WO2020129326A1

WO2020129326A1 - Turbo compressor and refrigeration cycle device

Info

Publication number: WO2020129326A1
Application number: PCT/JP2019/035115
Authority: WO
Inventors: 直芳庄山; 洪志孫; 文紀河野; 松井　大
Original assignee: パナソニック株式会社
Priority date: 2018-12-20
Filing date: 2019-09-06
Publication date: 2020-06-25
Also published as: JP2022028991A

Abstract

Provided is a turbo compressor comprising: a rotating shaft; a first impeller which is fixed to the rotating shaft and has a first suction port for suctioning an operating fluid and a first discharge port for discharging the operating fluid; a second impeller which is fixed to the rotating shaft and has a second suction port for suctioning the operating fluid and a second discharge port for discharging the operating fluid; a first bearing which is disposed between the first impeller and the second impeller in an axial direction of the rotating shaft and rotatably supports the rotating shaft; and a first flow path for guiding the operating fluid from the first discharge port to the second suction port. The first suction port and the second suction port are opened in the same direction.

Description

Turbo compressor and refrigeration cycle device

The present disclosure relates to a turbo compressor and a refrigeration cycle device.

Conventionally, a multi-stage turbo compressor for compressing low-pressure refrigerant is known.

Patent Document 1 describes a turbo compressor 200 as shown in FIG. The turbo compressor 200 includes a casing 221, an electric motor 213, a rotating shaft 225, a rolling bearing 227, and a slide bearing 228. Two stages of

impellers

223a and 223b are fixed to one end of the rotary shaft 225, and the

impellers

223a and 223b form a compression section 223 together with a compression passage structure. The rolling bearing 227 pivotally supports the rotary shaft 225 between the electric motor 213 and the

impellers

223a and 223b, and the other end of the rotary shaft 225 is pivotally supported by the slide bearing 228. The rolling bearing 227 has two

angular ball bearings

227a and 227b. In the plain bearing 228, the bearing metal 228a is press-fitted into the bearing boss 221d. The electric motor 213 includes a stator 213A and a rotor 213B fixed to the rotating shaft 225. When the compressor 223 is driven by the electric motor 213, the vaporized refrigerant is sucked into the compressor 223 and compressed.

On the other hand, conventionally, as a turbo compressor, a multi-stage centrifugal compressor having a return flow path is known. For example, Patent Document 2 describes a multistage centrifugal compressor 300 as shown in FIG. 7. The multi-stage centrifugal compressor 300 includes a first-stage impeller 301, a next-stage impeller 304, a diffuser 306, a return vane 307, a diffuser vane 308, a stage labyrinth 309, and a rotating shaft 310. The multi-stage centrifugal compressor 300 has a return flow passage through which the gas flow from the diffuser 306 flows into the next-stage impeller 304, and the return vane 307 is arranged in the return flow passage. In FIG. 7, white arrows indicate the main flow of gas. The return flow passage converts the velocity energy of gas into pressure energy. In FIG. 7, C is the axis of the rotating shaft 310.

JP, 2017-125434, A JP-A-7-127600

According to the techniques described in Patent Documents 1 and 2, for example, when compressing a working fluid such as a vapor phase refrigerant having a saturated vapor pressure equal to or lower than atmospheric pressure at room temperature (20° C.±15° C.: Japanese Industrial Standard JISZ8703) , There is room to increase the stability of rotation of the rotating shaft.

The present disclosure provides a turbo compressor that is advantageous for increasing the stability of rotation of the rotating shaft.

The present disclosure includes a first impeller fixed to the rotary shaft, the first impeller having a rotary shaft, a first suction port fixed to the rotary shaft, into which the working fluid is sucked, and a first discharge port discharging the working fluid. A second impeller having a second suction port for sucking the working fluid and a second discharge port for discharging the working fluid, and between the first impeller and the second impeller in the axial direction of the rotating shaft. And a first bearing that rotatably supports the rotating shaft,
A first flow path for guiding the working fluid from the first discharge port to the second suction port, wherein the first suction port and the second suction port are opened in the same direction, provide.

