DE69504961T2 - COMPRESSOR - Google Patents

Abstract

The compressor (101) has a driven rotatable shaft (2a, 2b, 7) carrying an impeller rotor stage (4, 5) arranged such that a thrust bearing (210a, 210b) acts directly on the impeller rotor stage (4, 5). Overheating of the bearing (210a, 210b) is thereby inhibited because heat generated at the bearing is transferred via the impeller stage (4, 5) to the working gas. As a preference, or alternatively, at least a portion of the shaft (2a, 2b, 7) is hollow to improve inertial characteristics and aid in bearing cooling.

Description

The present invention relates to a compressor.

When processing food, pharmaceuticals or other sensitive substances, it is desirable to provide compressed air or another working gas that is absolutely clean or "dry", i.e. completely free of oil or other bearing lubricants.

In the past, numerous attempts have been made to produce oil-free compressors, but such designs, such as dry spindle compressors, are costly, inefficient, consume a lot of electricity and are prone to failure.

The entire air compressor market covers a range of performance ranges, with each performance range including a range of working pressures in combination with a range of mass flows.

A working pressure of approximately 8.6 bara combined with a mass flow of 0.27 kg per second is within one of the market ranges for a dry air compressor. Such working pressures can currently be achieved without difficulty, However, the mass flow of a known turbo compressor of this type is much larger than the required mass flow.

In addition, turbo compressors mounted in conventional oil-lubricated roller or ball bearings would be unreasonably inefficient at the high shaft rotation speeds required to achieve the desired performance (typically 50,000 to 100,000 rpm). Conventional turbo compressors operating in this range would therefore be extremely expensive, large and inefficient.

EP-A-0 150 599 discloses a compressor comprising a rotatable shaft, bearing means provided for the shaft, drive means for rotating the shaft, at least two impeller stages mounted on the shaft in spaced relation, at least a portion of the rotatable shaft between the spaced impeller stages being substantially hollow.

According to the present invention, compressed air or working gas flows from a relatively higher pressure impeller stage to a relatively lower pressure impeller stage along a path that encompasses the hollow portion of the rotatable shaft.

For this purpose, flow passage means communicating with the hollow part of the rotating shaft are desirably provided. This is advantageous because the shaft is effectively cooled, resulting in further heat dissipation from the bearings.

The hollow portion of the shaft reduces the inertia torque of the combined impeller stage and shaft assembly, reducing the work required to rotate the shaft and thus increasing efficiency.

Preferably, thrust bearing means are provided which are arranged to act directly on the impeller stage such that, when the shaft rotates, a bearing contact is established between the thrust bearing and a bearing surface of the impeller stage.

This allows the heat generated at the thrust bearings to be transferred directly to the compressor's working gas via the impeller stage, thereby cooling the bearings and preventing overheating.

Preferably, the compressor is equipped with direct drive means arranged to rotate the shaft at high rotational speeds, preferably in the range of 50,000 to 100,000 rpm. The drive means therefore preferably comprise an electric motor with a rotor mounted on the shaft.

The electric motor is preferably arranged between the impeller stages. Preferably, the thrust bearing means are arranged so that they act directly on at least two impeller stages in order to absorb axial forces in opposite axial directions of the shaft.

It is desirable to provide intercooling means between impeller stages to increase the efficiency of the compressor.

It is desirable that the shaft comprises a composite shaft containing the hollow rotor portion interconnecting the spaced-apart parts of the shaft, the spaced-apart parts desirably carrying respective impeller stages. Advantageously, the hollow rotor portion of the shaft is made of magnetic or magnetizable material. The hollow construction of the rotor portion of the shaft keeps the mass (and hence the moment of inertia about the axis) of the composite shaft to a minimum.

Advantageously, fastening means are provided for fastening the hollow rotor part and the spaced-apart parts of the shaft relative to one another. Preferably, the fastening means comprise a connecting rod which passes through the hollow rotor part and the connected spaced-apart parts of the shaft.

Preferably, thrust bearing means are provided which act on impeller stages on both spaced parts of the shaft, and are arranged in such a way that axial thrust of the shaft is absorbed in opposite axial directions.

Preferably, the thrust bearing means are arranged to act on the corresponding impeller stage rotor such that heat generated at the bearing is dissipated to the impeller stage rotor. The thrust bearing means and the corresponding vane wheel stage are therefore preferably kept in thermally communicating bearing contact when the compressor is in operation.

This ensures that heat generated at the thrust bearing means is transferred to the corresponding impeller and then to the working gas flowing through the corresponding impeller stage of the compressor. The gas is then cooled as it enters the following intercooler means.

