WO2000017524A1

WO2000017524A1 - Two-stage centrifugal compressor driven directly by motor

Info

Publication number: WO2000017524A1
Application number: PCT/JP1998/004218
Authority: WO
Inventors: Haruo Miura; Kazuki Takahashi; Naohiko Takahashi; Hideo Nishida; Yasuo Fukushima; Minoru Yoshihara; Tsunehiro Endou
Original assignee: Hitachi, Ltd.
Priority date: 1998-09-18
Filing date: 1998-09-18
Publication date: 2000-03-30
Also published as: JP3918432B2

Abstract

A two-stage centrifugal compressor driven directly by a motor which has impellers directly mounted on both end portions of a shaft thereof, and which comprises an inverter-driven high-frequency motor rotatable at a high speed, the rotary shaft being supported rotatably on magnetic bearings, whereby the two-stage centrifugal compressor does not require an oil and maintenance work. A rotor of the motor comprises a permanent magnet, a ring-shaped carbon fiber-reinforced plastic member for retaining the permanent magnet, and a metal sleeve for holding the permanent magnet between the metal sleeve and the carbon fiber-reinforced plastic member, the occurrence of a clearance between the permanent magnet and the rotary shaft which is liable to be encountered during a high-speed rotation of the motor being prevented by utilizing the strength of the carbon fiber. The rotary shaft which is rotated at a high speed causes the motor to generate heat due to an iron loss, a copper loss and a windage loss, and also the magnetic bearings to generate heat in the same manner, a part of an operating gas which has been cooled in an intermediate cooler or a discharge cooler being therefore introduced into a central portion and a labyrinth of a stator of the motor to cool the motor and magnetic bearings.

Description

Description Two-stage centrifugal compressor driven directly by electric motor-Technical Field

TECHNICAL FIELD The present invention relates to a two-stage centrifugal compressor driven directly by an electric motor, and more particularly to a two-stage centrifugal compressor directly used to be driven by an electric motor suitable for a small-capacity air compressor and used for a power air source in a factory. . Background art

Screw compressors and turbo compressors are used as factory power air sources with a discharge pressure of about 8 OMPa in gauge pressure and a compressor capacity of less than 100 kW to several hundred kW. Among them, the use of turbo compressors is increasing due to easy maintenance. An example of this turbo compressor is described in Japanese Patent Application Laid-Open No. Hei 8-159904 / Japanese Patent Application Laid-Open No. Hei 9-287579. In the turbocompressors described in these publications, the rotation of a low-speed induction motor is increased at a high speed using a gear speed increaser, and blades are provided at both ends of an output shaft, which is a high speed rotation shaft of the gear speed increaser. The car is installed to form a two-stage air compressor.

In the case of an evening compressor, the impeller should be rotated at a high speed in order to compress the air efficiently with as little fluid power as possible. Conventionally, it was not easy to obtain a large-capacity motor capable of high-speed rotation, so the speed of the motor was increased by using a gear speed increaser. Use of the gearbox requires at least three shafts: a motor shaft, a bull gear shaft, and a pinion shaft for mounting the impeller. Each shaft has a contact-type rolling bearing that rotatably supports the shaft. Bearings are required. As compressors have become smaller and smaller at high speeds, the peripheral speed of sliding bearings has also increased significantly compared to conventional bearings, and the operating conditions of sliding bearings have become more severe, making it difficult to ensure bearing life. . On the other hand, there is a demand for smaller gears for gearboxes, and the gear ratio is gradually increasing. To increase the speed increase ratio, the module value of the gear is reduced, and the surface pressure and sliding speed are increased. As a result, the use conditions have become stricter than before.

By the way, the load on each component used in a turbo-type centrifugal compressor is still smaller than that of a positive displacement compressor, despite the increase in compact size. Therefore, the maintenance interval can be made relatively long, but it is difficult to achieve complete maintenance-free as long as there is contact between the rotating part such as a bearing and the stationary part.

In the case of a centrifugal compressor with a built-in gear speed increasing device, the impeller is loaded on the rotating shaft, so that the inertia force is large. If the inertia force is large, the transient torque will also be large, and the time required for starting and stopping will be long. In addition, the addition of the gearbox increases the complexity of the structure, and it is not easy to avoid the vibration eigenvalue. Therefore, with a turbo compressor equipped with a gearbox, it is difficult to frequently adjust the capacity by starting and stopping and controlling the number of revolutions.

When the turbo compressor is made compact, it is important to secure a heat radiation path to radiate the heat generated inside the compressor to the outside by the heat generated by compressing the working gas and the friction. Therefore, it is common to provide an intercooler or discharge cooler near the turbo compressor body. In order to reduce the size of the compressor as a whole, the size of the cooler also needs to be reduced. In a centrifugal compressor equipped with a gear speed increaser described in Japanese Patent Application Laid-Open No. H08-105538, an intercooler and The cooling fluid inlet and outlet of the discharge cooler are horizontal, which is consistent with the fact that the gear speed increaser necessarily requires a large horizontal width. However, in the method described in Japanese Patent Application Laid-Open No. Hei 8-105386, the drain generated by cooling in the cooler strike is discharged together with the flow of the working fluid, so that the pressure loss in the nest is reduced. There is a problem in that the working gas becomes large and the working gas contains drain.

