GB2089432A

GB2089432A - An Oil-injected Meshing-screw Gas-compressor

Info

Publication number: GB2089432A
Application number: GB8137422A
Authority: GB
Original assignee: Sullair Technology AB
Current assignee: Sullair Technology AB
Priority date: 1980-12-12
Filing date: 1981-12-11
Publication date: 1982-06-23
Also published as: SE8008760L; USRE32055E; DE3148579A1; GB2089432B; JPS57122188A; SE424760B

Abstract

At the high-pressure end of the compressor oil under pressure is supplied to bearing chambers (28a, 28b) whence some of the oil flows into the rotor chamber through gaps (32a, 32b) between the rotor shafts and the housing (11). Oil also flows through a passage (42) in the female rotor from the bearing chamber (28b) into a pressure chamber (38) at the low-pressure end of the said rotor for balancing the axial thrust caused by the oil in the bearing chamber (28b). <IMAGE>

Description

SPECIFICATION A Method at an Oil-injected Screw-compressor This invention relates to a method at oilinjected screw-compressors for balancing axial forces at at least one of the compressor rotors and for sealing the gaps between the rotor shafts and the rotor housing in order to prevent leakage from the compression space of the screw compressor through these gaps.

It was found very difficult at screwcompressors, at high pressure differences over the compressor to obtain a sufficient service life for the axial bearings, in which the shafts of the rotors are supported.

In order to increase the service life of the axial bearings, it was proposed to use balancing pistons such as shown, for example, in US-PS 3 161 349. This arrangement, however, has proved to involve great disadvantages, a.o. a substantial oil leakage over the outer diameter of the balancing piston. It was, therefore, difficult to maintain the desired balancing pressure, and the leakage has caused losses in the efficiency degree of the compressor.The way in which the balancing piston had been built-in in the compressor, viz. on the rotor shaft located on the high-pressure side closest to the compression space, gives rise to the further disadvantage, that the balancing pressure is transferred along the rotor shaft inward to the rotor body and thereby acts also on the ring area formed by the end surface plane of the rotor body between the shaft diameter and the core diameter of the rotor. The force hereby arising on this ring area acts in a direction opposed to the desired balancing force and, thus, reduces the desired balancing considerably.One object of the invention is to provide a method of axially balancing the rotor shaft, whereby the aforesaid disadvantages of conventional arrangements are eliminated, and which method especially can be applied to screwcompressors operating with high pressure differences, of the magnitude 2 MPa and higher, over the compressor.

A further object of the invention is to provide a method of the aforesaid kind, at which also sealing against leakage from the compression space of the compressor along the shafts of the rotors is obtained.

A still further object of the invention is at a method of the aforesaid kind to simultaneously effect lubrication and cooling of the bearings provided for the respective shaft. The improved balancing in combination with the lubrication and cooling possibility have been achieved in that the invention has been given the characterizing features defined in the attached claims.

The invention is described in greater detail in the following by way of an embodiment thereof and with reference to the accompanying drawing, which is a horizontal section through a screwcompressor provided with an arrangement according to the invention.

The rotor housing 11 of the screw-compressor 10 includes a compression space in the form of two rotor barrels forming two intersecting bores, with a low-pressure port 12 at one end 13 and a high-pressure port (not shown) at the other end 14. In the rotor barrels, two meshing rotors, viz.

one screw rotor 1 5 and one slide rotor 16, are mounted rotatably.

On the low-pressure side of the compressor, a radial bearing 17, preferably of roller bearing type, and an axial bearing 18, preferably of angular contact ball bearing type, are built-in for supporting the screw rotor 1 5. Outside said bearing package a balancing piston 1 9 is located a the rotor shaft end 20 for balancing the main part of the axial forces acting on the high-pressure end of the screw rotor 15. Said balancing piston 1 9 is located in a pressure space 21, to which oil under pressure can be supplied from the outside through an oil inlet openign 22. At the outer diameter of the balancing piston a meachanic sealing 23 is located which ensures that a constant pressure of the oil supplied is maintained.In order to effect oil circulation for cooling this sealing 23 and for cooling and lubricating the bearing package 1 7, 18, connections 24 are drilled from the pressure space 21 outside the balancing piston 1 9 into the bearing space at the rotor shaft. The oil can continue to pass from here along gaps 25 between the rotor shaft and the rotor housing into the compresion space for sealing these gaps 25, thereby eliminating leakage from the compression space.