The above turbo compressor is advantageous for improving the stability of rotation of the rotating shaft.

FIG. 1 is a sectional view showing an example of a turbo compressor of the present disclosure. FIG. 2 is a configuration diagram illustrating an example of the refrigeration cycle device of the present disclosure. FIG. 3 is a cross-sectional view showing another example of the turbo compressor of the present disclosure. FIG. 4 is a sectional view showing a part of another example of the turbo compressor of the present disclosure. FIG. 5 is a cross-sectional view showing a turbo compressor according to a reference example. FIG. 6 is a sectional view showing a conventional turbo compressor. FIG. 7 is a sectional view showing a conventional multistage centrifugal compressor.

(Findings that form the basis of this disclosure)
For example, it is conceivable to compress a working fluid having a saturated vapor pressure equal to or lower than atmospheric pressure at room temperature using a multi-stage turbo compressor having a return passage. In this case, it is necessary to increase the cross-sectional area of the return channel. This is because such working fluid flows through the return flow passage with a large specific volume, and the volume flow rate of the working fluid in the return flow passage is large. On the other hand, in order to increase the cross-sectional area of the return flow channel, it is necessary to lengthen the axial length of the rotary shaft. However, this tends to reduce the natural frequency based on the bending rigidity of the rotating shaft. When the natural frequency based on the bending rigidity of the rotating shaft is low, the natural frequency based on the bending rigidity of the rotating shaft and the rotating speed of the rotating shaft become close to each other, causing resonance, and rotating the rotating shaft at a desired rotating speed Becomes difficult. Therefore, in order to prevent the natural frequency from decreasing due to the bending rigidity of the rotary shaft, it is desirable to suppress the extension of the length of the rotary shaft in the axial direction. Therefore, the present inventors have earnestly studied the structure of a turbo compressor for increasing the cross-sectional area of the return flow channel while suppressing the extension of the axial length of the rotating shaft. As a result, the present inventors have newly found a desired positional relationship between the bearing that rotatably supports the rotating shaft and the return passage, and have devised the turbo compressor of the present disclosure. It should be noted that this finding is based on the study of the present inventors and is not admitted as prior art.

(Outline of One Aspect According to the Present Disclosure)
A turbo compressor according to a first aspect of the present disclosure includes a rotating shaft, a first suction port fixed to the rotating shaft, into which a working fluid is sucked, and a first discharge port from which the working fluid is discharged. A second impeller fixed to the rotary shaft, having a second suction port for sucking the working fluid and a second discharge port for discharging the working fluid, and the first impeller in the axial direction of the rotary shaft. A first bearing that is disposed between an impeller and the second impeller and rotatably supports the rotating shaft, and a first flow path that guides the working fluid from the first discharge port to the second suction port. The first suction port and the second suction port are opened in the same direction.

According to the first aspect, since the first bearing is arranged between the first impeller and the second impeller in the axial direction of the rotary shaft, the first flow is suppressed while suppressing the extension of the length of the rotary shaft in the axial direction. It is easy to increase the cross-sectional area of the road. As a result, it is possible to prevent the natural frequency from decreasing due to the bending rigidity of the rotating shaft. Therefore, the turbo compressor according to the first aspect is advantageous in increasing the stability of rotation of the rotary shaft.

In the second aspect of the present disclosure, for example, in the turbo compressor according to the first aspect, all of the region in which the first bearing exists in the axial direction overlaps with the region in which the first flow path exists in the axial direction. ing. According to the second aspect, it is possible to more reliably suppress the extension of the length of the rotating shaft in the axial direction.

In the third aspect of the present disclosure, for example, in the turbo compressor according to the first aspect or the second aspect, the working fluid having a saturated vapor pressure of atmospheric pressure or lower at room temperature may be accelerated and compressed. According to the third aspect, it is possible to compress the working fluid having a saturated vapor pressure of atmospheric pressure or less at room temperature in a state where the rotation of the rotary shaft is highly stable.

