differently, namely in the working phase high and low in the idling phase. As a result, the leak rate is kept low in the working phase and reduced in the idle phase, the seal wear, and the gaskets heat up only slightly. This makes the rotary union according to the invention particularly suitable for the above-described application of screwing boring bars of derricks, for example oil or gas wells, in particular in the offshore sector, generally speaking for all applications in which two different operations are used, of which one characterized by a low speed with pressurized channels and the other by a high speed with no-pressure channels.

In a preferred embodiment of the invention, a contact pressure is substantially equal to zero.

This measure has the advantage that at idle the wear and thus the heating of the seals are reduced to a minimum.

Preferably, the sealing body arranged in the stator is out of contact with the rotor at a contact pressure or vice versa. This measure has the advantage that during the working phase, the seal is fully biased and fully sealed almost without leakage. In the idle phase, however, the individual pressure rotary seals are not in contact with their respective mating surface.

It is particularly preferred if further seals are provided in the frontal region between the stator and the rotor.

This measure has the advantage that the rotary feedthrough is completely sealed to the outside even if the rotary pressure seals in the idle phase are not in contact with their respective mating surface.

In a practical realization of this embodiment, the further seals are designed as shaft seals.

This measure has the advantage that commercially available components can be used.

According to the invention it is further preferred if the stator has a relative to the stator movable sealing bush, wherein the sealing body is arranged in the sealing bush and the contact pressure varies depending on the position of the sealing bushing relative to the stator.

This measure has the advantage that the contact pressure can be changed by simply moving the sealing bushing.

This is especially true when the rotor is provided in the guide surface with regions of different radial height and the sealing body is moved in a movement of the sealing bushing over the areas. This measure has the advantage that the method of the sealing bushing and thus the change of the contact pressure can be effected by a simple mechanism.

In a preferred practical embodiment, the at least first fluid conduit in the sealing bushing contains a first annular groove, wherein the sealing bush is displaceable in the direction of the axis and the regions are formed as second annular grooves in the guide surface of the rotor adjacent to the first annular grooves.

This measure has the advantage that the mechanical adjustment can be performed by a simple axial movement.

A particularly good effect is achieved in this case that the annular grooves have a rounded cross-sectional shape, and that preferably the sealing body rests with a tapered bearing side on the guide surface.

This measure has the advantage that in the process of the sealing bush, the sealing body is continuous, i. without jerky transition, is relaxed from its biased position, in which he then preferably no longer touches its respective mating surface. In this case, the contact pressure can be reduced to virtually zero, when the tapered bearing side dips into the rounded cross-sectional shape.

In this embodiment, it is also particularly preferred if the sealing bush is hydraulically displaceable.

This measure has the advantage that the displacement of the sealing bush and thus the change in the contact pressure can be remotely controlled and quickly.

For this purpose, the sealing bush is preferably formed at the end as an annular piston, and the annular piston runs in an annular cylinder formed by the stator. This measure has the advantage that a compact size is created.

In a further embodiment of the invention, the sealing bush is hydraulically displaceable by means of a fluid tapped from one of the fluid lines.

This measure has the advantage that the need for separate fluid lines for moving the sealing bush is eliminated.

In another embodiment, in addition to at least one ring seal, a channel opens into the guide surface, and the channel is substantially depressurized.

This measure has the advantage that changes in shape in the area of the fluid lines and the seals can be kept low, in particular in the area of fluid lines which serve as control lines.

It is preferred if the channel leads to a lateral pressure chamber which is bounded on the one hand by a further seal and on the other hand by an annular seal.

This measure has the advantage that a possibly accumulating pressure by moving the sealing bushing can not affect the shaft seals, which typically only withstand a pressure of up to 5 bar.

In a further exemplary embodiment of the invention, the rotor is mounted in the stator by means of two bearings arranged axially on both sides of the sealing bush, wherein in each case a cavity is located between the bearings and the sealing bush, and the cavities are connected to one another via a leakage line. This measure has the advantage that during displacement of the sealing bush, pressure equalization can take place between the annular cavities at the two axial ends of the sealing bush.