On the other hand, in order to solve the above-mentioned problems of the centrifugal compressor with a speed increaser, an impeller is provided at the shaft end of the motor, and the direct drive centrifugal compressor drives the impeller directly without passing through the speed increaser. An example of this is disclosed in Japanese Patent Publication No. 5-36640. In the compressor described in this publication, the compressor casing and the motor casing are formed separately. For this reason, a gearbox is not required, and although a certain degree of downsizing has been achieved, it is still insufficient to reduce the overall size of the compressor.

SUMMARY OF THE INVENTION The present invention has been made in view of the problems and disadvantages of the related art described above, and has as its object to realize a two-stage centrifugal compressor that is directly driven by a compact electric motor that does not require a gear speed increasing device. It is in. Another object of the present invention is to realize a two-stage centrifugal compressor which is easy to assemble and has excellent maintainability. Still another object of the present invention is to provide a two-stage centrifugal compressor that is directly driven by an earth-friendly electric motor that can reduce power consumption and reduce the generation of carbon dioxide as much as possible. Disclosure of the invention

A first feature of the present invention for achieving the above object is that a rotating shaft having a motor rotor formed in an intermediate portion, and a centrifugal shaft attached to both ends of the rotating shaft. A two-stage motor driven directly by a motor having a compressor impeller, a motor stay that constitutes a motor together with the rotor, and radial magnetic bearing means and thrust magnetic bearing means rotatably supporting the rotating shaft. At the centrifugal compressor, a water-cooled jacket is provided around the outer periphery of the motor for stays "", and the two-stage centrifugal compressor is sucked in the middle of the stay. The gas flow path that leads the rare working gas is formed.

Preferably, the impellers mounted on both ends of the rotating shaft constitute a first compressor stage and a second compressor stage, and intermediate cooling for cooling the working gas compressed in these compressor stages. And an integral casing for accommodating the intermediate cooler and the discharge cooler together with the impeller, the radial magnetic bearing means, the thrust magnetic bearing and the electric motor. Things.

More preferably, a cooling path is formed such that the working gas compressed in the first compressor stage flows into the intercooler from above, is cooled by the intercooler, and then flows out from below. Alternatively, a cooling path is formed such that the working gas compressed in the second compressor stage flows into the discharge cooler from above, is cooled by the discharge cooler, and then flows out from below.

A second feature of the present invention to achieve the above object is that a rotating shaft having a rotor of an electric motor formed in an intermediate portion, impellers attached to both ends of the rotating shaft, and an electric motor together with the rotor A two-stage centrifugal compressor directly driven by an electric motor having a pair of radial magnetic bearing means and a thrust magnetic bearing means rotatably supporting the rotating shaft; Rotating the intercooler for cooling the working gas compressed by the impeller, and the discharge cooler for cooling the working gas compressed by the other impeller; An integral casing is provided below the shaft and substantially parallel to the rotating shaft, and houses the rotating shaft, the intercooler and the discharge cooler, and is cooled by either the discharge cooler or the intercooler. A cooling path is formed for guiding the working gas to an intermediate portion of the electric stay.

A window is formed at an axial end of the casing that houses the intermediate cooler and the discharge cooler. A cover is detachably attached to the window, and the intermediate cooler and the discharge cooler are inserted from the axial direction. Preferably it is possible. The motor is a permanent magnet type synchronous motor, and the rotor is a ring-shaped permanent magnet, a ring-shaped carbon fiber reinforced plastic holding member for holding the permanent magnet, and a ring-shaped holding member. It is preferable to have a metal sleeve for holding the permanent magnet together with the member.

Further, cooling switching means for selectively cooling the radial magnetic bearing means and the thrust magnetic bearing means with a working fluid cooled by the intercooler during load operation and with a working fluid cooled by the discharge cooler during no load operation. It is desirable to have

The rotor has two poles, means for detecting a magnetic pole position based on an induced voltage generated in a winding of a motor stator when the rotor rotates, and means for detecting the detected magnetic pole. Inverter control means for controlling the electric motor based on the position may be provided.

Further, an invertor for driving the electric motor, pressure detecting means for detecting a pressure on the downstream side of the second-stage impeller, and when the detected pressure is equal to or higher than a predetermined pressure, the working gas is discharged to the outside. A blow-off valve for blowing air, and control means for controlling the inverter based on the pressure detected by the pressure detection means, and the predetermined blow-off valve opening set pressure even when the set lower limit rotation speed is reached. that's all In the case where the detected pressure rises, a blow-off valve may be opened to form a flow path for returning the working fluid flowing through the blow-off valve to the upstream side of the first stage impeller. Further, the electric motor may be provided with cooling switching means for selectively cooling the motor with the working fluid cooled by the intercooler during the load operation and with the working fluid cooled by the discharge cooler during the load operation. BRIEF DESCRIPTION OF THE FIGURES