In the fourth aspect of the present disclosure, for example, the turbo compressor according to the first aspect to the third aspect is arranged at a position further away from the first impeller than the first bearing in the axial direction, and the rotation shaft is It may further include a second bearing rotatably supported, and a motor having a rotor fixed to the rotating shaft between the first bearing and the second bearing in the axial direction. According to the fourth aspect, it is possible to enhance the stability of rotation of the rotary shaft while the rotor of the motor is fixed to the rotary shaft.

In the fifth aspect of the present disclosure, for example, the turbo compressor according to the first aspect to the fourth aspect has a third suction port that opens in the same direction as the first suction port and the second suction port. And a third impeller arranged at a position apart from the first bearing in the axial direction and fixed to the rotating shaft, wherein the second impeller includes the first impeller and the third impeller. The rotary shaft may be disposed between the rotary shaft and the impeller, and the diameter of the rotary shaft at the fixed position of the third impeller may be equal to or larger than the diameter of the rotary shaft at the fixed position of the second impeller. According to the fifth aspect, it is easy to realize operation at a high pressure ratio by the third impeller. In addition, since the diameter of the rotary shaft at the fixed position of the third impeller is equal to or larger than the diameter of the rotary shaft at the fixed position of the second impeller, the bending rigidity of the rotary shaft tends to increase. Therefore, according to the fifth aspect, it is possible to prevent the natural frequency from decreasing due to the bending rigidity of the rotating shaft.

In the sixth aspect of the present disclosure, for example, the turbo compressor according to the fifth aspect is disposed at a position further away from the first impeller than the first bearing in the axial direction, and rotatably supports the rotary shaft. And a motor having a rotor fixed to the rotating shaft between the first bearing and the second bearing in the axial direction, the third impeller may be provided. , May be fixed between the second impeller and the rotor in the front axis direction. According to the sixth aspect, it is possible to enhance the stability of rotation of the rotating shaft while the rotor of the motor is fixed to the rotating shaft.

A refrigeration cycle apparatus according to a seventh aspect of the present disclosure is an evaporator that produces a vapor-phase refrigerant, and compresses the vapor-phase refrigerant produced in the evaporator as the working fluid. The turbo compressor according to any one of the aspects and a condenser for condensing the gas-phase refrigerant compressed by the turbo compressor.

According to the seventh aspect, the gas-phase refrigerant generated in the evaporator can be compressed in a state in which the rotation of the rotary shaft of the turbo compressor is highly stable.

In the eighth aspect of the present disclosure, for example, in the refrigeration cycle apparatus according to the seventh aspect, the evaporator may store the liquid-phase refrigerant therein, and the condenser may store the liquid-phase refrigerant therein. Alternatively, at least one of the liquid-phase refrigerant stored in the evaporator and the liquid-phase refrigerant stored in the condenser may be supplied to the first bearing. According to the eighth aspect, for example, at least one of the liquid-phase refrigerant stored in the evaporator and the liquid-phase refrigerant stored in the condenser can be used for lubricating or cooling the first bearing.

Hereinafter, embodiments of the present disclosure will be described with reference to the drawings. Note that the following description relates to an example of the present disclosure, and the present disclosure is not limited to the following embodiments.

(Embodiment 1)
As shown in FIG. 1, the turbo compressor 1a includes a rotary shaft 10, a first impeller 21, a second impeller 22, a first bearing 41, and a return flow passage 82a. The first impeller 21 is fixed to the rotating shaft 10. The first impeller 21 has a first suction port 21a for sucking the working fluid and a first discharge port 21b for discharging the working fluid. The second impeller 22 is fixed to the rotating shaft 10. The second impeller 22 has a second suction port 22a for sucking the working fluid and a second discharge port 22b for discharging the working fluid. The first bearing 41 is arranged between the first impeller 21 and the second impeller 22 in the axial direction of the rotary shaft, and rotatably supports the rotary shaft 10.

