WO2022268356A1

WO2022268356A1 - Screw compressor

Info

Publication number: WO2022268356A1
Application number: PCT/EP2022/000056
Authority: WO
Inventors: Stefan Losch; Ole Fredrich; Winfried RENWANZ; Christian Ramsch
Original assignee: Gea Refrigeration Germany Gmbh
Priority date: 2021-06-22
Filing date: 2022-06-21
Publication date: 2022-12-29
Also published as: JP2024525208A; DE102021003198A1; EP4359672A1

Abstract

The invention relates to a screw compressor (10), comprising: - at least one suction gas inlet (12); - at least one screw rotor (14), which has a screw-rotor center axis extending in an axial direction; - at least one compressed gas outlet (16); - a housing (22), which at least partly surrounds the screw compressor (10); and - a balancing piston (28), which has a balancing-piston center axis, can be moved along the balancing-piston center axis, and is designed to produce a counterforce to a force which acts, in the axial direction, on the screw rotor (14); wherein: the balancing piston (28) has at least a first balancing-piston portion (28a) and a second balancing-piston portion (28b); the first balancing-piston portion (28a) has a first effective surface (29a), to which pressure can selectively be applied by a pressurized fluid, more particularly oil, more particularly by means of a first valve (60, 74); and the second balancing-piston portion (28b) has a second effective surface (29b), to which pressure can selectively be applied by a pressurized fluid, more particularly oil, more particularly by means of a second valve (62, 74).

Description

screw compressor

The present invention relates to a screw compressor according to the preamble of patent claim 1.

In a screw compressor, gas is drawn in from a suction port arranged on a suction side and transported by at least one screw rotor, in certain embodiments also two screw rotors, in the direction of a gas outlet arranged on a pressure side (i.e. pressure side), the gas being compressed to a higher pressure in the process. In the following, the screw rotor is also simply referred to as “rotor”.

The difference in pressure between the gas on the suction side of the rotor and the higher pressure on the pressure side of the rotor creates an axial force that pushes the rotor towards the suction side.

The rotor is supported by radial bearings on the suction and pressure side and by axial bearings. Typically, the axial bearings are designed as roller bearings.

CONFIRMATION COPY In a typical operating case, an axial force acts as described, which pushes the rotor in the direction of the suction side. This axial force is directly absorbed by the axial bearings, usually one or more deep groove or angular contact ball bearings, which are directed towards one another, and is derived via a housing enclosing the screw compressor. Especially when high pressure differences occur as a result of the compression between the suction and pressure sides, the forces that occur can be so great that the axial bearings are overloaded or their service life is reduced to an undesirably small value.

Screw compressors usually have an oil system, as described for example in DD 108 797 A1, which provides oil for lubricating bearings, for moving a control slide for power control of the screw compressor and for other tasks. The pressure of this oil is usually almost as high as the final pressure at the gas outlet on the pressure side, and even higher if an oil pump is used.

Screw compressors also often have a compensating piston, as is described, for example, in DD 80 961 A1 and DE 10 2009 038 937 A1.

This is often arranged on the pressure side, as shown in the documents described in more detail above. The compensating piston is attached to the shaft of the screw rotor and is sealed to the housing by means of a labyrinth seal. Pressurized oil is introduced into a space next to the balance piston, so that the pressure of the oil creates a force in the direction of the pressure side, which opposes the axial force of the rotor resulting from a gas force in the direction of the suction side and thus the load to be absorbed by the thrust bearings reduced.

Under certain operating conditions, it can happen that the axial force in the direction of the pressure side, which is generated by the balancing piston, is greater than the axial force of the rotor due to the gas force in the direction of the suction side. In this case, the rotor is moved in the direction of the pressure side, one speaks of an "axial thrust reversal".

This can happen when the control slide is pushed towards the pressure side, to operate the compressor at partial load with reduced displacement. This forms a bypass to the suction side and increases the suction pressure. Axial thrust reversal can also occur at full load when the pressure of the aspirated gas at a suction gas inlet is relatively high to the pressure of the gas at a pressure gas outlet.

In the case of axial thrust reversal, the axial bearing must also absorb the axial forces, otherwise the rotor would shift in the direction of the pressure side and collide with the housing.

In a conventional screw compressor design, the axial bearing is supported by disk springs in the direction of the force, i.e. in the direction of the pressure side. The axial bearings are preloaded safely and without play by these elastic disc springs, whereby relatively large dimensional tolerances of the individual components are accepted and individual components, e.g. roller bearings, can be exchanged very easily during maintenance work.

The force of the disc springs must be dimensioned in such a way that it is greater than the greatest possible force in the case of axial thrust reversal.

A large disk spring force is disadvantageous, however, as it preloads the thrust bearings in addition to the gas force, reducing the expected life of the thrust bearings to a fraction of what bearings of this size would achieve with optimum preload.