FIG. 1 is a front view of an embodiment of a two-stage centrifugal compressor directly driven by an electric motor according to the present invention, and FIGS. 2 and 3 are side views of the centrifugal compressor shown in FIG. 2 is a left side view, and FIG. 3 is a right side view. FIG. 4 is a longitudinal sectional view around the rotation axis of the embodiment of FIG. 1, FIG. 5 is a circuit diagram of an embodiment of the motor driving circuit, and FIG. 6 is a motor driven by an inverter. FIG. 7 is a circuit diagram showing one embodiment of a position detection circuit when driving the motor, and FIG. 7 is an explanatory diagram for explaining position detection. FIG. 8 is a detailed view of the electric motor portion of the rotating shaft shown in FIG. 4, and FIG. 9 is a view as viewed from C in FIG. FIG. 10 is a piping flow diagram of the compressor shown in FIG. 1; FIGS. 11 and 12 are diagrams for explaining control of the centrifugal compressor; FIG. FIG. 12 is a diagram showing a pressure-flow rate characteristic. FIG. 13 is a detailed view of a cooling structure of a bearing portion in one embodiment of the present invention, and FIG. 14 is a system diagram for sending a cooling gas to the cooling structure shown in FIG. BEST MODE FOR CARRYING OUT THE INVENTION

An embodiment of the present invention will be described with reference to the drawings. 1 to 3 are external views of one embodiment of a two-stage centrifugal compressor according to the present invention, and FIG. 2 is a left side view, and FIG. 3 is a right side view. Further, FIG. 4 is a longitudinal sectional view around the rotation axis of the two-stage centrifugal compressor according to the present invention. As shown in Figure 4, the two-stage centrifugal compressor of the present embodiment, the rotary shaft 2 0 the electric motor rotor f 0 _a is formed in the central portion, the radial magnetic bearing 2 3 a in the axial direction two positions, 2 3 b supports in the radial direction. Also, thrust magnetic bearings 24a and 24b provided adjacent to one radial magnetic bearing 23b support the rotating shaft 20 in a non-contact manner in the axial direction. A first-stage impeller 21a and a second-stage impeller 21b are attached to both ends of the rotating shaft 20, respectively. A motor stator 22 is arranged opposite the motor rotor 20a. Cooling water that cools the heat generated by the copper loss and iron loss that occur during the stay, or the wind loss that occurs between the stay 22 and the rotor 20a around the stay 22 Jacket 14 is provided. This stage 22 and the cooling water jacket 14 are assembled and integrated in advance before assembling the compressor.

The impellers 21a and 21b at each stage are of an open shroud type having no shroud. The first-stage impeller 2 1a is kept between the inner casing 17b and the second-stage impeller 2 1b is kept between the inner casing 18b and the minute gap. It is spinning.

Further, both the first stage compressor 1a and the second stage compressor 1b are axially covered with head covers 17 and 18 at their axial ends. The head cover forms a suction stationary flow path wall surface and a discharge stationary flow path wall surface. The first-stage head cover 17 includes pressure-resistant closing portions 17a and 17d, an inner casing portion 17b that forms a stationary flow path, and ribs 17c. The pressure-resistant closing portion 17d may be integrated with the pressure-resistant closing portion 17a. Ma In addition, the rib 17c secures a flow path between the pressure-resistant closing portion 17a and the inner casing 17b. The head cover 18 of the second-stage compressor 1b has pressure-resistant closing portions 18a and 18d, an inner casing 18 and ribs 18c as in the first stage.

The outer diameter of the motor cooling water jacket is larger than at least one of the outer diameters of the magnetic bearing housings 28a and 28b in which the magnetic bearings are assembled. In the embodiment shown in FIG. 4, the outer diameter of each of the magnetic bearing housings 28a and 28b is larger. Furthermore, the outer diameter of the head covers 17 and 18 is larger than the outer diameter of the magnetic bearing housing. By determining the relationship between the size of the motor, bearing, and compressor section in this manner, the internal components such as the rotating shaft and the impeller accommodated in the casing can be easily inserted into the casing from the axial direction. It can be assembled. Each part can be incorporated into the motor casing 2 and the compressor casings 1a and 1b from the axial direction, so that the motor casing 2 and the compressor casings 1a and 1b are circles without division. It is possible to have a cylindrical and axially integral structure.

In order to rotate a two-stage centrifugal compressor directly driven by an electric motor configured as described above at the same speed as a conventional gear-increasing compressor, the shaft length must be as short as possible. However, the diameter must be as large as possible. (This is because the motor rotor is almost at the center of the rotating shaft and is much longer than the tooth width of the gears. The operating speed of the compressor according to the present invention is 40,000 RPM or more, and ordinary oil bearings cannot control such high-speed rotating shafts. In this embodiment, a magnetic bearing is used as a bearing, and the operating range of the compressor is thereby reduced to the third and fourth critical speeds of the rotating shaft. P

It is stored between 9 degrees.

In this embodiment, a two-pole permanent magnet synchronous motor driven by an inverter is used as the motor. Then, the induced voltage generated in the winding of the electric stay is used for detecting the magnetic pole position. This eliminates the need for a dedicated magnetic pole position detection sensor and shortens the rotating shaft length. Conventionally, a resolver, an optical encoder, or the like is used to detect the magnetic pole position. However, these detectors can prevent the shaft length from becoming longer and the dangerous speed from lowering. Detecting the magnetic pole position using the induced voltage is often used for small motors with a capacity of about several kW or less, but 400 000 r r ρ π ! It was difficult to apply to a compressor rotating up to 800,000 rpm. In this embodiment, the magnetic pole position detection using the induced voltage is enabled by the method described below. The permanent magnet type brushless synchronous motor used in the two-stage centrifugal compressor of this embodiment is a type of DC brushless motor. Fig. 5 shows an example of a drive circuit for this DC brushless motor.