The first bearing 41 is an essential member for rotatably supporting the rotary shaft 10. The first bearing 41 is distinguished from the seal member which is not indispensable for rotatably supporting the rotating shaft 10 in this respect. The first suction port 21a and the second suction port 22a open in the same direction.

As shown in FIG. 1, the return flow passage 82a is arranged so as to approach the rotary shaft 10 from the first discharge port 21b, and is connected to the second suction port 22a. The return channel 82a may be called the first channel.

The first flow path is a flow path that guides the working fluid from the first discharge port 21b to the second suction port 22a, and may be a flow path through which the working fluid passes through pipes or structures.

For example, the entire region where the first bearing 41 exists in the axial direction of the rotating shaft 10 overlaps the region where the return flow passage 82a exists in the axial direction of the rotating shaft 10.

The first impeller 21 is fixed to one end of the rotary shaft 10, for example. Since the first bearing 41 is arranged between the first impeller 21 and the second impeller 22 in the axial direction of the rotary shaft 10, the first impeller 21 is fixed to one end of the rotary shaft 10. Also, the first bearing 41 can be rotatably supported.

The turbo compressor 1a includes, for example, a casing 60, and the first bearing 41 is fixed to the casing 60. The casing 60 has a wall surface that forms a stationary flow channel, and the return flow channel 82a exists inside the casing 60.

As shown in FIG. 1, the turbo compressor 1 a further includes, for example, a motor 70. The motor 70 includes, for example, a rotor 71 and a stator 72. The rotor 71 is fixed to the rotating shaft 10, for example. The rotor 71 is preferably fixed to the rotary shaft 10 at a position farther from the first impeller 21 than the second impeller 22 in the axial direction of the rotary shaft 10. In other words, the first impeller 21, the second impeller 22, and the rotor 71 are arranged in this order in the axial direction of the rotary shaft 10. The stator 72 is fixed to the casing 60, for example. When electric power is supplied to the motor 70, a rotating magnetic field is generated by the stator 72, and a rotating torque is generated in the rotor 71. As a result, the first impeller 21 and the second impeller 22 fixed to the rotating shaft 10 rotate.

The turbo compressor 1a accelerates and compresses a working fluid having a saturated vapor pressure equal to or lower than atmospheric pressure at room temperature, for example. The working fluid is typically in the gas phase. Water may be preferably used as such a working fluid. Since such a working fluid easily flows through the return passage with a large specific volume, it is highly necessary to increase the cross-sectional area of the return passage. In some cases, the turbo compressor 1a may accelerate and compress a working fluid having a saturated vapor pressure exceeding atmospheric pressure at room temperature.

The turbo compressor 1a further includes a second bearing 42, for example. The second bearing 42 is arranged at a position farther from the first impeller 21 than the first bearing 41 in the axial direction of the rotating shaft 10. The second bearing 42 rotatably supports the rotating shaft 10. The rotor 71 of the motor 70 is fixed to the rotary shaft 10 between the first bearing 41 and the second bearing 42 in the axial direction of the rotary shaft 10, for example. The center of the rotary shaft 10 in the axial direction is located, for example, between the second impeller 22 and the rotor 71.

The rotary shaft 10 desirably has a highly symmetrical axisymmetric shape in order to minimize unbalance during rotation. The rotor 71 is fixed to the rotary shaft 10 by shrink fitting, for example. A sufficiently large shrink fit margin is secured in the shrink fit of the rotor 71 so that the shrink fit margin remains even if the rotary shaft 10 rotates at a high speed and the diameter of the rotor 71 is slightly increased by centrifugal force. .. The rotation speed of the rotary shaft 10 is, for example, 5000 rpm (revolutions per minute) to 100000 rpm. The material of the rotating shaft 10 is not limited to a particular material. The material of the rotating shaft 10 may be, for example, a high-strength iron-based material such as SNCM630.