Accordingly, the object of the present invention is to specify a screw compressor in which the axial force to be absorbed by the axial bearings can be reduced, but it is also ensured that no axial thrust reversal occurs in all possible operating or load states or at least the axial force that occurs is reduced to an acceptable level value is reduced.

This problem is solved by a screw compressor with the features of claim 1. Accordingly, the object of the present invention is achieved by a screw compressor, which has at least one suction gas inlet, at least one screw rotor, which has a screw rotor central axis that extends along an axial direction, and at least one compressed gas outlet, as well as a housing at least partially enclosing the screw compressor and has a compensating piston which has a compensating piston center axis, the compensating piston being arranged to be displaceable along the compensating piston center axis and being designed to generate a counterforce to a force occurring on the screw rotor directed in the axial direction, the compensating piston having at least one first compensating piston section and has a second compensating piston section, wherein the first compensating piston section has a first active surface, which is selectively, in particular via a first valve, with a pressurized fluid, in particular oil , can be pressurized, and wherein the second compensating piston section has a second effective surface, which can be pressurized selectively, in particular via a second valve, with a pressurized fluid, in particular oil. Optionally, the first active area and the second active area have different sizes.

A desired counterforce can thus be generated in one of four force levels by selectively applying pressure, in particular oil pressure, to the first or the second active surface or both active surfaces or neither of the two active surfaces. As a result, the axial force to be absorbed by the axial bearings can be reduced, with the operating or load states of the compressor generally not causing any reversal of axial thrust or the axial force occurring being at least reduced to an acceptable value.

In one possible embodiment, the compensating piston extends in the axial direction, with the first compensating piston section and the second compensating piston section being arranged one after the other in the axial direction, and with the first compensating piston section having a circular outer circumference with a diameter dl in its radial direction and with the second compensating piston section in a circular outer periphery in its radial direction with a diameter d2, where dl is smaller than d2.

Optionally, the compensating piston is mounted in a compensating piston bushing, which is at least partially arranged in the housing of the screw compressor, in particular designed as a recess in this.

In one possible embodiment, a labyrinth seal, in particular a labyrinth seal on one side, is arranged between at least one compensating piston section, in particular between each compensating piston section and the compensating piston bushing. In addition or as an alternative to this, a contacting seal, in particular a radial shaft seal or a lip seal, can be arranged between at least one compensating piston section, in particular between each compensating piston section and the compensating piston bushing.

The compressor can have a controller, in particular an electronic controller, which is designed to selectively apply the pressurized fluid, in particular oil, to the first effective surface and/or the second effective surface.

Furthermore, the first active surface can be designed as a circular ring and have a first active surface inner diameter dli and as a first active surface outer diameter the outer diameter dl and the second active surface can be designed as a circular ring and a second active surface inner diameter d2i and as a second active surface outer diameter the outer diameter d2, where the relation dli<dl=d2i<d2 optionally applies.

Optionally, the compressor has a control slide in a

Slider guide recess is arranged to be movable back and forth, for power control of the screw compressor, which is designed to apply the first effective surface and / or the second effective surface with the pressurized fluid, in particular oil, to apply, in particular selectively to apply. One or more directional control valve actuation elements can be at least partially integrated in the slide guide recess, which are designed to act on the first effective surface and/or the second effective surface with the pressurized fluid, in particular oil, by opening and closing a respective assigned directional control valve , In particular to apply selectively. The one or more directional control valve actuators may be mechanically operable, extend at least partially into the spool guide recess of the control spool, and be operable by the control spool.

Alternatively or additionally, one or more oil line recesses can open into the slide guide recess, which are designed to apply the pressurized fluid, in particular oil, to the first effective surface and/or the second effective surface, in particular to apply it selectively, with the The control spool has a recess and is reciprocally disposed in the spool guide recess to positions in which the control spool closes all of the oil passage recesses, in which one or more oil passage recesses open into the recess of the control spool, and in which all of the oil passage recesses open into the recess of the control spool.

Optionally, the pressurized fluid with which the first effective surface can be pressurized and the pressurized fluid with which the second effective surface can be pressurized can have an approximately constant pressure during operation of the screw compressor, in particular approximately the same pressure exhibit.

The respective features described can be implemented individually or in any combination. The invention is described below with reference to the accompanying drawings using exemplary embodiments. Further optional features of the invention are also specified in the following description of the figures.

In the drawings show:

Fig.l is a sectional view of an embodiment of an inventive screw compressor;

Fig. 2 shows an enlarged view of a section of the upper right area of Fig. 1, including the compensating piston;

3 shows a detailed illustration of an arrangement for controlling or regulating the application of pressure to the compensating piston;

4 shows a detailed illustration of an alternative arrangement for controlling or regulating the application of pressure to the compensating piston;

5 shows a detailed representation of a control slide for a capacity control of the compressor of FIG. 1 in a full-load position; and

6 shows a detailed representation of the control slide for the capacity control of the compressor of FIG. 1 in a partial load position for a small partial load.