Depending on the rotational position of the motor shaft or the size of the load, the power switch is turned ON and OFF by, for example, a 120 ° conduction method, and current is applied to each of the phases a, b, and c of the motor winding. Shed. The drive unit of the DC brushless motor includes an inverter unit 47, an energization switching unit 46 that instructs the inverter unit 47 to switch energization, and a position detection circuit 45. The position detection circuit 45 is for generating an energization switching signal, and detects the circumferential position of the permanent magnet rotor from the voltage induced in the motor. The inverter unit 47 includes a switching transistor and a switching transistor in accordance with the current switching timing to the windings 44 a, 44 b, and 44 c of each phase of the motor designated by the current switching unit 46. Switching elements consisting of diodes 48a, 48b, 48c and switching Operate the elements 49a, 49b, 49c.

In the 120 ° energization method, a positive voltage is applied to the a-phase winding 44a only for a section of about 120 ° in the rotation angle of the rotation axis 20 and then the rotation angle of about 60 ° Section, stop energizing this a-phase. After that, a negative voltage is applied only for a section of about 120 °, and then the energization is stopped for a section of about 60 °. Energize the b-phase and c-phase in the same way. However, the energization timing for each of the a-phase, b-phase, and c-phase is varied by 120 ° each. As a result, a rotating magnetic field is generated, and the brushless motor is driven.

Next, FIG. 6 shows details of the position detection circuit 45 shown in FIG. In FIG. 6, a block shows a rotational position detecting function based on the terminal voltage of the motor. The terminal voltages Va, Vb, and Vc of the motor are input to an integrator 42 provided for each phase and integrated. By integrating with the integrator 42, the phases of the terminal voltages Va, Vb, and Vc are shifted by 90 °. At the same time, the ripple component of harmonics generated by pulse width modulation (PWM) control is removed, and the signal approximates a sine wave. Here, if the integrator 42 performs a pure integration process of integrating all the signal components, the minute DC voltage is amplified in a meaningless manner. To avoid this problem, the integrator 42 has a high-pass filter for cutting the DC voltage. The signal passed through the integrator 42 is compared with the phase terminal voltages V a, V b, and V c at the neutral potential V n by the comparator 43, and a pulse with a duty ratio of 1: 1 is used to make 0 NZO FF at the zero crossing point. A signal is obtained. Here, the duty ratio is the ratio of the ON and OFF times. This pulse signal is obtained for each phase and is shifted by 120 ° each.

The source of the pulse signal is the phase terminal voltage whose phase changes according to the phase current of the motor. Because it is pressure, the rotation position pulse signal changes with the phase terminal voltage as it is, and the rotation position of the motor cannot be specified. To obtain the rotational position of the motor, the rotational position pulse signal is corrected with the phase current of the motor. The method of correction using the phase current will be described with reference to FIG.

The phase terminal voltage V is represented by the vector sum of the induced voltage E ₀ and the impedance drop generated when the phase current Im flows through the phase resistance r and the phase inductance L. In other words, the induced voltage E 0 is the vector difference between the phase terminal voltage V and the impedance drop. By obtaining the phase difference 5 between the phase terminal voltage V and the induced voltage E ₀ and subtracting this value from the phase of the phase terminal voltage V, the phase of the induced voltage, that is, the rotation angle information of the rotating shaft can be obtained. The phase corrector 41 in FIG. 6 calculates the phase difference S, and corrects the phase of the rotational position pulse signal based on the obtained phase difference <5. As a result, a pulse signal whose phase matches the induced voltage is obtained regardless of the magnitude of the motor current. The inverter for controlling the motor current switches the energization of the switching element based on the rotational position signal whose phase has been corrected by the phase corrector 41.

By the way, in order to increase the critical speed of the rotating shaft 20, it is common to increase the rotating shaft diameter as much as possible or to shorten the rotating shaft length. Fig. 8 shows an example of increasing the critical speed by increasing the diameter of the rotating shaft. FIG. 8 shows a schematic structure of a rotor part of a permanent magnet type synchronous motor. FIG. 9 is a view as viewed from C in FIG. A ring-shaped permanent magnet 61 is held by a holding member 62. The holding material is required to have properties such as being a non-magnetic material, having high specific strength, having low electric conductivity, that is, having high specific resistance. CFRP (carbon fiber) Fiber-reinforced plastic material).

By the way, since the rotating shaft of this embodiment rotates beyond the third critical speed, it is necessary to prevent the permanent magnet from separating from the rotating shaft due to centrifugal force. An interference is formed between CFRP and rotating shaft. Since the strength of the permanent magnet is very small, the permanent magnet itself cannot bear the centrifugal force. If the above-mentioned interference is insufficient, the permanent magnet separates from the shaft during operation, causing imbalance. If unbalance occurs, unbalance vibration will increase and the compressor will not be able to operate stably. For this reason, under the operating conditions of the present embodiment, a closing margin of about 3 mm for 100 minutes is required. The coefficient of thermal expansion in the circumferential direction of CFRP is almost 0, and shrink-fit which is generally used is impossible. Hydraulic fitting is also difficult because of the low tensile strength of the CFRP material in the axial direction. In addition, a cold fit that cools the rotating shaft does not cause temperature deformation corresponding to the above-mentioned interference. In addition, in the case of the present embodiment, the press-fitting method used in the processing of -CFRP is inappropriate because the compressive strength of CFRP is insufficient because the axial length of CFRP is long.

