WO1991017345A1

WO1991017345A1 - Rotary screw compressor with thrust balancing means

Info

Publication number: WO1991017345A1
Application number: PCT/SE1991/000309
Authority: WO
Inventors: Frits SÖDERLUND
Original assignee: Svenska Rotor Maskiner AB
Current assignee: Svenska Rotor Maskiner AB
Priority date: 1990-05-08
Filing date: 1991-04-30
Publication date: 1991-11-14
Anticipated expiration: 1992-11-08
Also published as: SE9001640L; SE9001640D0; SE466166B

Abstract

In a rotary screw compressor with capacity regulating means (14) the axial gas forces on the rotors (12) are balanced by hydraulic thrust balancing means (18). When the compressor is working at reduced capacity below a certain level the thrust balancing force is inactivated. This is accomplished in that the pressure oil supplied to the thrust balancing means (18) is drained by opening a cut-off valve (28) controlling a drainage channel (26) connected to the oil supply channel (24). According to the invention the cut-off valve (28) is directly governed by the capacity regulating means (14) so that when these means are in a position representing a capacity below a certain level, typically about 75 % of full capacity, the drainage channel (26) is connected to a low pressure source. The capacity regulating means can preferably act as the cut-off valve which in the open position connects the drainage channel (26) to the low pressure port (44) of the compressor through the compressor working space (13).

Description

ROTARY SCREW COMPRESSOR WITH THRUST BALANCING MEANS

The present invention relates to a rotary screw compressor having a pair of intermeshing rotors in a work¬ ing space and axial slide means regulating the capacity of the compressor in a valve boring facing the working space and extending axially along the same, at least one of the rotors being provided with hydraulic or pneumatic thrust balancing means including a first pressure chamber facing a pressure surface, a supply channel for the supply of pressure fluid to said first pressure chamber and a drainage channel connected to said supply channel, said drainage channel being controlled by a cut-off valve having an open position connecting said drainage channel to a low pressure source and a closed position.

A compressor having such thrust balancing means is known from US 3,388,854. In this known device the cut-off valve is governed by means sensing the outlet or interstage pressure of the compressor.

When a compressor of the kind in question is oeprating, the gas exerts an axial load on the rotors in the direction from the outlet end to the inlet end of the compressor.

This load is taken up by thrust bearings. In many applica¬ tions of the compressor, e.g. when used in heat pump systems an appreciable running life of these bearings is desired, often more than 50.000 hours. In order to attain this the compressor can be provided with a hydraulic thrust balancing device acting on the rotors in a direction contrary to the direction of the axial gas forces. When the compressor, however, is capacity regulated the thrust balancing fore? required at full capacity will exceed the gas forces when running at reduced capacity. Therefore the thrust balancing force should be in-activated or reduced when running at lower capacity in order to avoid over- compensation of the axial load. The object of the present invention is to find a simple and reliable way to adapt the thrust balancing force to the capacity of the compressor.

This has according to the invention been achieved in that the position of the slide valve means governs said cut-off valve so that the cut-off valve is open when the slide valve means are in a position corresponding to a capacity at or below a certain level and is closed when the slide valve means are in a position corresponding to a capacity above said level.

Advantageous embodiments of the invention are specified in the dependent claims.

The invention will be explained through the following detailed description of two preferred embodiments of the compressor and with reference to the accompanying figures of which

Figure 1 is a partial section through a compressor according to a first embodiment of the invention and

Figure 2 is a section similar to that of figure 2 but according to a second embodiment of the invention.

In figure 1 the upper part is a longitudinal section through one of the rotors 12 of a rotary screw compressor according to the invention. The lower part of the figure shows the slide valve means 14 regulating the capacity and the internal volume ratio of the compressor. The screw rotor 12 intermesh with a second screw rotor (not shown) to form chevron-shaped compression chambers in a working space 13 limited by a low pressure end wall 15, a high pressure end wall 17, a barrel wall 19 and slide valve means 14. The working space 13 has in a transversal section the shape of two intersecting borings formed by the barrel wall 19 and those surfaces of the slide valve means 14 which face the working space 13. The slide valve means 14 include a slide stop member 32 and a slide valve member 34. The slide stop member 32 and the slide valve member 34 can independently of each other be displaced axially in an axial boring 16 intersecting the working space 13 and are guided by the wall 42 of the boring 16. In the figure, the slide valve means 14 are in a position representing full capacity. For attaining reduced capacity the slide stop member 32 is moved leftwards in the figure or the slide valve member 34 is moved rightwards.