In a first development of this embodiment, the leakage line runs axially through the sealing bushing.

This measure has the advantage that the cable routing is particularly simple and short.

In a second embodiment of the embodiment, however, the leakage line extends axially through the rotor.

This measure has the advantage that more installation space is available. As a result, the leakage line can be dimensioned with a larger diameter. Accordingly, a pressure equalization is possible even with rapid changes in position of the sealing bush with a small time constant.

The bearings are preferably sealed on their side facing away from the sealing bushing to the outside via the other seals and preferably designed as shaft sealing rings. Further preferably, the shaft seals are provided with an axially inner sealing lip and with an axially outer dust protection lip.

This measure has the advantage that air ingress is avoided in the area of the shaft sealing rings, in particular in the event of a rapid change in position of the sealing bushing. The dust protection lip then has the additional advantage that it rests on the vacuum side and provides protection against air ingress.

In embodiments of the method according to the invention, the small peripheral speed is below 0.5 m / s, and the high contact pressure is between 100 and 200 bar, preferably about 150 bar, further preferably is the high peripheral speed above 0.5 m / s, and the low contact pressure is in particular less than 0.5 bar.

Further advantages will become apparent from the description and the accompanying drawings.

It is understood that the features mentioned above and those yet to be explained below can be used not only in the particular combination given, but also in other combinations or in isolation, without departing from the scope of the present invention.

Exemplary embodiments of the invention are illustrated in the drawing and will be explained in more detail in the following description. Show it:

Figure 1 is a side view of an embodiment of an inventive

Rotary union;

Figure 2 shows a longitudinal section through the rotary feedthrough of Figure 1 along the line II-II;

FIG. 3 shows a detail of FIG. 2 in a working phase on a greatly enlarged scale;

Figure 4 is a view like Figure 3, but for an idle phase; and

Figure 5 in greatly enlarged scale a cross-sectional view of a

Shaft seal, as it can be preferably used for axial sealing of the bearings between the stator and rotor of the rotary feedthrough according to the invention. In Fig. 1 and 2, 10 designates a rotary union as a whole. The rotary leadthrough 10 has a stator 12 and a rotor 14 rotatable in the stator 12. The axis of rotation of the rotor 14 is denoted by 16, and the rotation is indicated by an arrow 18. The stator 12 and the rotor 14 both preferably have substantially the shape of a hollow cylinder, as can be seen clearly from FIG. The longitudinal axis of the hollow cylinder coincides with the axis of rotation 18.

In a practical embodiment for the use according to the invention in the offshore sector, the rotary feedthrough is about 600 mm long and has a diameter of about 420 mm. However, it should be understood that this is just one of many possible examples.

The rotor 14 has a first sleeve 20. Axial channels extend in the first sleeve 20, one of which is shown at 22 in FIG. From the inner end of the axial channel 22 is a radial channel 24, which opens at an outer surface of the first sleeve 22. The outer end of the axial channel 22 terminates in a port 26 in a left end side 28 of the rotor 14 in FIG.

It is understood that a plurality of such channels 22, 24, 26 or lines over the circumference of the first sleeve 20 and the end face 28 may be distributed, for example, seven or nine such channels. A distinction is made between working channels and control channels. Working channels usually have a larger cross-section and conduct a working fluid, which is needed for example for actuating piston-cylinder units. Control lines, on the other hand, usually have a smaller cross-section and conduct a fluid which is needed for actuating control elements, in particular valves.

On an outer peripheral surface 30 of the rotor 14, which also serves as a guide surface between the stator 12 and the rotor 14, end bearings 32a and 32b are arranged, which rotatably support the rotor 14 in the stator 12. The bearings 32a and 32b are in Figure 2 shown only schematically, because this type of storage is known in the art.

The stator 12 has a second sleeve 40, which is arranged coaxially with the first sleeve 20. The second sleeve 40 is provided with a window 42.