An embodiment of a screw compressor 10 according to the invention is shown in sectional view in FIG. In the screw compressor 10, gas is drawn in from a suction gas inlet 12 (suction connection, gas inlet on the suction side) and transported by a screw rotor 14 in the direction of a pressure gas outlet 16 (gas outlet on the pressure side), the gas being compressed to a higher pressure. The screw rotor 14 is rotationally driven via a shaft 13, to which it is non-positively connected, by means of a drive device (not shown), for example a motor, in particular an electric motor. The shaft 13 has a central axis extending in the radial direction. The screw rotor 14 has a screw rotor center axis extending along an axial direction.

Due to the pressure difference between the pressure of the gas on the suction side (pressure on a suction side of the screw rotor 14, ie at or in the area of the suction gas inlet 12) and the higher pressure on the pressure side (pressure on a pressure side of the screw rotor 14, ie on or in the area of the pressure gas outlet 16) creates an axial force that Screw rotor 14 pushes in the direction of the suction gas inlet 12 or towards the suction side. This direction is marked by an arrow A and is also referred to simply as direction A in the following.

The shaft 13 and thus the screw rotor 14 is mounted on the suction side by a first radial bearing 18 and on the pressure side by a second radial bearing 20 in a housing 22 enclosing the screw compressor 10 . Furthermore, the screw rotor 14 is also mounted on the pressure side by an axial bearing 24 in the housing 22 . In the embodiment described, the axial bearing 24 is designed as a roller bearing, more precisely in the form of a plurality of angular contact ball bearings which are directed towards one another. In alternative embodiments, usually one or more deep groove or angular contact ball bearings (or also combinations thereof, which in particular are directed towards one another) but also plain bearings are conceivable as the axial bearing 24 .

When the screw compressor 10 is in operation, an axial force acts which urges or pushes the screw rotor 14 in direction A. The axial force is taken up directly by the axial bearing 24 and dissipated via the housing 22 .

The screw compressor 10 also has an oil system that provides oil for lubricating the radial bearings 18, 20 and the axial bearing 24, the movement of a control slide 26 for power control of the screw compressor 10 and other tasks. The pressure of the oil is usually approximately as high as the discharge pressure of the compressed gas at the compressed gas outlet 16. In alternative embodiments in which the screw compressor 10 has an oil pump, the pressure of the oil can also be higher.

The screw compressor 10 also has a balancing piston 28 which is arranged on the pressure side of the screw compressor 10 in a balancing piston bushing 30 which is arranged in a recess 32 in the housing 22 which corresponds thereto. The compensating piston 28 has a compensating piston center axis and is arranged to be displaceable along the compensating piston center axis and is designed to generate a counterforce to a force occurring on the screw rotor 14 directed in the axial direction. The balancing piston 28 is on the shaft 13 of the screw rotor 14 attached, more precisely connected in a non-positive manner to the shaft 13 .

Like the shaft 13, the compensating piston 28 also extends in the axial direction and, viewed in the axial direction, has a first compensating piston section 28a with a first outside diameter dl extending in the radial direction and a second compensating piston section 28b with a second outside diameter extending in the radial direction d2, where dl is less than d2. The first balance piston section 28a and the second balance piston section 28b are arranged one after the other in the axial direction. The first compensating piston section 28a has a first effective surface 29a which can be pressurized selectively (a more detailed explanation of the structural details in this regard follows below) with the pressurized oil, and the second compensating piston section 28b has a second effective surface 29b which can be selectively (a more precise explanation of the design details for this also follows below) can be pressurized with the pressurized oil. The first active surface 29a and the second active surface 29b have different sizes. A counterforce can be generated in one of four force levels by selectively applying oil pressure to the first active surface 29a or the second active surface 29b or both active surfaces 29a, 29b or neither of the two active surfaces 29a, 29b.

In alternative embodiments, it is also conceivable that the first active surface 29a and the second active surface 29b have the same size, i.e. have the same area.

The first active surface 29a is designed like a circular ring and has a first active surface inner diameter dli and the first active surface outer diameter the outer diameter dl. The second active surface 29b is designed in the manner of a circular ring and has a second active surface inner diameter d2i and, as the second active surface outer diameter, the outer diameter d2. The relation applies: dli < dl = d2i < d2.

The compensating piston bushing 30 has, also viewed in the axial direction, a first compensating piston bushing section 30a with a first inner diameter dl* and a second compensating piston bushing section 30b with a second inner diameter d2*. It should be noted that the outer diameter dl designates the furthest or largest radial extent of the first balance piston section 28a and that the outside diameter d2 designates the furthest or largest radial extent of the second balance piston section 28b.