Therefore, in this embodiment, a steel sleeve 64 is interposed between the magnet 51 and the rotating shaft 20a. Before assembling the rotating shaft 20, carbon fibers are wound around the outer periphery of the multilayer ring magnet while impregnating the adhesive resin, and the resin is set at a high temperature in a high-temperature furnace. This forms an integral sleeve. — On the other hand, prepare a steel sleeve 64 having the same outer diameter as the inner diameter of the integral sleeve, and attach the above-mentioned integral sleeve to the outer periphery of this steel sleeve. Make the diameter smaller than the outer diameter of the rotating shaft by an amount equivalent to a minute. Next, hydraulic pressure is applied to the inner diameter side of the steel sleeve to forcibly expand the inner diameter of the steel sleeve, and the expanded steel sleeve is fitted to the rotating shaft. Put in.

To avoid gaps during operation, carbon fibers having a large modulus of elasticity may be used. However, carbon fibers having a high elastic modulus have a low tensile strength, and conversely, a high tensile strength has a low elastic modulus. Therefore, the structure shown in Fig. 8 and Fig. 9 is used to secure the required interference and to enable the use of carbon fiber with high tensile strength to realize a large-diameter rotary shaft. Although the number of slots is set to six in the station core 2a shown in FIG. 9, the number of slots is not limited to this and can be increased or decreased as needed.

The CFRP rings 63a and 63b are attached to both ends of the permanent magnet. This ring is provided to reduce the leakage of the rotating magnetic field generated by the stator winding and to reduce the eddy current loss due to the harmonic components of the rotating magnetic field. The mechanical properties of CFP and magnets decrease with increasing temperature. Therefore, if the heat generated by the loss in the rotor part is reduced and the mechanical characteristics are secured, the speed of the rotating shaft can be increased.

The built-in portion of the compressor described above is housed in a casing having an integral structure shown in FIG. The first-stage compressor section 1a has a first-stage suction nozzle 4 having a suction filter silencer 19, a discharge nozzle, and a one-stage discharge flow forming a communication flow path to the intercooler 3a. Road 5 The suction nozzle 4 and the discharge nozzle flow path 5 are formed so that the working fluid is sucked and discharged from the radial direction of the rotation shaft.

Similarly, the second-stage compressor section lb has a suction passage 6 from the intercooler 3b, a discharge nozzle, and a second-stage discharge forming a communication flow passage to the discharge cooler 3b. And a flow path 7. The suction nozzle 6 and the discharge nozzle flow path 7 The working fluid is formed so as to be sucked and discharged from the radial direction of the rotating shaft.

The intercooler 3a is stored in the intercooler case 1c, and the discharge cooler 3b is stored in the discharge cooler case 1d. The intercooler case 1c and the discharge cooler case 1 are arranged in parallel so that the flows from the inlet to the outlet are opposite to each other in opposite directions. And this flow direction is made to coincide with the axial direction of the rotating shaft. The intercooler case 1 c and the discharge cooler case 1 are joined at opposing mating surfaces.

The intercooler 3a is arranged so that the flow in the nest of the intercooler 3a goes from the upper surface to the lower surface. Discharge cooler 3b is also placed in case 1d in the same way.

The first-stage discharge passage 5 is connected to the upper surface of one end of the intercooler case 1c. The flow that has flowed into the intercooler case 1c flows out from the lower surface of the intercooler. For example, as shown in the embodiment shown in FIG. 1, this flow passes through the lower part of the other end of the inlet of the intermediate cooler case 1C, the lower part of the discharge cooler, and the outer end face of the discharge cooler case 1d. Be guided. A pipe is connected between the outlet and the second stage suction nozzle 6. The flow that has flowed out of the second-stage compressor is guided to the upper surface of the discharge cooler case 1d at the outlet side of the intercooler through the second-stage discharge passage. The flow that has flowed into the discharge cooler case 1d flows out from the lower surface of the discharge cooler. Then, this flow is discharged from the other end side surface of the inlet of the discharge cooler case Id.

In this embodiment in which the compressor stage, the intercooler and the discharge cooler are configured as described above, the width of the cooler case in the direction perpendicular to the rotation axis direction can be reduced, and the casing can be integrally structured. Making it possible. In other words, this implementation According to the example, the motor case 2, the compressor casings 1a and 1b, and the cooler cases lc and 1d are integrated by adopting a natural structure.

If the casing is made of stuff, it is necessary to set a parting surface between the upper and lower rust molds. In this embodiment, it is desirable that the surface pass through the center of the joint between the rotating shaft and the intercooler and discharge cooler. In order to set the parting plane at this location on the premise of a monolith, it is necessary to reduce the depth dimension from the parting plane. For this purpose, the working fluid may be flowed downward from above in the cooler.