For example, the diameter of the rotary shaft 10 at the fixed position of the second impeller 22 is larger than the diameter of the rotary shaft 10 at the fixed position of the first impeller 21. In this case, the rotary shaft 10 has a large diameter near the center, and the natural frequency based on the bending rigidity of the rotary shaft 10 tends to be high. Therefore, the rotation stability of the rotary shaft 10 is likely to increase.

In the turbo compressor 1a, the first impeller 21 and the second impeller 22 typically form a centrifugal compression mechanism. The outer diameter of the portion of the first impeller 21 forming the first suction port 21a is smaller than the outer diameter of the portion of the first impeller 21 forming the first discharge port 21b, for example. The first suction port 21a is open, for example, in front of the first impeller 21. The first discharge port 21b opens to the outside of the first impeller 21 in the radial direction. The outer diameter of the portion of the second impeller 22 forming the second suction port 22a is smaller than the outer diameter of the portion of the second impeller 22 forming the second discharge port 22b. The second suction port 22a is open, for example, in front of the second impeller 22. The second discharge port 22b opens to the outside of the second impeller 22 in the radial direction.

The return passage 82a is, for example, an annular passage located outside the first bearing 41 in the direction perpendicular to the axis of the rotating shaft 10. The outer diameter of the portion of the first impeller 21 forming the first discharge port 21b is smaller than the outer diameter of the portion of the second impeller 22 forming the second suction port 22a, for example. Therefore, the return passage 82a is formed so as to narrow toward the second impeller 22 in the axial direction of the rotating shaft 10.

The working fluid sucked from the first suction port 21a is accelerated by the first impeller 21 and discharged from the first discharge port 21b. After that, the working fluid passes through the return flow path 82a, is sucked into the second suction port 22a, is accelerated by the second impeller 22, and is discharged from the second discharge port 22b. Thereby, the working fluid is compressed. A flow path 83 is connected to the turbo compressor 1a, and the working fluid discharged from the second discharge port 22b is supplied to the outside of the turbo compressor 1a through the flow path 83.

The first bearing 41 is not limited to a specific bearing as long as it supports the rotating shaft 10 in a rotatable manner. The first bearing 41 is, for example, a slide bearing. The first bearing 41 has, for example, a radial support portion 41a and a thrust support portion 41b. The radial support portion 41a supports the load of the rotating body including the rotating shaft 10 in the radial direction. The thrust support portion 41b supports the thrust force generated by the rotation of the first impeller 21 and the second impeller 22. When the first bearing 41 is a slide bearing, the fluid that lubricates the first bearing 41 may be a lubricating oil or a fluid that is compatible with the working fluid. The fluid that lubricates the first bearing 41 may desirably be a fluid having the same composition as the working fluid. The first bearing 41 may be a rolling bearing such as a ball bearing or a magnetic bearing.

The second bearing 42 is not limited to a particular bearing as long as it supports the rotating shaft 10 in a rotatable manner. The second bearing 42 is, for example, a slide bearing. The second bearing 42 is configured, for example, so as not to support thrust force during normal operation of the turbo compressor 1a. The second bearing 42 allows, for example, in the axial direction of the rotary shaft 10, a dimensional change of about one thousandth or more of the entire length of the rotary body including the rotary shaft 10 in the axial direction of the rotary shaft 10. Thereby, even if the rotary shaft 10 expands due to the temperature rise of the rotary shaft 10 during the operation of the turbo compressor 1a, the second bearing 42 can cope with the expansion of the rotary shaft 10 due to the expansion of the rotary shaft 10. The second bearing 42 may be a rolling bearing such as a ball bearing or a magnetic bearing.