The first compensating piston bushing section 30a is provided to receive the first compensating piston section 28a, the second compensating piston bushing section 30b is provided to receive the second compensating piston section 28b. The diameters dl and dl* are formed such that a first gap (in the radial direction) is formed between them. In other words, dl* is greater than dl by the extent of a first (radial) gap. The diameters d2 and d2* are formed such that a second gap (in the radial direction) is formed between them. In other words, d2* is greater than d2 by the extent of a second (radial) gap. In the embodiment described, the gap width or the gap dimension in the radial direction is approximately 200 pm for both the first gap and the second gap, i.e. for the diameters dl and dl* as well as the diameters d2 and d2* (i.e. 200 pm is more usual in the frame Tolerances in manufacture and measurement of the gap size). Expressed mathematically, this results in the following relations: dl* = dl + 200pm and d2* = d2 + 200pm.

A first gap seal 32 is thus formed or arranged between the first compensating piston section 28a and the first balancing piston bushing section 30a and a second gap seal 34 is formed or arranged between the second compensating piston section 28b and the first compensating piston bushing section 30b. The first gap seal 32 and the second gap seal 34 are each designed as a one-sided labyrinth seal.

For this purpose, groove-shaped recesses are formed on the radial outer sides of first compensating piston section 28a and second compensating piston section 28b, so that a first labyrinth seal 32 is formed or arranged between first compensating piston section 28a and first compensating piston bushing section 30a, and between second compensating piston section 28b and second compensating piston bushing section 30b a second labyrinth seal 34 is formed or arranged.

In alternative embodiments, gap dimensions of 50 μm to 400 μm, in particular from 100 μm to 300 μm, furthermore in particular from 150 μm to 250 μm, are provided. It remains to be noted that all diameters dl, dl*, d2 and d2* as well as the gap dimensions extend in the radial direction, as already mentioned at least in part above. As an alternative to a one-sided labyrinth seal, simple gap seals (in particular for small gap dimensions) or two-sided labyrinth seals, as well as contact seals, for example radial shaft seals, lip seals and the like are also conceivable.

The screw compressor 10 further includes a first feed or port 36 associated with the first balance piston portion 28a and a second feed or port 38 associated with the second balance piston portion 28b. Pressurized oil can be fed via the first port 36 into a first space 40 associated with the first balance piston portion 28a and at least partially bounded by the first balance piston portion 28a and the first balance piston sleeve portion 30a. The pressurized oil can also be fed via the second port 38 into a second space 42 associated with the second balance piston section 28b and at least partially delimited by the second balance piston section 28b and the second balance piston bushing section 30b. By feeding pressurized oil into the first space 40 and/or the second space 42, the pressure of the oil creates a force in the direction of the pressure side, which corresponds to the axial force of the screw rotor 14, which is generated by gas force in the direction of the suction side. opposed and thus the load to be absorbed by the thrust bearing 24 is reduced.

It remains to be mentioned that the thrust bearing 24 is supported by disk springs 44 against a force directed in the direction of the pressure side. The thrust bearings 24 are preloaded securely and without play by the disc springs 44, with relatively large dimensional tolerances of the individual components being accepted and individual components, for example roller bearings, in maintenance work can be exchanged very easily. The force of the plate springs 44 can be selected to be lower than in the prior art due to the design of the compensating piston and the possibilities of applying pressure to it selectively (via two feeds, pressure can be applied to different piston diameters in each case). The corresponding construction and its operation will be described in more detail below.

As discussed above, the screw compressor 10 of the present invention includes a double balance piston, i.e., a balance piston 28 having two sections 28a, 28b of different radial diameters, as shown in FIG. For this purpose, the first compensating piston section 28a can be supplied with oil, which is under pressure, via a first oil connection 36 . For this purpose, the second compensating piston section 28b can be supplied with oil that is under pressure via a second oil connection 38 . The oil under pressure pressurizes in particular a surface extending in the radial direction and facing the suction side, which then provides the force in direction B, i.e. towards the pressure side. By switching on and off or opening and closing the two oil connections 36 and 38, which are equipped with appropriate valves for this purpose, the force generated by the double compensating piston 28 can be adjusted in 4 stages as required.

As already shown, the gas pressure on the screw rotor 14 causes an axial thrust before and after compression, i.e. it generates a force directed in the axial direction, which acts in the direction of the suction side. This axial force is diverted to the axial bearing 24 via a tension washer 44 via the compensating piston bushing 30 into the housing 22 .