Windows are provided on the axial end faces of the intercooler case 1c and the discharge cooler case 1d, and the intercooler 3a and the discharge cooler 3b are inserted from one of the windows, and then the cooler head cover Attach 15a, 15b, 15c and 15d. The cooler head covers 15a and 15b may be integral or may be divided. The windows at the other end are closed with covers 16a and 16b. The covers 16a and 16b may also be integral or may be divided.

The cooling water of the intercooler 3a and the discharge cooler 3b flows in from the head covers 15c and 15d, and is turned back by the end covers 16c and 16d at the tips of these coolers and again. It flows into the head covers 15c and 15d. With this configuration, the head covers 15c, 15d, and the closing covers 16a, 1a can be used while the nests of the intercooler 3a and the discharge cooler 3b are attached to the compressor integrated casing. Simply removing the 6b and the end covers 16c and 16d can clean the water-side cooler that is the most dirty. Therefore, maintenance becomes easier.

Wind loss, copper loss, and iron loss occur in the rotating shaft bearing and the motor. 75 P

16 Generates heat. In order to dissipate this heat, a working fluid cooled by a cooler after being compressed by a compressor, that is, compressed air, is used. However, if the amount of cooling air is not minimized, the unit consumption of the compressor, that is, the power consumption will increase. Therefore, the working fluid leaking from the back of the impeller, which was originally expected as a loss, is used as cooling air for the bearing. In FIG. 4, a labyrinth 29 is formed on the back of the first-stage impeller 21a, and a labyrinth 30 is formed on the back of the second-stage impeller 21b. The first-stage labyrinth 29 has a passage hole 29a for blowing cooling air into a plurality of labyrinth chambers formed in the labyrinth.

Next, a method of blowing cooling air will be described with reference to FIG. FIG. 10 is a diagram schematically showing one embodiment of a piping system diagram of a two-stage centrifugal compressor. In the integral casings 1a, 1b, 1c, 1d and 2, a rotating shaft 20 to which the impellers 21a and 21b are mounted is rotatably accommodated. The fluid sucked from the suction nozzle 4 is compressed by the first stage compressor, cooled by the intercooler 3a, and guided to the second stage suction nozzle. The fluid sucked from the second-stage suction nozzle is further compressed by the second-stage impeller 21b and then cooled by the discharge cooler 3b. Pumped to

On the other hand, the cooling fluid is taken out from the intermediate between the outlet of the intercooler 3a and the suction nozzle 6 of the second stage, passes through the filter 72 and the check valve 73, and passes through the cooling hole 29a of the first stage lapel It is led to. The cooling fluid is also guided to the motor cooling holes 32 and the bearing cooling holes 31 on the second stage side. Here, the 0 N / 0 FF valve 74 adjusts the flow rate of the fluid for cooling the electric motor, and is operated by the electromagnetic valve 82. It should be noted that when the compressor is operating under no load, When the wind valve 77 is open, the pressure of the cooling fluid taken out from between the first-stage compressor and the second-stage compressor decreases. Since cooling is necessary even in such a state, a part of the working fluid flowing out of the discharge cooler 3b is taken out and used as a cooling fluid. After passing through the filter 75 and the ONZOFF valve 76, the cooling fluid joins a cooling line taken out from between the first-stage compressor and the second-stage compressor.

An example of switching of the cooling line will be described below. The pressure detector 92 shown in FIG. 1 detects the pressure downstream of the discharge check valve 71 of the compressor. When the detected pressure exceeds the set pressure P d H T at the minimum rotation speed N min for avoiding surging, the arithmetic processing unit 93 issues a command to blow air through the blow-off valve 77. When a command to open the blow-off valve 77 is issued for blowing air, the solenoid valve 80 operates and the blow-off valve 77 opens. At this time, the arithmetic processing unit 93 simultaneously issues an operation command for the solenoid valve 81. When the solenoid valve 81 opens, the ON / OFF valve 76 also opens. It should be noted that a flow regulating valve 78 and a filter 79 are arranged upstream of the solenoid valve 80.

The cooling air supply line branches off from the downstream of the discharge cooler and is always open.In addition, it branches off from the downstream of the intercooler and opens with the blow-off valve 77 opened. Two cooling lines in the line are operational. In this state, the pressure of the cooling line branched from the downstream of the intercooler is higher than the pressure of the cooling line branched from the downstream of the intercooler. Back pressure is applied to valve 73 and this line is automatically closed.

In this embodiment, no cooling air is blown into the lapiling chamber formed on the back of the second-stage impeller. This is a labyrinth from downstream of discharge cooler 3b. When fluid is blown into the suction chamber, the static pressure at the back of the second-stage impeller increases, and the thrust force by the fluid from the back side of the second-stage impeller toward the suction side increases. If the thrust force is large, the thrust of the magnetic axis g of the thrust must be increased, which causes a decrease in the strength of the thrust force and an increase in the loss. Is no longer effective.