Between the second sleeve 40 and the first sleeve 20 is a sealing bush 50. The sealing bushing 50 is axially displaceably mounted between the sleeves 20 and 40, as indicated by a double arrow 51. In the direction of rotation 18, however, it is connected to the stator 12, so that it forms part of the stator 12. The sealing bushing 50 is provided with a plurality of ports, namely working ports 52 and control ports 54 associated with the above-mentioned working or control ports. The terminals 52 and 54 are accessible through the window 42, wherein the window 42 is dimensioned so that this accessibility over the full displacement (arrow 51) of the sealing bushing 50 is maintained. The fluid lines connected to the ports 52 and 54 from the outside are formed at least over a certain length as flexible hose lines to follow the displacement of the sealing bushing 50 can.

Between the sealing bushing 50 and the bearings 32a and 32b, there are a right pressure chamber 58 and a left pressure chamber 60 in FIG. 2. By introducing a fluid into one of the two pressure chambers 58 and 60, the sealing bushing 50 can be displaced axially in the direction of the arrow 51 , The required for the displacement fluid pressure in the pressure chambers 58 and 60 can be conveniently provided by the working or the control pressure is tapped in one of the working or control channels (not shown). This can be effected, for example, in that the pressure chambers 58 and 60 are pressurized earlier in time than a tapped annular groove 62 or 64. Then the sealing bushing 50 is already in the desired end position, if the corresponding other connections 52 and 54, respectively Pressure is given. On the inner peripheral surface of the sealing bush 50, first annular grooves 62 are designed as working channels and second annular grooves 64 as control channels. The first annular grooves 62 are outside the plane of the drawing of Figure 2 with the working ports 62, and the second annular grooves 64 are connected to the control terminals 54 in connection.

On both sides of the annular grooves 62 and 64, ring seals 66 may still be located in the inner peripheral surface of the sealing bushing 50, of which only four are shown next to the second annular grooves 64 in FIG. 2 for the sake of clarity. Preferably, one will provide the ring seals 66 only on both sides of the larger pressure annular grooves 62 and for reasons of space on both sides of the smaller control annular grooves 64 renounce.

In the end region of the rotary feedthrough 10 further seals 68a, 68b are provided on the bearings 32a, 32b, which are preferably formed as shaft seals, as will be explained below with reference to FIG 5. The further seals 68a, 68b are designed for the high rotational speeds of the leelaufphase and, for example, for an operating pressure of 5 bar maximum.

Furthermore, 50 radial pressure relief lines 70 may be provided in the sealing bushing, which open between the optional ring seals 66 in the inner peripheral surface of the sealing bushing 50. The leftmost pressure relief line 70 in Figure 2 has a left axial branch 72a leading to a left pressure relief cavity 73a. The cavity 73a is located between the left in Figure 2 bearing 32a and the adjacent thereto radial end face of the sealing bushing 50. Similarly, on the right side of the sealing bushing 50, a right axial branch 72b and a right cavity 73b can be seen, the right Cavity 73b in the operating position shown in Figure 2 has the volume zero. In the sealing bush 50, an axial manifold 78 is provided. The optionally provided pressure relief lines 70 are connected to the manifold 78. This is in turn connected to a terminal 80, which is also accessible through the window 42 from the outside. The port 80 is connected to a pressureless fluid tank. Therefore, the line system formed from the lines 70, 72, 78 is depressurized or is at a very low pressure level of less than 0.5 bar.

The optional pressure relief lines 70 are arranged in particular between the annular grooves 64 belonging to the control connections 54. The channels 70 ensure that, for example, an increase in pressure in an annular groove 64, the optional annular seal 66 deformed, but this does not affect the adjacent annular groove 64 in such a way that changes by a change in volume, the pressure in the adjacent annular groove 64, because the channel 70th depressurized and connected to a tank. The pressure relief lines are useful in many cases, but they can be omitted in other cases.