Furthermore, the compressor preferably contains axle bearing means for supporting the shaft, which preferably have at least one block axle bearing, which is advantageously self-generating and air or gas lubricated and preferably has bearing blocks which are provided with a ceramic bearing surface. The bearing blocks can comprise homogeneous blocks made of ceramic material.

The shaft is preferably equipped with hardened or ceramic surface parts with which the ceramic bearing surface of the corresponding tilting blocks of the axle bearing means interacts.

Advantageously, the bearing means comprise at least two journal bearings, each of which is preferably an air or gas lubricated pad journal bearing and includes bearing blocks provided with corresponding ceramic bearing surfaces. Alternatively, blade journal bearings may be used. Desirably, the journal bearings are arranged to support spaced-apart parts of the shaft, advantageously at opposite ends of the electric motor. Preferably, at least an axle bearing is arranged in the space between a corresponding end of the motor and a corresponding impeller stage.

The thrust bearing means preferably includes a thrust bearing with rocker pads cooperating with the impeller stage. Desirably the thrust bearing is self-generating and air or gas lubricated and includes pads having ceramic bearing surfaces.

Advantageously, the impeller stages are overhung at opposite ends of the shaft. Preferably, each impeller stage comprises a corresponding compressor impeller, with intercooler means being communicatively connected between the impeller stages.

It is desirable to provide three impellers so that the compressor comprises three compression stages. Preferably, the corresponding intercooling means are provided in the space between successive compression stages. This increases the efficiency of the compressor. Advantageously, the flow of the working gas into each corresponding impeller is axial, preferably in the direction of the electric motor.

Accordingly, it is preferred that at least two of the impeller stages are arranged in opposite formation relative to each other such that the respective flows into the respective impeller stages flow in opposite directions, preferably towards each other. This has the advantage that the axial thrust loads exerted on the shaft by the respective impeller stages approximately cancel each other out, thereby the axial pressure absorbed by the axle bearing means is reduced.

Preferably, sealing means for the shaft are provided, preferably with corresponding labyrinth seals, so that they prevent access of the working gas from the impeller stages to the motor and the bearing means.

Advantageously, the electric motor comprises a solenoid or permanent magnet electric motor, preferably designed to rotate the shaft at a speed of more than 50,000 rpm, more preferably more than 70,000 rpm. Desirably, the electric motor is a DC motor, preferably controlled by a variable frequency source.

The invention is explained below only by way of example and with reference to the accompanying drawings. In them:

Fig. 1 is a schematic representation of a compressor;

Fig. 2 is an enlarged detail of a part of the compressor of Fig. 1;

Fig. 3 is a schematic representation of a compressor according to the invention; and

Fig. 4 is an enlarged detail of a part of the compressor according to Fig. 3.

The drawings show a compressor, generally designated 1, as generally described in PCT publication WO94/05913. Although the compressor shown in Figures 1 and 2 is similar in general construction and illustrates the preferred features, it does not have the specific structural features which result in the improved performance of a compressor according to the invention. The compressor 1 comprises an axially rotatable shaft 2 mounted in a housing 3 on which are mounted molded aluminum impellers 4, 5 and 6.

The supply, first stage, wheel 4 is overhung at one end of the shaft, whereas the wheels 5 and 6 of the second and third stages are overhung at the opposite end. In the space between impellers 4 and 5 is a brushless DC motor with a rotor 7 with permanent magnets mounted on the shaft 2 and a stator 23 built into the housing. A solid state thyristor based inverter/controller (not shown) generates a variable high frequency current from a standard 415 V/50 Hz power source. The high frequency current drives the motor (thus driving the shaft 2 directly without the need for an intermediate gear) at the required high operating speed, which is typically in the order of 50,000 to 100,000 rpm. Since no gear is required to connect shaft 2 to the drive, only minimal power loss occurs.

The shaft 2 is mounted in the housing 3 on axle bearings 8, 9, which are located at both ends of the electric motor, adjacent to the blade wheels 4 and 5 respectively. A thrust bearing 10 is also incorporated in the housing to cooperate with the thrust ring 11 provided on the shaft. The axle bearings 8, 9 comprise tilting block bearings which are self-generating and air-lubricated. The tilting blocks 12 of the two axle bearings 8, 9 are mounted on flexible pivots 24 and are provided with ceramic bearing surfaces 13 which are arranged to cooperate with immediately adjacent bearing surface parts of the shaft. The bearing surface parts of the shaft are coated with a hardened coating to increase resistance to wear.