An example of the operation and action of the turbo compressor 1a will be described. In the space in front of the first suction port 21a of the first impeller 21, for example, a vapor-phase working fluid having a pressure of 1 kPa exists. When the rotating shaft 10 rotates due to the operation of the motor 70, the first impeller 21 and the second impeller 22 rotate together with the rotating shaft 10. As a result, the working fluid in the space in front of the first suction port 21a passes through the first suction port 21a, is accelerated by the first impeller 21, and is discharged from the first discharge port 21b. Thereby, the working fluid is compressed. The pressure ratio of the working fluid in the first impeller 21 may be 2, for example, and the pressure of the working fluid discharged from the first discharge port 21b may be 2 kPa. The working fluid discharged from the first discharge port 21b is sucked into the second suction port 22a through the return flow passage 82a. For example, when the working fluid is steam, the specific volume of steam having a pressure of 2 kPa at room temperature is 67 m3/kg, and the return flow passage 82a requires a large cross-sectional area. In the turbo compressor 1 a, the first bearing 41 is arranged between the first impeller 21 and the second impeller 22 in the axial direction of the rotary shaft 10. Therefore, the rotary shaft 10 originally has a length for disposing the first bearing 41 between the first impeller 21 and the second impeller 22. In the turbo compressor 1a, the first bearing 41 overlaps the return flow passage 82a in the axial direction of the rotary shaft 10, so that the cross-sectional area of the return flow passage 82a is suppressed while suppressing the extension of the axial length of the rotary shaft 10. It is easy to make Therefore, the natural frequency based on the bending rigidity of the rotary shaft 10 tends to increase. As a result, according to the turbo compressor 1a, it is possible to prevent resonance due to the natural frequency based on the bending rigidity of the rotating shaft 10 and the rotating speed of the rotating shaft 10 becoming close to each other.

Note that the meaning of the cross-sectional area of the return flow passage 82a is also meant to include the cross-sectional area A having a constant radius from the rotation axis as shown in FIG. In the present disclosure, since the return passage 82a is held by the first bearing 41, the cross-sectional area A of the return passage 82a can be expanded. As a result, the high-pressure vapor-phase refrigerant can be passed inside the return flow passage 82a.

A working fluid may be used as the lubricant of the first bearing 41. For example, as for the supply of the working fluid to the first bearing 41, in FIG. 1, a passage passing through the inside of the rotating shaft 10 from the right side of the rotating shaft 10 to the first bearing 41 is provided, and the working fluid is supplied to the first bearing through the passage. 41.

For example, a refrigeration cycle device can be provided using the turbo compressor 1a. As shown in FIG. 2, the refrigeration cycle device 100 includes an evaporator 30, a turbo compressor 1 a, and a condenser 50. The evaporator 30 produces a gas-phase refrigerant. The turbo compressor 1a compresses the vapor phase refrigerant generated in the evaporator 30 as a working fluid. The condenser 50 condenses the vapor phase refrigerant compressed by the turbo compressor 1a. According to the refrigeration cycle apparatus 100, the vapor-phase refrigerant generated in the evaporator 30 can be compressed in a state where the rotation shaft 10 of the turbo compressor 1a is highly stable in rotation. The refrigeration cycle apparatus 100 can be made to function as an air conditioner, for example.

The evaporator 30 stores, for example, a liquid-phase refrigerant therein. The condenser 50 stores the liquid phase refrigerant therein, for example. In the refrigeration cycle apparatus 100, at least one of the liquid-phase refrigerant stored in the evaporator 30 and the liquid-phase refrigerant stored in the condenser is supplied to the first bearing 41. In this case, for example, at least one of the liquid-phase refrigerant stored in the evaporator 30 and the liquid-phase refrigerant stored in the condenser 50 can be used for lubricating or cooling the first bearing 41.

As shown in FIG. 2, the refrigeration cycle apparatus 100 includes, for example, a main circuit 80, a first circulation circuit 31, and a second circulation circuit 51. The main circuit 80 is a circuit in which the evaporator 30, the turbo compressor 1a, and the condenser 50 are connected in this order. A flow passage 81 connects the internal space of the evaporator 30 and the internal space of the casing 60 of the turbo compressor 1 a. A flow passage 83 connects the internal space of the casing 60 of the turbo compressor 1 a and the internal space of the condenser 50. The internal space of the condenser 50 and the internal space of the evaporator 30 communicate with each other through a flow path 84.