The branches 72a and 72b, together with the manifold 78 as a leakage line connect the cavities 73a and 73b with each other. In the axial movement of the sealing bushing 50, an oil volume is moved because the volumes of the cavities 73a and 73b change in opposite directions. If the sealing bushing 50 is moved to the left, for example, from the position shown in FIG. 2, an overpressure is created in the left-hand cavity 73a and a negative pressure in the right-hand cavity 73b. The pressure compensation then takes place via the leakage line 72a-78-72b. Of course, this pressure equalization depends on how fast the displaced oil volume can flow through the leakage line 72a-78-72b. Even if the cross section of the leakage line 72a-78-72b is made large, the time constant of the pressure compensation remains finite.

In the embodiment shown in Figure 2, the leakage line 72a-78-72b passes through the sealing bushing 50. Although this results in a short line course without kinking, but the dimensioning of the cross section are limited, because the sealing bushing 50 can only have a limited radial thickness.

Alternatively, it can therefore be provided to arrange the leakage line in the rotor 14 (not shown), wherein the leakage line then similar to the axial channel 22, so with a larger cross section and the cavities 73a and 73b would connect. In this case, the non-pressurized port 80 would continue accordingly radially inward.

In this case, one has in the dimensioning and arrangement of the annular grooves 62 and 64 more degrees of freedom. Preferably, one will then place the larger pressure annular grooves 62 on one axial side and the smaller control annular grooves 64 on the other axial side and concentrate all the annular grooves 62, 64 in the axial center to avoid complicated holes near the end surfaces ,

The cavity 73a - and correspondingly also the cavity 73b - is bounded by the further seal 68a and in the working phase by the optional ring seal 66. This has the consequence that a possibly accumulating pressure by moving the sealing bushing 50 can not affect the further seal 68a, which, as mentioned, withstands only a lower operating pressure.

Details of the ring seals 66 according to the invention are shown in FIGS. 3 and 4.

The ring seals 66 are formed by third, rectangular annular grooves 82, in whose base an elastic O-ring 84 is inserted. On the O-ring 84, an annular, largely inelastic sealing body 86 is placed. The sealing body is provided on its inner circumference with a support side 88, which is preferably designed to be tapered. Such ring seals are known in the art. FIG. 3 shows a working state of the rotary feedthrough in which the working channels are pressurized, for example with an operating pressure of 150 bar, and in which the rotor 14 rotates in the stator 12 with low peripheral speed v of the outer peripheral surface of less than 0.5 m / s. In rotary unions of conventional size, as exemplified above, this corresponds to a speed range of, for example, 15 to 25 U / min.

Figure 4 shows an idle state of the rotary feedthrough in which the working channels are substantially unpressurized and in which rotates the rotor 14 in the stator 12 at high circumferential speed v of the outer peripheral surface 30 of more than 0.5 m / s, which in the aforementioned size one Speed range of, for example, 90 to 100 rev / min corresponds.

In the working state of FIG. 3, the rotary feedthrough 50 is preferably axially positioned so that the first annular grooves 62 are substantially aligned with the radial channels 24. The working fluid can flow unhindered from the working ports 52 in the stator 12 to the axial channels 22 and the terminals 26 in the rotor 14, as indicated by arrows 92.

In this working state, the bearing sides 88 of the ring seals 66 rest on the cylindrical outer peripheral surface 30 of the rotor 14. The depth and the width of the third annular grooves 82 are dimensioned such that the O-rings 84 are radially compressed in this state, so that a high contact pressure of the bearing sides 88 on the outer peripheral surface 30 is formed and thus a high sealing effect or a low leakage rate in particular between the annular grooves 62 is effected.

In this state of the high fluid pressure acted upon first and second annular grooves 62, 64, the pressure relief acts through the pressure relief lines 70 between the annular seals 66. The pressure relief bends, as mentioned, in particular an influence of the control annular grooves 64 with each other by changing the behavior of one ring seal 66 to the next ring seal 66 by deformation of the latter, because the adjacent ring seals 66 are separated from each other by one of the unpressurized pressure relief lines 70.