An important feature of the design is that friction losses in the bearings have been minimized to maximize compressor efficiency. Typically, when using fluid lubricated journal bearings (such as oil lubricated bearings) or ball or roller journal bearings in high speed rotating machinery, friction losses in the bearings account for between 5% and 10% of the drive power. The use of self-generating rocker block, air (or gas) bearings reduces friction losses to approximately 0.5% of the drive power. However, due to the fact that the rotational speed of the shaft is extremely high (e.g. 80,000 rpm for a compression of 1 bara to 8.5 bara at a mass flow of 0.27 kg/s for air), the temperature developed at the bearings is extremely high, which can cause problems with material expansion at the bearing/shaft due to the necessarily small bearing-shaft gap widths required for the operation of air or gas lubricated self-generating tilting block axle bearings (typically 0.003" diametric gap width for axle bearings). This problem is overcome by the use of ceramic materials materials for the bearing surfaces of the tilting blocks 12; the use of a surface coating with a hardened coating for the bearing parts of the shaft 2 also contributes to solving this problem.

The thrust bearing 10 is also provided with tilting pad thrust members 10a, 10b having ceramic bearing surfaces. The pads 10a are arranged to take up normal thrust loads transmitted from shaft 2 through thrust ring 11 during normal operation of the compressor. The pads 10b act on the opposite side of ring 11 and take up opposite thrust loads during "ramp-up" of the motor and shaft to normal operating speed.

To increase efficiency, an intercooler 15 is provided in the space between the first stage wheel 4 and the second stage wheel 5. A second intercooler 16 is provided in the space between the second stage wheel 5 and the last (third) stage wheel 6. It is an important feature of the compressor that the flow of the working gas into the first stage wheel 4 is opposite to the flow of the working gas into the second and third stage wheels 5, 6. This "balances" the axial pressure acting on the shaft and reduces the normal axial pressure acting on the thrust bearing 10. This minimizes bearing losses in the thrust bearing 10.

During operation, the electric motor runs up to an operating speed of about 80,000 rpm. Working gas is then sucked axially into the first impeller stage 4 and pressed out through channel 17 into the intercooler 15. The working gas leaves the intercooler 15, enters channel 18 and then travels axially into the second impeller stage 5. The working fluid leaves impeller 5 radially and enters the second intercooler 16 via channel 19. Intercoolers 15 and 16 are essentially identical except that intercooler 16 is arranged longitudinally at an angle of 90º to the longitudinal direction of intercooler 15 (ie the longitudinal direction of intercooler 16 runs out of the side in Fig. 1).

The working gas leaves the intercooler 16 via channel 20 and is directed to enter the third (and final) impeller stage 6 axially. The working gas leaves the final impeller stage 6 radially via the outlet channel 21 (the outflow through channel 21 is out the side in Fig. 1).

Due to the combination of the directly driven high-speed shaft together with the minimization of bearing losses and the arrangement of the impellers in separate stages with intercooling, a highly efficient compressor is achieved. The compressor makes a compact turbomachine possible, which can be used in applications that were previously reserved mainly for spindle compressors, since - unusually for a turbocompressor - high pressure outputs (typically 8.5 bara) can be achieved at relatively low mass flows (typically 0.27 kg/s for air).

With reference to Fig. 3, the embodiment shown of the compressor 101 corresponds generally in construction and operation to the arrangement shown in Figs. 1 and 2, and like reference numerals have been used for like components of the compressor.

In the embodiment shown in Fig. 3, the thrust ring 11 of the compressor design shown in Fig. 2 is omitted and a pair of spaced apart thrust bearings 210a, 210b are provided adjacent the first and second impeller stages 4, 5 respectively to accommodate axial forces acting on the shaft 2 in respective opposite directions. It has been found that in the compressor shown in Fig. 1, excessive heat is generated at the thrust bearing 10, resulting in reduced efficiency in terms of compressor performance and operating life. By replacing the thrust ring 11 and bearing assembly 10 with thrust bearings 210a, 210b acting directly on the rear substantially flat surfaces of the impeller stages 4 and 5 respectively (as shown in the embodiment of Fig. 3), overheating problems are significantly reduced. Heat generated at the thrust bearings 210a and 210b is transferred directly to the corresponding impeller stage rotor 4, 5 and then to the working gas flowing through the corresponding impeller stage rotor. The working gas is then cooled by passing through the corresponding intercoolers (not shown in Fig. 3) arranged between all the impeller stage rotors as in the device shown in Fig. 1. This effectively removes heat from the thrust bearings. The thrust bearings 210a, 210b comprise bearing pads 110a, 110b mounted in a corresponding circular support ring 37a, 37b held by housings 35a, 35b. The pads may be made of homogeneous ceramic material or may be provided with a ceramic bearing surface.