Oil, which has approximately the same pressure as the compression end pressure, is introduced via the first oil connection 36 and/or the second oil connection 38 into the chambers 40 and 42 of the screw compressor 10 assigned to them and connected to them and causes a force on the compensating piston 28 , which acts in the direction of the pressure side. This force is passed from the compensating piston 28 to the inner rings 46 of the axial bearing 24 via their rolling elements 48 in the outer rings 50 which are in operative engagement with the disk springs 44 . The disk springs 44 which press against the outer rings 50 are supported on a cover 52 of the housing 22 . As mentioned above, the compensating piston 28 is stepped in two diameters. The diameters are preferably designed as a labyrinth. Corresponding to this is the compensating piston bushing 30, which also has at least two diameters which run without contact, but at a minimal distance from the compensating piston 28. This creates a sealing effect, so that before or at the first

Compensating piston section 28a and/or the second compensating piston section 28b can build up an oil pressure and only a relatively small amount of oil runs as a loss through the gap. This flow of oil runs through the axial bearing 24, lubricates it and can be sucked off again through an opening 54.

By grading the diameter of the compensating piston 28, i.e. by designing the compensating piston 28 with a first compensating piston section 28a and a second compensating piston section 28b of different radial diameters, a desired force directed towards the pressure side can be generated in several stages. This is explained in more detail below:

In a zeroth stage, also called stage 0 in the following, none of the oil connections is supplied with oil and thus neither the first active surface 29a nor the second active surface 29b is pressurized.

In a first stage, also called stage 1 in the following, the first oil connection 36 is supplied with oil, the pressure of which usually corresponds approximately to the end pressure of the compression, but is at least significantly above the intake pressure of the compression. The oil supply at the second oil port 38 remains blocked or, in an alternative embodiment, is connected to a low pressure point. Thus, the oil pressure, more precisely, the oil, the pressure of which usually corresponds approximately to the end pressure of the compression, acts only on the small-diameter part of the balance piston 28, ie the first balance piston section 28a. Since the area (effective area 29a) delimited or defined by the first compensating piston section 28a is small, the axial force generated by the compensating piston 28 in the direction of the pressure side is also relatively small. In a second stage, also called stage 2 below, the second oil port 38 is supplied with oil, the oil supply at the first oil port 36 remains blocked or, in the alternative embodiment, is connected to a point of low pressure. Thus the oil pressure, or more precisely, the oil, the pressure of which usually corresponds approximately to the end pressure of compression, acts only on the part of the compensating piston 28 with the large diameter, ie the second compensating piston section 28b (effective surface 29b). In sum, a medium-sized area remains, on which oil pressure acts in direction B. This creates a resultant axial force of medium magnitude that acts in direction B.

In a third stage, also called stage 3 below, both the first oil port 36 and the second oil port 38 are supplied with oil. This acts like a combination of stages 1 and 2, which means that the surfaces (effective surface 29a and effective surface 29b) of both compensating piston sections 28a and 28b are subjected to oil pressure, which in both cases is effective in direction B. In total, the axial force induced by the compensating piston in direction B is maximum.

When operating at full load, operation in stage 3 is usually advantageous, since the axial thrust generated by the gas during the compression process is at its maximum here and therefore requires a large balancing piston force to compensate.

When operating at high part load or at full load with relatively high suction pressure, stage 3 operation could cause a balance piston force in excess of the amount of axial thrust produced by the gas in the compression process. Level 2 operation is usually advantageous here.

When operating at low partial load, the axial thrust generated by the gas during the compression process is very small or non-existent. Level 1 operation is usually advantageous here.

There are various alternatives for controlling the described stages with oil that is under pressure: The operating conditions of the screw compressor 10, in particular the pressures of the sucked-in gas, the compressed gas and the oil, are measured by sensors 56, sent to an electronic controller 58 (a mechanical controller is also conceivable in alternative embodiments) and monitored by the latter. This determines the optimal level of the compensating piston pressure and switches on or two directional control valves 60, 62 accordingly, which release the oil supply to the first oil port 36 and/or to the second oil port 38. In other words, the controller 58 is designed to selectively apply the pressurized oil to the first effective surface 29a and the second effective surface 29b.

In order to save the cost of a directional control valve 60, 62, it may be sufficient for a screw compressor 10 with a high, but not very high load, to permanently connect one of the two oil connections 36 or 38 to the oil supply, while the other oil connection (38 or 36) runs through a directional valve 60, 62, controlled by the controller 58, is released or blocked.

As an alternative or in addition to this, the activation of the stages described with oil that is under pressure can be determined by the position of the control slide 26 of the power control. In other words, the control slide 26 is designed to selectively apply the pressurized oil to the first effective surface 29a and/or the second effective surface 29b.