Thereby the slide stop member 32 and the slide valve member 34 will become spaced apart and the position of the rear end surface 38 of the slide valve member 34 will determine the moment when each compression chamber is closed off from communication with the low pressure port 44 of the compressor. This way of regulating the capacity is well-known in the rotary screw compressor technology and therefore requires no further explanation.

Each rotor 12 is provided with shaft journals 21, 23 for the journalling of the rotors. The shaft journal 23 at the high pressure end is journalled in journal roller bearing 25, thrust ball bearings 27, 29, and thrust balancing bearing 31. Hydraulic thrust balancing means 18 act on the thrust balancing bearing 31. These means include a pressure chamber 20 facing a pressure surface 22 on an actuating member 33 abutting the outer ring 35 of the thrust balancing bearing 31. Oil under pressure is supplied to the pressure chamber 20 through a supply channel 24. From the supply channel 24 a drainage channel 26 leads to an orifice 30 in the wall 42 of the valve boring 16. The oil supplied through the channel 24 can preferably be pressurized by the outlet pressure of the compressor. Although a liquid is used as pressure medium in the described embodiment it should be understood that gaseous pressure medium can be used as well.

In operation of the compressor the compressed gas exerts an axial load on eadch rotor 12 in a direction from the high pressure end to the low pressure end of the compressor, i.e. leftwards in the figure as indicated by the arrow F. Through the thrust balancing means 18 a force can be applied to the outer ring 35 of the thrust balancing bearing 31, which transmits this force to the rotor 12. The thrust balancing force is directed contrary to the axial gas forces, i.e. rightwards in the figure. Thereby the axial load on the thrust ball bearings 27, 29 will be reduced so that a longer running life of these bearings is attained.

When the compressor is running at reduced capacity, the axial gas forces will be smaller. There is therefore no longer any need to balance these forces when the capacity is below a certain level. In contrary the thrust balancing force under such condition might over-compensate the axial gas forces so that the resulting force on the rotor 12 momentary can be directed rightwards with the risk that the rotor 12 contacts the high pressure end wall 17 as a consequence. In order to avoid this the compressor is provided with means to in-activate the thrust balancing means 18 when running at a capacity below a certain level. At reduced capacity the slide stop member 32 and the slide valve member 34 are spaced apart. Thereby communication is opened between the low pressure port 44 and the orifice 30 in the wall 42 of the valve boring 16 through the working space 13 and the interspace between the slide stop member 32 and the slide valve member 34. This, however, occurs only if the slide valve member 34 is so located that it does not cover this orifice 30, i.e. when the rear end surface 38 of the slide stop member 34 is in a position to the right of the orifice 30, which represents a capacity below said certain level. If the slide stop member 32 is moved leftwards from the position shown in the figure said communication is established because the orifice 30 then directly faces the interspace. If the position of the slide stop member 32 is remained and the slide valve member 34 is moved rightwards the orifice 30 communicates with the interspace through the recess 40 in the front end surface 36 of the slide stop member 32. When communication between the orifice 30 and the low pressure port 44 is open, oil supplied to the pressure chamber 20 through the supply channel 24 will be drained to the low pressure port 44 through the drainage channel 26, the orifice 30, the interspace between the slide stop member 32 and the slide valve member 34 and through the working space 13 of the compressor. Thereby the pressure acting on the pressure surface 22 is relieved and as a consequence the thrust balancing force is in-activated. The axial location of the orifice 30 can easily be calculated in order to correspond to the capacity condition where there is no need for thrust balancing. At reduced capacity the position of the rear end surface 38 of the slide valve member 34 determines the capacity. The more to the right this end surface is positioned in the figure, the lower is the capacity. The orifice thus is located adjacent the rear end surface 38 of the slide valve member 34 to the left side (in the figure) thereof when the slide valve member 34 is in a position representing the capacity at or below which the thrust balancing means 18 shall be in-activated. In typical cases this lies somewhere in the range from 60 % to 85 % , normally about 75 % of full capacity.

With the described device the thrust balancing force can be adapted to the capacity of the compressor in a very simple and reliable way. No extra means for governing this is required since it is governed directly by the capacity regulating means.