In the outer peripheral surface 30 fourth annular grooves 94 are now introduced in the region of the annular seals 66, which preferably have a rounded axial profile at the bottom. These fourth annular grooves are inactive in the working state shown in FIG. The annular grooves 94 are adapted to receive a non-preloaded annular seal 66, wherein the annular seal 66 preferably in a working position, which is determined by an axial end position of the sealing bushing 50 does not touch their mating surface in the outer peripheral surface 30.

If now is changed from the working state to the idle state, then the sealing bush 50 shifts to the position shown in Figure 4. In this position, the working port 52 and the first annular groove 62 are offset axially by an amount .DELTA.z with respect to the radial channel 24. The sealing bodies 86 can now constantly slide into this 94 due to the rounded shape of the fourth annular grooves.

In this neutral position, the fourth annular grooves 94 are aligned with the annular seals 66. This results in that the sealing bodies 86 have entered with their support side 88 in the fourth annular grooves 94. As a result, the O-rings relax from 84 (FIG. 3) to 84 '(FIG. 4). The contact pressure of the ring seals 66 is thereby significantly reduced and may even drop to zero. It is particularly preferred if the sealing body 86 in the idle position have no contact with the rotor 14, as shown in Figure 4.

It is understood that the invention is not limited by the above description of an embodiment. In particular, the invention also includes embodiments in which the respective elements explained are arranged in kinematic reversal instead of on the stator on the rotor and vice versa. Figure 5 shows an enlarged view of an embodiment of the further seal 68a, which embodiment can of course also be selected for the further seal 68b.

The seal 68a is designed with an annular groove 100 into which a shaft seal 102 is inserted. Shaft seal of this type are known per se, for example from DE 867 189 C. The shaft seal 102 has axially inside a sealing lip 104 which is biased by a sealing spring. Furthermore, a dust protection lip 108 is provided axially on the outside. The lips 104 and 108 ride on a surface 110 of the first sleeve 20.

As already stated above, a pressure equalization between the cavities 73a and 73b in the rapid process of the sealing bushing 50 only finally finishes well, so that a negative pressure forms in one of the cavities. This negative pressure has the consequence that air can penetrate from the outside into the cavity in question.

In the case of the shaft sealing ring 102 according to FIG. 5, a negative pressure in the cavity 73a now causes the dust protection lip 108 to be attracted against the surface 110, so that an air inlet is prevented. Besides, the dust-proof lip 108 naturally has its own effect of preventing the ingress of dirt.

Claims

Godfathers

1. Rotary feedthrough with a stator (12) and in the stator (12) by means of a rotationally symmetrical guide surface (30) mounted about an axis (16) rotatable rotor (14), wherein the stator (12) at least a first fluid line (52, 54, 62, 64, 90) and the rotor (14) at least one second fluid line (22, 24, 26), each in the guide surface (30) open and upon rotation of the rotor (14) in Stator (12) communicate with each other, and further wherein axially adjacent to the mouth at least one annular seal (66) is provided, in which in the stator (12) or the rotor (14) arranged, annular sealing body (86) with a predetermined contact pressure is pressed against the rotor (14) or the stator (12), characterized in that means are provided to selectively press the sealing body (86) with at least two different contact pressures.

2. Rotary feedthrough according to claim 1, characterized in that a contact pressure is substantially equal to zero.

3. Rotary feedthrough according to claim 1 or 2, characterized in that in the stator (12) arranged sealing body (86) at a contact pressure out of contact with the rotor (14) or vice versa.

4. Rotary feedthrough according to one or more of claims 1 to 3, characterized in that in the frontal region between the stator (12) and rotor (14) further seals (68 a, 68 b) are provided.

5. Rotary feedthrough according to claim 4, characterized in that the further seals (68 a, 68 b) as shaft seals (102) are formed.

6. Rotary feedthrough according to claim 5, characterized in that the shaft sealing rings (102) are provided with an axially inner sealing lip (104) and with an axially outer dust protection lip (108).

7. rotary feedthrough according to one or more of claims 1 to 6, characterized in that the stator (12) has a relative to the stator (12) movable sealing bushing (50) that the sealing body (86) in the sealing bushing (50) is and that the contact pressure in dependence on the position of the sealing bushing (50) relative to the stator (12) varies.