As described above, the optimal level 1, 2 or 3 of the activation of the multi-level balance piston 28 is often determined by whether the compressor 10 is operating at full load, medium or small part load. The load is adjusted in the screw compressor 10 by the control slide 26 of the capacity control, which is pushed to the stop in the suction direction at full load and which, in order to reduce the delivery volume of the screw compressor 10, is increasingly pushed in the pressure direction. Pushed completely in the direction of pressure, the compressor works in minimal partial load.

Fig. 3 shows one way of controlling the oil using the directional control valves 60, 62. In the full-load position, the control slide 26 is pushed all the way in the suction direction (direction a). The two directional control valves 60, 62 are not actuated and are placed in their initial position by respective return springs 64, 68 assigned to them. Oil flows through the first directional control valve 60 to the second oil port 38 on the balance piston. Furthermore, oil flows through the second directional control valve 62 to the first oil port 36 on the balance piston. Thus, the first and the second oil connection 36 and 38 are active, the compensating piston 28 works with maximum power in stage 3.

When the control slide 26 is moved in the direction of the pressure side (reduction of the delivery rate), it actuates the first directional valve 60 after covering a predetermined distance, which for this purpose has a beveled first directional valve actuating element 70 which is actuated by the control slide 26. As a result, the oil inflow to the first oil connection 36 is shut off. Thus, only the second oil connection 38 is active, the compensating piston 28 works with medium force in stage 2.

If the control slide 26 is moved further in the direction of the pressure side (further reduction in the delivery rate), it also actuates the second directional valve 62 by moving over a second directional valve actuating element 72 assigned to this, namely the second directional valve 62 . The second directional control valve 62 changes the oil flow from the second oil port 38, which is now blocked, to the first oil port 36. Only port 36 is therefore active, and the compensating piston works with the smallest force in stage 1.

In other words, the directional valve actuating elements 70, 72 are integrated in the slide guide recess 76 and are designed to selectively close the first active surface 29a and the second active surface 29b with the pressurized oil by opening and closing the respectively assigned directional valve 60, 62 pressurize The directional valve actuating elements 70, 72 can be actuated mechanically, extend at least partially into the slide guide recess 76 of the control slide 26 and can be actuated by the control slide (26) in order, as mentioned above, to open and close the respectively assigned directional valve 60, 62 to apply the pressurized oil selectively to the first active surface 29a and the second active surface 29b. Balancing piston pressure control could be applied not only to a multi-stage, but also to single balancing piston compressors, according to the current state of the art.

The simple compensating piston has a connection for oil with which it is supplied. This connection can be switched via a directional valve as described, whereby the compensating piston force is switched on or off in one step.

Fig. 4 shows a basic possibility of controlling the oil by a directional control valve 62'.

This circuit is also conceivable with a multi-stage compensating piston 28, as is implemented in the subject of the present description of the figures, in that one connection (e.g. the first oil connection 36) is permanently supplied with oil and another connection (e.g. the second oil connection 38) by a valve , For example, solenoid valve 74 is either supplied with oil or shut off.

The compensating piston force can be switched in 2 stages. Maximum balance piston force at full load and reduced balance piston force at part load of screw compressor 10.

In an alternative embodiment, the control slide 26 can be designed in such a way that it itself takes over the function of one or both directional control valves 60, 62 and no further components are required.

The control slide 26 moves in a slide guide recess 76 in the housing 22, whereby it comes to a stop on this slide guide recess 76 when moving in the direction of the intake side (direction a), this position is the full load position. When the control slide 26 moves in the direction of the pressure side (direction b), the output of the compressor is reduced more and more until the control slide 26 comes to a stop in the minimum partial load position. This stop is shown in a conventional version of the power control by the piston, on the right in Fig. 6. formed which drives against the wall of the hydraulic cylinder.

Fig. 5 shows the screw compressor 10, the control slide 26 is in the full load position and Fig. 6 shows the same in the small part load position. The control slide 26, which is guided in a corresponding slide guide recess 76 in the housing 22, as in the embodiments shown in FIGS. 1 to 4, has a recess 78 in the form of a pocket in an area directed towards the housing 22, which control slide 26 is arranged in such a way that it is sealed off from its surroundings in the best possible way by the control slide 26 and the housing 22, i.e. that the pocket is completely covered by the housing 22 and there is thus at most one gap between the areas of the control slide 26 that delimit the pocket and the Housing 22 is formed, which acts as a gap seal, since oil is present in this area.

Furthermore, a first oil line recess 80 and a second oil line recess 82 are arranged in the housing 22, which in the present embodiment are present in the form of channels which are already specifically produced during the production of the housing, for example in a casting or injection molding process. Alternatively, bores that are made, for example, after the manufacture of the housing 22, are conceivable.