The embodiment of the invention shown in Fig. 3 differs from the arrangement shown in Fig. 1 in addition in that the shaft effectively comprises a hollow cross-section composite shaft having a first shaft portion 2a (which carries the impeller stage rotor 4), a second shaft portion 2b (which carries the impeller stage rotor 5) and an intermediate motor rotor section 7 extending between shaft portions 2a and 2b. Shaft portions 2a and 2b are connected to opposite ends of the motor rotor section 7, the entire composite shaft being held together by the axially disposed connecting rod 25. First and second shaft portions 2a, 2b are provided with respective cylindrical cavities 31, 32 which intersect the axis of the shaft.

The connecting rod 25 is provided along its entire length with groups of circumferentially spaced projections 40 adjacent internal axial bores of the shaft parts 2a, 2b and the motor rotor 7. Circumferential spaces between corresponding projections in each group 40 allow air communication substantially along the entire length of the interior of the assembled shaft in the area adjacent to the connecting rod 25. Compressed air or working gas flows back from the relatively higher pressure stage 5 (through the outflow connecting channel 42) and flows internally along the entire length of the composite shaft towards the relatively lower pressure stage 4. The flow of air or transport gas in the internal cavities 31, 32 and along the connecting rod causes heat dissipation from the shaft parts 2a, 2b (and thus the bearings 210a, 210b, 108, 109) and the motor rotor 7.

Furthermore, the axial cylindrical cavities provided in the shaft parts 2a and 2b reduce the moment of inertia of the assembled shaft about its axis of rotation, which improves the efficiency of the electromagnetic motor drive. The journal bearings 108, 109 are provided at opposite ends of the shaft and have bearings 112 which cooperate with corresponding shaft parts 2a, 2b. Favoured by the presence of the cavities 31, 32, heat generated in the shaft by bearing contact with the journal bearings is transferred directly to the impeller 5, 4, from where it is transferred to the working gas of the compressor. At each end of the shaft, the corresponding thrust bearings 210a, 210b and journal bearings 108, 109 are provided in a corresponding common housing unit 35a, 35b. The compressor shown in Fig. 3 and 4 operates in an almost identical manner to the compressor shown in Fig. 1. Intercoolers (not shown) are arranged between all impeller rotor stages 4, 5, 6 and the working gas flow through the compressor corresponds essentially to that described with respect to the compressor shown in Fig. 1.

Claims (10)

1. Compressor (101) comprising a rotatable shaft (12), bearing means (8, 9) provided for the shaft, drive means (7, 23) for rotating the shaft, at least two impeller stages (4, 5, 6) arranged at a distance from one another on the shaft, at least one substantially hollow part (31, 32) of the rotatable shaft between the spaced impeller stages (4, 5, 6), characterized in that compressed air or working gas flows from an impeller stage (5) of relatively higher pressure to an impeller stage (4) of relatively lower pressure, along a path which includes the hollow part (31, 32) of the rotatable shaft. 2. Compressor (101) according to claim 1, characterized in that the shaft comprises a composite shaft which includes a rotor part (7) which interconnects the spaced-apart parts (2a, 2b) of the shaft (23), the spaced-apart Parts are provided with corresponding axial cavities (31, 32) and carry corresponding impeller stages (4, 5). 3. Compressor (101) according to claim 1 or 2, characterized in that the hollow part (31, 32) of the shaft consists of a magnetic or magnetizable material. 4. Compressor (101) according to one of the preceding claims, characterized in that the drive means (7, 23) comprise electromagnetic drive means, preferably an electric motor with a rotor (7) mounted on the shaft (2). 5. Compressor (101) according to one of the preceding claims, characterized in that pressure bearing means (210a, 210b) are arranged so that they act directly on the impeller stages (4, 5) in order to absorb forces acting in the axial direction of the shaft. 6. Compressor (101) according to one of the preceding claims, characterized in that intercooler means (15, 16) are provided between the impeller stages (4, 5). 7. Compressor (101) according to one of the preceding claims, characterized in that the compressor further comprises axle bearing means (108, 109) for supporting the shaft. 8. Compressor (101) according to claim 7, characterized in that the axle bearing means (108, 109) comprise at least one block axle bearing. 9. Compressor (101) according to one of the preceding claims, characterized in that impeller stages (4, 5, 6) are arranged overhung at opposite ends of the shaft (2). 10. Compressor (101) according to one of the preceding claims, characterized in that impeller stages are arranged in opposite formation (4, 6 and 4, 5) relative to each other, so that the corresponding flows flow into the corresponding impeller stages in opposite directions.