The first circulation circuit 31 has a first pump 35 and a first heat exchanger 33. In the first circulation circuit 31, the liquid working fluid stored in the evaporator 30 is supplied to the first heat exchanger 33 by the first pump 35, and the working fluid absorbed in the first heat exchanger 33 is supplied to the evaporator 30. It is configured to return. In the first circulation circuit 31, the evaporator 30 and the inlet of the first pump 35 are connected by a flow path 31a. The outlet of the first pump 35 and the inlet of the first heat exchanger 33 are connected by the flow path 31b. The outlet of the first heat exchanger 33 and the evaporator 30 are connected by the flow path 31c.

The second circulation circuit 51 has a second pump 55 and a second heat exchanger 53. In the second circulation circuit 51, the liquid working fluid stored in the condenser 50 is supplied to the second heat exchanger 53 by the second pump 55, and the working fluid radiated by the second heat exchanger 53 is fed to the condenser 50. It is configured to return. In the second circulation circuit 51, the condenser 50 and the inlet of the second pump 55 are connected by the flow path 51a. The outlet of the second pump 55 and the inlet of the second heat exchanger 53 are connected by the flow path 51b. The outlet of the second heat exchanger 53 and the condenser 50 are connected by the flow path 51c.

In the refrigeration cycle device 100, for example, the liquid-phase refrigerant stored in the evaporator 30 is used as the lubricating liquid to be supplied to the first bearing 41 and the second bearing 42. In the refrigeration cycle apparatus 100, for example, the liquid-phase refrigerant stored in the condenser 50 may be used as the lubricating liquid to be supplied to the first bearing 41 and the second bearing 42. The liquid-phase refrigerant supplied to the first bearing 41 and the second bearing 42 is returned to the evaporator 30 or the condenser 50.

The turbo compressor 1a can be changed from various viewpoints. For example, the turbo compressor 1a may be changed to the turbo compressor 1b shown in FIG. The turbo compressor 1b is configured in the same manner as the turbo compressor 1a, except for the part particularly described. The constituents of the turbo compressor 1b that are the same as or correspond to the constituents of the turbo compressor 1a are designated by the same reference numerals, and detailed description thereof will be omitted. The description of the turbo compressor 1a also applies to the turbo compressor 1b unless technically contradictory.

As shown in FIG. 3, the turbo compressor 1b further includes a third impeller 23. The third impeller 23 is arranged at a position apart from the first bearing 41 in the axial direction of the rotary shaft 10 and is fixed to the rotary shaft 10. The second impeller 22 is arranged between the first impeller 21 and the third impeller 23. According to the turbo compressor 1b, it is easy to realize the operation at a high pressure ratio by the third impeller 23. In addition, in the turbo compressor 1b, the diameter of the rotary shaft 10 at the fixed position of the third impeller 23 is equal to or larger than the diameter of the rotary shaft 10 at the fixed position of the second impeller 22. Thereby, the rotating shaft 10 tends to have a large diameter near the center in the axial direction, and the bending rigidity of the rotating shaft 10 tends to increase. As a result, it is possible to prevent the natural frequency based on the bending rigidity of the rotating shaft 10 from decreasing, and prevent the natural frequency based on the bending rigidity of the rotating shaft 10 and the rotation speed of the rotating shaft 10 from becoming close to each other to cause resonance. it can.

The third impeller 23 is fixed to the rotary shaft 10 between the second impeller 22 and the rotor 71 in the axial direction of the rotary shaft 10, for example. In this case, the stability of rotation of the rotary shaft 10 can be enhanced with the rotor 71 of the motor 70 fixed to the rotary shaft 10. The center of the rotary shaft 10 in the axial direction is located, for example, between the second impeller 22 and the rotor 71.