The first oil route recess 80 and a second oil route recess 82 open into the slider guide recess 76 in a first oil route recess mouth 84 forming the mouth of the first oil route recess 80 and a second oil route recess mouth 86 forming the mouth of the second oil route recess 82. The first oil passage notch opening 84 and the second oil passage notch opening 86 are located in a region of the spool guide notch 76 that is covered by the spool 26 or notch 78 (pocket) at any position of the spool 26 .

When, as shown in FIG. 5, the control valve 26 is in the full load position, the cutout 78 connects the first oil passage cutout 80 and the second oil passage cutout 82 so that the balance piston 28 is supplied with oil at the corresponding port. The pocket (recess 78) is surrounded all around by sufficient material of the control slide 26 so that a sealing effect is achieved in combination with the guide bore (slide guide recess 76).

The pocket (recess 78) is arranged so far in the direction of the pressure side of the compressor that the bores that feed or drain oil into the pocket are also covered by the control slide 26 when it is maximally in the direction of the pressure side, i.e. in the smallest partial load is shifted.

By moving the control spool 26 toward the pressure side, the pocket (recess 78) no longer connects the first oil passage recess 80 and the second oil passage recess 82, so that the balance piston 28 is not supplied with oil.

Furthermore, the control slide covers the first oil line recess 80 and the second oil line recess 82 so that they are sealed from the interior of the compressor and neither oil from the openings of the same can get into the space from which gas flows into the screw rotor, nor gas into the openings of the first oil line recess 80 and the second oil line recess 82 can reach.

In other words, this can be briefly expressed in such a way that in one possible embodiment, a plurality of oil line recesses 80, 82 open into the slide guide recess 76, which are designed to selectively apply the pressurized oil to the first active surface 29a and the second active surface 29b , the control spool having a recess 78 and being reciprocally disposed in the spool guide recess 76 to positions in which the control spool 26 closes all of the oil line recesses 80, 82, in which one or both of the oil line recesses 80, 82 fit into the recess 78 of the control spool 26 flow. As an alternative to this, only one oil line recess can be provided in order to be able to act selectively on one active surface, with the other active surface either being permanently pressurized with oil or being pressurized with oil via another mechanism, for example a valve controlled by the controller 58 . In summary, it can be stated that the subject matter of the present application has the following features, among others:

1. Screw compressor, which has a balance piston, which generates a counter force to the axial force occurring at the screw rotor, which is designed with at least two diameters, both of which are arranged in the housing and sealed to this, which is usually by a gap or preferably a labyrinth takes place, whereby the gap of approx. 0.1 ... 0.3mm is selected in such a way that the piston does not come into contact with the housing.

2. Screw compressor according to 1., wherein the compensating piston or compensating piston sections are each surrounded on the side by a chamber which is supplied via a connection with oil that is under pressure, so that the pressure force generated by the piston surface occurs on the screw rotor counteracting axial force.

3. Screw compressor according to 2., wherein the diameter of the compensating piston sections is selected in such a way that by activating the oil supply into none, the first, the second chamber or into both chambers, solely due to the effective areas of the compensating piston or the respective compensating piston section, differs strong axial forces are caused, whereby the pressure of the supplying oil can remain approximately constant, so no regulation of the oil pressure is required.

4. Screw compressor according to 3., wherein the oil supply to the first or the second chamber is blocked or activated in a targeted manner, so that the axial force acting on the rotor through the compressed gas is reduced by the force generated by the compensating piston, but it is avoided that a Reversal of the direction of the resulting force occurs.

5. Screw compressor according to 4., wherein the targeted shutting off or activating the oil supply into the chamber or chambers is determined by an electronic control. 6. Screw compressor according to 4. with a capacity control with a control slide, the targeted shutting off or activating the oil supply to the chamber or chambers being effected by the movement of the control slide

7. Screw compressor according to 6., wherein one or more directional control valves are integrated in a guideway of the control slide, which direct the flow of the oil to at least one chamber of the compensating piston. These directional valves are switched by mechanical actuating elements (directional valve actuating elements) which are arranged at a suitable point in such a way that the control slide either overruns them during movement or the control slide during movement releases an actuating element (directional valve actuating element) that has previously been passed.

8. Screw compressor according to 7., wherein the function of the directional control valve or valves is fulfilled directly by the control slide, in that one or more pockets are introduced into it in such a way that they are in one position, or in a certain area of the travel, of the control slide once is connected to a first passage in the housing which provides oil and is connected by a second passage in the housing which is connected to a chamber of the balancing piston so that communication to the chamber of the balancing piston is via the second passage, the pocket and the first channel is made. By moving the control slide a certain distance from this position, the pocket of the control slide covers only one or neither of the two channels in the housing, so that the oil supply to the balancing piston chamber is interrupted.

Although the invention is described on the basis of embodiments with fixed combinations of features, it also includes other conceivable advantageous combinations, as specified in particular, but not exhaustively, by the dependent claims. All features disclosed in the application documents are claimed to be essential to the invention insofar as they are new compared to the prior art, either individually or in combination.