8. Rotary feedthrough according to claim 7, characterized in that the rotor (14) in the guide surface (30) is provided with regions of different radial height and that the sealing body (86) is moved in a movement of the sealing bush (50) over the areas.

9. Rotary feedthrough according to claim 8, characterized in that the at least first fluid line (52, 54, 62, 64, 90) in the sealing bushing (50) includes a first annular groove (62, 64) that the sealing bush (50) in the direction the axis (16) is displaceable and that the areas as second annular grooves (94) in the guide surface (30) of the rotor (14) adjacent to the first annular grooves (62, 64) are formed.

10. Rotary feedthrough according to claim 9, characterized in that the annular grooves (94) have a rounded cross-sectional shape, and that preferably the sealing body (86) rests with a tapered bearing side (88) on the guide surface.

11. Rotary feedthrough according to claim 9 or 10, characterized in that the sealing bush (50) is hydraulically displaceable.

12. Rotary feedthrough according to claim 11, characterized in that the sealing bushing (50) is designed as an end ring piston and that the annular piston in an annular cylinder formed by the stator (12).

13. Rotary feedthrough according to claim 11 or 12, characterized in that the sealing bushing (50) by means of a fluid from one of the fluid lines (52, 54, 62, 64, 90) tapped fluid is hydraulically displaceable.

14. Rotary feedthrough according to one or more of claims 9 to 13, characterized in that in addition to at least one annular seal (66), a channel (70) opens into the guide surface (30) and that the channel (70) is substantially depressurized.

15. Rotary feedthrough according to claim 14, characterized in that the channel (70) leads to a lateral pressure chamber (73) which is bounded on the one hand by a shaft seal (68 a, 68 b) and on the other hand by a ring seal (66).

16. Rotary feedthrough according to one or more of claims 7 to 15, characterized in that the rotor in the stator by means of two axially on both sides of the sealing bush (50) arranged bearing (32a, 32b) is mounted, that between the bearings (32a, 32b) and the seal bushing (50) each having a cavity (73a, 73b) is located, and that the cavities (73a, 73b) via a leakage line (72a, 78, 72b) communicate with each other.

17. Rotary feedthrough according to claim 16, characterized in that the leakage line (72 a, 78, 72 b) extends axially through the sealing bushing (50).

18. Rotary feedthrough according to claim 16, characterized in that the leakage line extends axially through the rotor.

19. Rotary feedthrough according to claim 4 and one or more of claims 16 to 18, characterized in that the bearings (32 a, 32 b) on its side remote from the sealing bush (50) side to the outside via the further seals (68 a, 68 b) are sealed.

20. Rotary feedthrough according to claim 19, characterized in that the further seals (68 a, 68 b) as shaft seals (102) are formed.

21. Rotary feedthrough according to claim 20, characterized in that the shaft sealing rings (102) are provided with an axially inner sealing lip (104) and with an axially outer dust protection lip (108).

22. A method for operating a rotary feedthrough (10) according to one or more of claims 1 to 13, characterized in that the rotary feedthrough (10) with high contact pressure and small peripheral speed (v) of the guide surface (30) is operated when the fluid lines ( 52, 54, 62, 64, 90, 22, 24, 26) are acted upon by a fluid under working pressure, and operated at low contact pressure and high peripheral speed (v), when the fluid lines (52, 54, 62, 64, 90, 22, 24, 26) are depressurized.

23. The method according to claim 22, characterized in that the small peripheral speed (v) is less than 0.5 m / s and the high contact pressure between 100 and 200 bar, preferably about 150 bar, is.

24. The method according to claim 22 or 23, characterized in that the high peripheral speed (v) is more than 0.5 m / s and the low contact pressure is less than 0.5 bar.

25. Use of a rotary feedthrough (10) according to one or more of claims 1 to 14 for screwing boring bars of derricks, at- For example, for oil or gas wells, especially in the offshore sector.