During load operation, the working fluid extracted from the upstream side of the second stage suction section is used for cooling, and during no-load operation, the working fluid extracted from the discharge cooler outlet downstream is used for cooling. This cooling fluid is blown into the motor section, near the inlet side of the labyrinth on the back side of the first stage impeller, and into the downstream side of the labyrinth on the back side of the second stage impeller, thereby forming the entire compressor. Reduce the power used for cooling

Further, in the two-stage centrifugal compressor of the present invention, the total power required for operation is reduced by the method described below. FIG. 11 shows an embodiment of the rotation speed control of a motor in a two-stage centrifugal compressor driven directly by a motor. The first-stage suction temperature 91 is detected by a temperature sensor (not shown), and the detected signal is input to the arithmetic processing circuit 93. In this arithmetic processing circuit, the ratio between the temperature obtained from the detected signal and the design suction temperature (T s de s) stored in advance therein is obtained. Then, the rated speed (N max) at that temperature (T s) is calculated by multiplying the design speed by approximately the 1/3 power of the temperature ratio. Further, a design minimum rotation speed (N min) in the load operation is calculated corresponding to the rated rotation speed.

On the other hand, a sensor for detecting the brand pressure 92 is provided downstream of the check valve 71, and a signal detected by this sensor is also input to the arithmetic processing unit 93. Then, the output of the inverter 94 is adjusted so that the plant pressure 92 becomes a predetermined value between the rated rotation speed Nmax and the design minimum rotation speed Nmin. That is, the current supplied from the inverter 94 to the motor 2 is controlled in accordance with the output adjustment signal of the inverter obtained by the arithmetic processing device 93. As a result, the number of revolutions of the motor changes, and the capacity of the compressor is adjusted.

A method of controlling the compressor speed by controlling the number of revolutions of the compressor will be described with reference to FIG. The set air volume of the compressor is determined under the summer suction temperature conditions. Therefore, if the rotation speed does not change, the flow rate at which the set discharge pressure (P dp) is obtained shifts to the larger flow rate side from the design point when the suction temperature decreases. In this case, the power of the compressor increases. If more working fluid is supplied to the compressor than necessary, the no-load operation time of the compressor becomes longer, and the power consumption also increases in this case.

Therefore, the cube value of the ratio between the suction temperature and the design suction temperature is calculated, and the rotation speed obtained by multiplying the design rotation speed is set as the rated rotation speed under the suction temperature condition. Thereafter, the compressor is controlled using this rated rotation speed as the upper limit rotation speed for compressor operation. On the other hand, the minimum rotational speed of the compressor is obtained as follows. In a speed-controlled compressor, the flow rate decreases as the speed decreases. If the flow rate is lower than a certain value, surging will occur and unstable operation will occur. Therefore, surging must be avoided. The operating speed at which this surging can be avoided is the minimum speed. Specifically, the value obtained by multiplying the previously obtained rated speed by a certain ratio is defined as the minimum speed (N min). Since this minimum rotation speed has already been corrected for the suction temperature, this value can be used as the minimum rotation speed even if the suction temperature changes. Therefore, by changing the compressor speed between the rated speed and the above minimum speed, The capacity of the compressor can be easily controlled. If the capacity can be adjusted by changing the rotation speed of the compressor, the no-load operation time of the compressor can be reduced, and the total power consumption of the compressor can be reduced. -When the rotation speed is controlled and the lower limit rotation speed is reached, the blow-off valve 7 will also be set if the pressure (P dp) downstream of the check valve rises above the set pressure of the blow-off valve (P d _H n). Open 7. Then, the discharged working fluid is returned to the first-stage suction flow path from the direction perpendicular to the rotation axis direction, that is, the operation flowing out of the blow-off valve 7 7 to the nozzle 33 in Fig. 1. In this way, a swirl component is given to the inlet flow of the first stage impeller, the suction flow rate can be reduced, and the power of the compressor can be reduced. When the pressure downstream of the valve (P dp) drops below the blow-off valve closing set pressure (P d L), close the blow-off valve 77.

In conventional compressors, a throttle valve was provided on the suction side of the first stage compressor, and during blowing operation, the throttle valve was throttled to reduce compressor power. The compressor of the present embodiment does not require a throttle valve on the suction side, so that the number of parts can be reduced as compared with the conventional method. Also, the compressor can be packaged in a compact form, which is economically advantageous. It goes without saying that a suction valve may be provided in the above embodiment.

Another embodiment of the present invention will be described with reference to FIGS. 13 and 14. FIG. In the present embodiment, cooling air is blown into the inside of the rear labyrinth on the second stage side. A labyrinth 35 having an inner diameter larger than the inner diameter of the lapiling 30 formed on the back of the second-stage impeller is provided on the inner peripheral portion of the housing that holds the second-stage side 24 b of the thrust bearing. A cooling fluid blowing hole is formed between the labyrinth 35 and the lapiling 30 of the second stage impeller. Further In addition, a cooling fluid blowing hole 31b is formed between the magnetic bearing sensor 36 and the labyrinth 30 of the second stage impeller. The cooling hole 31b may be formed between the magnetic bearing 23b and the sensor 36. With this configuration, as shown in FIG. 14, cooling air is taken in from the downstream of the discharge cooler 23 b through the filter 75 and is introduced into the blowing hole 31 a of the casing 1. Guided <Thrust from the second stage to the first stage at the thrust force receiving part determined by the inner diameter of the back labyrinth 30 of the second stage impeller and the labyrinth inner diameter of the thrust bearing Force acts. As a result, it is possible to reduce the imbalance of the thrust force generated by the first-stage impeller and the second-stage impeller, and to realize a high-speed rotating shaft system in which the thrust bearing is reduced in size. In the present embodiment, since the amount of leakage from the rear lapel of the second-stage impeller can be reduced, there is also an effect that the power consumption of the compressor can be reduced.