Claims

patent claims

1. Screw compressor (10) with at least one suction gas inlet (12), at least one screw rotor (14) having a screw rotor central axis extending along an axial direction, and at least one pressure gas outlet (16), and a screw compressor (10 ) at least partially enclosing the housing (22) and with a compensating piston (28) which has a compensating piston center axis, which is arranged to be displaceable along the compensating piston center axis and which is designed to counteract an axial force occurring on the screw rotor (14). direction directed force, characterized in that the compensating piston (28) has at least a first compensating piston section (28a) and a second compensating piston section (28b), wherein the first compensating piston section (28a) has a first effective surface (29a) which selectively, in particular via a first valve (60, 74), with a pressurized fluid, esp especial oil, can be pressurized, and wherein the second compensating piston section (28b) has a second active surface (29b) which is selectively, in particular via a second valve (62, 74), with a pressurized fluid, in particular oil, with pressure is acted upon.

2. Screw compressor (10) according to claim 1, characterized in that the balance piston (28) extends in the axial direction, the first balance piston section (28a) and the second balance piston section (28b) being arranged one after the other in the axial direction, and the first The balance piston portion (28a) has a circular outer periphery in its radial direction with a diameter dl and the second balance piston portion (28b) has a circular outer periphery in its radial direction with a diameter d2, where dl is smaller than d2.

3. Screw compressor (10) according to one of the preceding claims, characterized in that the compensating piston (28) is mounted in a compensating piston bush (30) which is at least partially arranged in the housing (22) of the screw compressor (10), in particular as a recess in this is trained.

4. Screw compressor (10) according to Claim 3, characterized in that a labyrinth seal, in particular a one-sided labyrinth seal, is provided between at least one compensating piston section (28a, 28b), in particular between each compensating piston section (28a, 28b) and the compensating piston bushing (30). is arranged.

5. Screw compressor (10) according to claim 3 or 4, characterized in that a contacting seal, in particular a radial shaft seal or a Lip seal is arranged.

6. Screw compressor (10) according to one of the preceding claims, characterized in that the screw compressor (10) has a controller, in particular an electronic controller (58) which is designed to the first active surface (29a) and / or the second Active surface (29b) to act selectively with the pressurized fluid, in particular oil.

7. Screw compressor (10) according to any one of the preceding claims, characterized in that the first active surface (29a) is circular and has a first active surface inner diameter dli and a first active surface outer diameter is the outer diameter dl and that the second active surface (29b) is circular and has a second active surface inner diameter d2i and a second active surface -Outer diameter has the outer diameter d2.

8. Screw compressor (10) according to claim 7, characterized in that the relation dli<dl=d2i<d2 applies.

9. Screw compressor (10) according to one of the preceding claims, characterized in that the screw compressor (10) has a control slide (26), which is arranged in a slide guide recess (76) so that it can be moved back and forth, for controlling the output of the screw compressor (10), which is designed to apply the pressurized fluid, in particular oil, to the first effective surface (29a) and/or the second effective surface (29b), in particular to apply it selectively.

10. Screw compressor (10) according to Claim 9, characterized in that one or more directional valve actuating elements (70, 72) are at least partially integrated in the slide guide recess (76) and are designed to open and close a respectively assigned directional valve (60 , 62) to apply, in particular to selectively apply, the first active surface (29a) and/or the second active surface (29b) to the pressurized fluid, in particular oil.

11. Screw compressor (10) according to claim 10, characterized in that the one or more directional valve actuating elements (70, 72) are mechanically operable, extend at least partially into the spool guide recess (76) of the control spool (26) and through the control spool ( 26) are operable.

12. Screw compressor (10) according to one of claims 9 to 11, characterized in that one or more oil line recesses (80, 82) open into the slide guide recess (76), which are designed to surround the first effective surface (29a) and/or to apply the pressurized fluid, in particular oil, to the second active surface (29b), in particular to apply it selectively, the control slide having a recess (78) and being arranged in the slide guide recess (76) so that it can be moved back and forth in positions in in which the control slide (26) closes all oil line recesses (80, 82), in which one or more oil line recesses (80, 82) open into the recess (78) of the control slide (26) and in which all oil line recesses (80, 82) flow into the Recess (78) of the control slide (26) open.

13. Screw compressor (10) according to one of the preceding claims, characterized in that the pressurized fluid with which the first effective surface (29a) can be pressurized and the pressurized fluid with which the second effective surface (29b) can be pressurized, have an approximately constant pressure during operation of the screw compressor (10), in particular approximately the same pressure.

14. Screw compressor (10) according to one of the preceding claims, characterized in that the first effective surface (29a) and the second effective surface (29b) have different sizes.