The screw-compressor 10 is driven via the drive shaft 26 of the screw rotor 1 5 which extends outward through the rotor housing 11 on the high-pressure side of said housing and is supported in a radial bearing 27 located in a bearing space 28a. In this bearing space also a mechanic shaft sealing 29 for the drive shaft 26 is provided. The bearing space 28a on the highpressure side of the screw rotor 1 5 is in direct connection with a bearing space 28b on the highpressure side of the slide rotor 16. In the bearing space 28b of the slide rotor a radial bearing 30 is located for supporting the shaft 31 of the slide rotor on the high-pressure side. As can be seen, no special sealings against the compression space are built-in.Oil is suppoied under pressure to the bearing spaces 28a, 28b through a throttling 33, which is adjusted so as to deliver pressure of the magnitude of the arithmetic mean value of the inlet and outlet pressure of the screwcompressor so that an oil flow if ensured through the bearings 27, 30 along the gaps 32a , 32b formed between the rotor shafts 26, 31 and the rotor housing into the compression space, whereby cooling and lubriction of these bearings 27, 30 are obtained and at the same time gas leakage out of the compression space along the rotor shafts 26, 31 is prevented.

For supporting the slide rotor 1 6 on the lowpressure side, a bearing package like the one for the screw rotor 1 5 is built-in the form of a radial bearing 34, preferably of roller bearing type, and an axial bearing 35, preferably of angular contact ball bearing type. Outside this bearing package, a balancing piston 36 is attached by screws one screw 43 indicated in the Figure) to the shaft end 37 of the slide rotor. (A corresponding attachment applies to the balancing piston 1 9 at the shaft end of the screw rotor). The balancing piston 36 is located in a pressure space 38, to which oil under pressure is supplied. At the outer diameter of the balancing piston 36 a mechanic sealing 39 is provided to ensure that a constant pressure of the oil supplied is maintained.In order to effect oil circulation for cooling this sealing 39 as well as for cooling and lubricating the bearing package 34, 35, connections 40 are drilled from the pressure space 38 into the bearing space at the shaft end 37 of the slide rotor. The oil can continue to pass from here along gaps 41 between the rotor shaft and rotor housing into the compression space for sealing these gaps 41 against leakage from the compression space.

For supplying oil to the pressure space 38 on the high-pressure side of the slide rotor 1 6, a connection 42 is drilled axially along the central line of the slide rotor, so that in the bearing space 28b of the slide rotor and the pressure space 38 a common pressure is obtained.

The pressure space 21 for the balancing piston 1 9 of the screw rotor 1 5 is supplied with oil, the pressure of which corresponds to the outlet pressure of the screw-compressor reduced by the pressure drop in the oil cooler and oil filter. The oil pressure in this pressure space 21, thus, is substantially higher than in the corresponding pressure space 38 on the slide rotor side. This higher pressure is desirable in view of the substantially higher axial forces acting on the screw rotor 1 5 compared with those acting on the slide rotor 1 6.

Due to the fact that the screw-compressor is driven from the high-pressure side of the screw rotor, no free shaft journal area does exist here and, thus, no additional axial forces from the bearing space 28a on the high-pressure side are obtained. It was hereby possible to limit the area on the balancing piston of the screw rotor and to design it with the same size as on the slide rotor side. It was hereby possible that both the bearing package and the balancing piston system inclusive of the mechanic sealings at the lowpressure ends on the screw and slide rotor sides could be designed identical.

As regards the dimensioning of the connections 24 and 40 drilled from the presure spaces 21 and, respectively, 38, the hole area obtained by one or more bores is calculated on the basis of available oil pressure difference and oil viscosity, so that a suitable oil amount for cooling and lubricating the bearing package is obtained. This oil amount normally is of the magnitude 5 litres/min per bearing package.

As regards the throttling 33, this is to be calculated so that the oil supply therethrough slightly exceeds the oil amount calculated for the flow in the aforesaid connections. Hereby an oil supply along the gaps 32a, 32b on the highpressure side into the compression space always is obtained. It is, however, an essential and characterizing feature of the invention, that a more accurate dimensioning of this throttling is not required, because an oil supply in excess of the aforesaid minimum amount implies only that the pressure in the bearing spaces 28a, 28b and in the pressure space 38 of the slide rotor increases relatively insignificantly, because due to the higher pressure the oil flow above all along the gaps 32a, 32b on the high-pressure side increases simultaneously.

A further essential and characterizing advantage of the invention is that, if the pressure according to the aforesaid increases in the bearing spaces 28a, 28b, this does not affect the axial forces neither of the screw rotor nor of the slide rotor, because the ingoing shaft 26 is the drive shaft, and the sealing area of the mechanic sealing 29 is located on the same level as the diameter of the rotor shaft extending into the compression space and, therefore the pressure in the bearing space 28a does not yield any axial force, but the axial forces on the rotor shaft 26 substantially completely balance each other in this space 28a.