Figure 2 illustrates in a view similar to that of figure 1 an alternative embodiment of the invention, in which there are provided thrust balancing means at both ends of the compressor. The thrust balancing means 118 at the high pressure side are of similar construction as those 18 of figure 1. The supply channel 124, however, has no drainage channel so that these thrust balancing means 118 are active under all working conditions. The thrust balancing means 58 at the low pressure end have a pressure chamber 50 facing the end surface 52 of the shaft journal 21. The thrust balancing force thus acts directly on the rotor when oil under pressure is supplied to the pressure chamber 50 through the supply channel 54. The shaft journal 21 is journalled in a roller bearing 46, and there is a seal 48 between the roller bearing 46 and the pressure chamber 50. A drainage channel 56 connects the supply channel 54 to the orifice 30 in the wall 42 of the valve boring. At reduced capacity the pressure in the pressure chamber 50 is relieved since the drainage channel undere such conditions communicate with the low pressure prot 44 through the interspace between the slide stop member 32 and the slide valve member 34 and through the working space 13 of the compressor. This is accomplished in the same manner as described in relation to f gure 1. As can be understood from above both the thrust balancing means 58 and 118 are active at full capacity, whereas at reduced capacity the thrust balancing means 118 at the high pressure end are active and the thrust balancing means 58 at the low pressure end are in-active. Through this embodiment of the invention it is thus possible to maintain a certain thrust balancing force also at reduced capacity but at a lower level corresponding to the smaller axial gas forces.

If a larger balancing force is required than what can be attained through the pressure surface 52 at the end of the shaft journal 21, a piston having a larger pressure surface can be fixed to the end of the shaft journal. It is possible also to provide the thrust balancing means 118 at the high pressure end with a drainage channel connected to a second orifice in the wall of the valve boring, through which this thrust balancing means can be in-activated at a still lower capacity level.

Claims

1. Rotary screw compressor having a pair of intermeshing rotors (12) in a working space (13) and axial slide valve means (14) regulating the capacity of the compressor in a valve boring (16) facing the working space and extending axially along the same, at least one of the rotors (12) being provided with hydraulic or pneumatic thrust balancing means (18; 58, 118) including a first pressure chamber (20; 50) facing a pressure surface (22; 52), a supply channel (24; 54) for the supply of pressure fluid to said first pressure chamber (20; 50) and a drainage channel (26; 56) connected to said supply channel (24; 54), said drainage channel (26; 56) being controlled by a cut-off valve (28) having an open position connecting said supply channel to a low pressure source (44) and a closed posi- tion, c h a r a c t e r i z e d i n that the position of said slide valve means (14) governs said cut-off valve (28) so that said cut-off valve (28) is open when said slide valve means (14) are in a position corresponding to a capacity at or below a certain level and is closed when said slide valve means (14) are in a position corresponding to a capacity above said level.

2. Compressor according to claim 1 , wherein said drainage channel ends in said boring (16) through an orifice (30) being so located that said slide valve means (14) allow communication between said orifice (30) and the low pressure port of the compressor through said boring (16) and said working space (13) when said slide valve means (14) are in said position corresponding to a capacity below said level and prevent such communication when said slide valve means (14) are in a position corresponding to a capacity above said level, said slide valve menas (14) thereby forming said cut-off valve (28).

3. Compressor according to claim 2, wherein said slide valve means (14) include a slide stop member (32) having a front end surface (36) and a slide valve member (34) having a rear end surface (38), which members (32, 34) are dis- placeable in said boring (16) independently of each other, said end surfaces (36, 68) abutting each other at full capacity and being spaced apart at reduced capacity whereby said orifice (30) communicates with the low pressure port (44) of the compressor through the interspace between said end surfaces.

4. Compressor according to claim 3, wherein the position of said rear end surface (38), when said end surfaces (36, 38) being spaced apart, determines the capacity of the compressor and said orifice (30) is located adjacent said rear end surface (38) but uncovered by said slide valve member (34) when said rear end surface (38) is in a position corresponding to a capacity at said level.

5. Compressor according to claim 4, wherein said slide stop member (32) is provided with a recess (40) at the edge of said front end surface adjacent the wall (42) of said boring (16), said recess (40) being axially aligned with the location of said orifice (30) .

6. A compressor according to any of claims 1 to 5, wherein said thrust balancing means (58; 118) further include a second pressure chamber (120) facing a second pressure surface (122) and a second supply channel (124) for the supply of pressure fluid to said second pressure chamber (120) .

7. A compressor according to any of claims 1 to 5, wherein said capacity level is within a range between 60 to 85 % of full capacity.