As described above, in the present invention, there is no contact between the rotating part and the stationary part, and there is no need to use oil in any part of the compressor. A centrifugal compressor can be obtained. In addition, since no oil is used, a two-stage centrifugal compressor that is friendly to the global environment can be obtained. _(In addition, since the operating power can be reduced, the energy saving effect is large, and the generation of carbon dioxide to the global environment can be suppressed. Also, since the centrifugal compressor can be downsized at high speed, the installation area can be reduced.

Claims

Scope of Claims: A rotating shaft with a rotor of an electric motor formed in its intermediate portion, a centrifugal compressor impeller attached to both ends of the rotating shaft, and a steering wheel of the electric motor that constitutes an electric motor together with the rotor. In a two-stage centrifugal compressor that is directly driven by an electric motor having a radial magnetic bearing means and a thrust magnetic bearing means that rotatably support the rotary shaft, the rotary shaft is rotatably supported. Directly driven by an electric motor, the stator is provided with a water-cooled jacket for water-cooling the stator, and a gas flow path is formed in the middle of the stator for guiding the working gas sucked into the two-stage centrifugal compressor. Two-stage centrifugal compressor.

Impellers attached to both ends of the rotating shaft constitute a first compressor stage and a second compressor stage, and an intercooler cools the working gas compressed by these compressor stages. A discharge cooler is provided, and an integrated casing is provided that accommodates the intercooler and the discharge cooler together with the impeller, the radial magnetic bearing means, the thrust magnetic bearing, and the electric motor. A two-stage centrifugal compressor directly driven by the electric motor according to claim 1.

A cooling path is formed in the intercooler so that the working gas compressed in the first compressor stage flows from above, is cooled by the intercooler, and then flows out from below. A two-stage centrifugal compressor directly driven by the electric motor according to claim 2.

The working gas compressed by the second compressor stage flows into the discharge cooler from above, and after being cooled by this discharge cooler, it flows out from below. A two-stage centrifugal compressor directly driven by an electric motor according to claim 2, characterized in that a cooling path is formed in the compressor.

5. A rotating shaft with a rotor of an electric motor formed in an intermediate portion, an impeller attached to both ends of the rotating shaft, a stator of the electric motor that constitutes an electric motor together with the rotor, and the rotating shaft. In a two-stage centrifugal compressor directly driven by an electric motor equipped with a pair of rotatably supported radial magnetic bearing means and thrust magnetic bearing means, intercooling is performed to cool the working gas compressed by one of the impellers. and a discharge cooler for cooling the working gas compressed by the other impeller are arranged below the rotating shaft and almost parallel to the rotating shaft, and these rotating shafts, intercooler and discharge cooler The present invention is characterized by providing an integrated casing that houses a cooler, and forming a cooling path that guides the working gas cooled by either the discharge cooler or the intercooler to an intermediate portion of the stator of the electric motor. A two-stage centrifugal compressor that is directly driven by an electric motor.

6. A window is formed at the axial end of the casing that accommodates the intercooler and the discharge cooler, and a cover is removably attached to the window, so that the intercooler and the discharge cooler can be inserted from the axial direction. A two-stage centrifugal compressor directly driven by the electric motor according to claim 5, characterized in that:

7. The electric motor is a permanent magnet type synchronous motor, and the rotor includes a ring-shaped permanent magnet, a ring-shaped carbon fiber reinforced plastic holding member that holds the permanent magnet, and the ring-shaped holding member. 6. The two-stage centrifugal compressor directly driven by an electric motor according to claim 5, further comprising a metal sleeve that holds the permanent magnet together with the member.

8. The radial magnetic bearing means and the thrust magnetic bearing means are operated under load. Claim 5, further comprising a cooling switching means that selectively cools the working fluid with the working fluid cooled by the intercooler at times and with the working fluid cooled by the discharge cooler during no-load operation. A two-stage centrifugal compressor directly driven by the electric motor described above.

9. The rotor has two poles, and means for detecting the magnetic pole position based on the induced voltage generated in the winding of the motor's steering wheel when the rotor rotates; 6. The two-stage centrifugal compressor directly driven by the electric motor according to claim 5, further comprising: an inverter overnight control means for controlling the electric motor based on the magnetic pole position.

10. An inverter that drives the electric motor, a pressure detection means that detects the pressure on the downstream side of the second stage impeller, and an external device that operates if the detected pressure is higher than a predetermined pressure. A blow-off valve for blowing off gas and a control means for controlling the inverter overnight based on the pressure detected by the pressure detection means are provided, and the blow-off valve is configured to prevent the blow-off from occurring at a predetermined speed even when a set lower limit rotation speed is reached. When the detected pressure rises above the wind valve opening set pressure, the blowoff valve is opened, and a flow path is formed in which the working fluid that has passed through the blowoff valve is returned to the upstream side of the first stage impeller. A two-stage centrifugal compressor directly driven by an electric motor according to claim 5.

1 1. The motor is characterized by comprising a cooling switching means for selectively cooling the electric motor with the working fluid cooled by the intercooler during load operation and with the working fluid cooled by the discharge cooler during no-load operation. A two-stage centrifugal compressor directly driven by the electric motor according to claim 5.