What has been stated above for the screw rotor 1 5 applies in principle also to the slide rotor 16, viz. that the axial forces substantially are independent of the pressure in the bearing space 28b, because the axial force obtained on the rotor shaft 31 in this space 28b is balanced substantially completely by the axial force obtained on the balancing piston 36 on the opposite side of the slide rotor, due to the fact that, as described above, the same pressure prevails on both sides, and that the pressure area on both sides substantially is of the same size, because the diameter of the rotor shaft 31 on the high-pressure side is substantially the same as the diameter of the balancing piston 36 on the low-pressure side.

The arrangement described above has brought about in an operationally very reliable way an oil circulation system for the cooling and lubrication of bearings and mechanic sealings and an axial balancing system for both rotors, thereby rendering it possible to use simple and inexpensive bearings with a good service life. On the screw rotor where the axial forces are high, the arrangement via the balancing piston yields a balancing force, which increases with increasing counterpressure in the compressor, whereby the bearing force and, thus, the bearing service life substantially are constant. A change of the inlet pressure to the compressor does not affect the axial forces acting on the screw rotor from the balancing piston or bearing space. As regards the slide rotor, the axial forces from the balancing piston and bearing space are not affected, either, by changes in the inlet or outlet pressures of the compressor.

Claims

1. A method at on oil-injected screwcompressor for balancing axial forces at at least one of the rotors of the compressor, for sealing the gaps between the rotor housing and the shafts of the rotors and for cooling and lubricating the bearings of the rotor shafts, where a rotor housing includes a compression space in the form of two rotor barrels defined by two intersecting bores with a low pressure port at one end and a high-pressure port at the other end, and two meshing rotors, viz. a screw rotor and a slide rotor, mounted rotatably in the rotor barrals, which compressor is driven on the shaft of the screw rotor on the high-pressure side, characterized in that at the high-pressure end of the compressor oil under pressure of such size is supplied to bearing spaces at the ends of both rotors, that an oil flow inward to the compression space along the gaps between the rotor shafts and the rotor housing is obtained for sealing against leakage from the compression space, that oil at this pressure is supplied via a connection from the bearing space of the slide rotor to a pressure space at the low-pressure end of the slide rotor for balancing the axial force arising on the shaft end of the slide rotor on the highpressure side due to the oil supplied to the bearing space of the slide rotor.

2. A method as defined in claim 1, characterized in that the oil is supplied to the bearing spaces though a common inlet conduit, and that the two bearing spaces are in direct connection with each other.

3. A method as defined in claim 1 or 2, characterized in that the supply of oil from the bearing space of the slide rotor to the pressure space is effected via a bore extending centrally through the slide rotor shaft.

4. A method as defined in any one of the preceding claims, characterized in that roller bearings provided for supporting the rotor shafts in the bearing spaces are cooled and lubricated by the oil supplied to the bearing spaces.

5. A method as defined in any one of the preceding claims, characterized in that for preventing axial forces from being supplied to the shaft of the screw rotor due tithe oil supplied to the bearing space of the screw rotor, the oil acts on a mechanic sealing located about the drive shaft in such a manner, that the sealed surface has about the same diameter as the diameter of the rotor shaft where the shaft passes through the rotor housing.

6. A method as defined in any one of the preceding claims, characterized in that the pressure of the oil supplied to the bearing spaces and the pressure space of the slide rotor has the magnitude of the arithmetic mean value of the inlet and outlet pressures of the screwcompressor.

7. A method as defined in any one of the preceding claims, characterized in that for axial balancing the screw rotor, oil under pressure is supplied to a pressure space adjacent to the shaft end of the screw rotor on the low-pressure side.

8. A method as defined in claim 7, characterized in that for supporting the rotor shafts on the low-pressure side, the shaft ends of the screw and slide rotors are supported in roller bearings and angular contact ball bearings.

9. A method as defined in claim 7 or 8, characterized in that for balancing the shaft ends of the screw rotor and, respectviely, slide rotor, balancing pistons located in the pressure spaces are pressed against the shaft ends, and that the outer diameters of the balancing pistons are sealed by mechanic sealings so that the oil pressures supplied to the pressure spaces are maintained.

10. A method as defined in claim 8 or 9, characterized in that from the pressure spaces oil is supplied via connections equipped with throttlings to the bearings for lubricating and cooling the same.

11. A method as defined in any one of claims 7-1 0, characterized in that the pressure of the oil supplied to the pressure space of the screw rotor is about the same as the outlet pressure of the screw-compressor.

1 2. A method as defined in any one of the preceding claims, characterized in that the oil is included in the same oil circulation system as the oil injection to the compression space, and that the pressure of the oil supplied to the bearing spaces and the pressure of oil supplied to the bearing spaces and the pressure space of the slide rotor is reduced through a throttling.

1 3. A method at an oil-injected screwcompressor substantially as described herein with reference to the accompanying drawing.

14. An oil-injected screw-compressor substantially as described herein with reference to the accompanying drawing.