The third impeller 23 typically has a third suction port 23a for sucking the working fluid and a third discharge port 23b for discharging the working fluid. The third suction port 23a opens in the same direction as the opening direction of the first suction port 21a and the second suction port 22a, for example. The third impeller 23 constitutes, for example, a centrifugal compression mechanism together with the first impeller 21 and the second impeller 22. The outer diameter of the portion of the third impeller 23 forming the third suction port 23a is smaller than the outer diameter of the portion of the third impeller 23 forming the third discharge port 23b, for example. The third suction port 23a opens, for example, in front of the third impeller 23. The third discharge port 23b opens to the outside in the radial direction of the third impeller 23.

The turbo compressor 1b further includes, for example, a return flow path 82b. The return flow path 82b is a flow path for guiding the working fluid from the second discharge port 22b to the third suction port 23a. As a result, the working fluid having a small specific volume that has been compressed by passing through the first impeller 21 and the second impeller 22 is guided to the third impeller 23. Therefore, the third impeller 23 can be manufactured so that the width of the working fluid passage in the third impeller 23 is narrower than the width of the working fluid passage in the first impeller 21.

(Embodiment 2)
The difference from the first embodiment is that the first bearing 41 has a load capacity of the thrust load of the rotary shaft 10.

When the turbo compressor 1a is operated for a long time, the temperature of the rotating shaft 10 rises. As a result, the total length of the rotary shaft 10 increases. Since the first bearing 41 has a load capacity of the thrust load of the rotary shaft 10, the positional relationship between the first bearing 41 and the rotary shaft 10 hardly changes in the axial direction of the rotary shaft 10. The influence of thermal expansion of the rotary shaft 10 becomes more serious as the distance from the first bearing 41 having the load capacity of thrust load increases. The first bearing 41 is arranged between the first impeller 21 and the second impeller 22 in the axial direction of the rotating shaft 10. Therefore, it is easy to shorten the distance between the first impeller 21 and the first bearing 41 in the axial direction of the rotating shaft 10 and the distance between the second impeller 22 and the first bearing 41 in the axial direction of the rotating shaft 10. Therefore, the amount of movement of the first impeller 21 due to thermal expansion of the rotating shaft 10 in the axial direction of the rotating shaft 10 and the amount of movement of the second impeller 22 due to thermal expansion of the rotating shaft 10 in the axial direction of the rotating shaft 10 can be reduced. As a result, it is possible to set a small clearance design value to be considered for preventing contact between the first impeller 21 and a member such as a shroud or a diffuser ring arranged near the first impeller 21. In addition, the design value of the clearance to be considered for preventing contact between the second impeller 22 and a member such as a shroud or a diffuser ring arranged near the second impeller 22 can be set small.

The first bearing 41 is not limited to a particular bearing as long as it has a load capacity of the thrust load of the rotary shaft 10. The first bearing 41 is, for example, a slide bearing. In this case, the first bearing 41 has a bearing surface 41b. The bearing surface 41b is formed so as to form a projection surface when projected on a plane perpendicular to the axis of the rotating shaft 10. The first bearing 41 may be a rolling bearing such as a ball bearing or a magnetic bearing.

The first bearing 41 may have a load capacity of the radial load of the rotary shaft 10. For example, the first bearing 41 has a radial bearing surface 41a that supports the rotating shaft 10 in the radial direction. The first bearing 41 may be configured to have a bearing surface that supports the rotating shaft 10 in the thrust direction and the radial direction. Such a bearing surface can be formed, for example, by forming a tapered hole in the first bearing 41.

The second bearing 42 is not limited to a particular bearing as long as it supports the rotating shaft 10 in a rotatable manner. The second bearing 42 is, for example, a slide bearing. The second bearing 42 is configured, for example, so as not to support thrust force during normal operation of the turbo compressor 1a. The second bearing 42 allows, for example, in the axial direction of the rotary shaft 10, a dimensional change of about one thousandth or more of the entire length of the rotary body including the rotary shaft 10 in the axial direction of the rotary shaft 10. Thereby, even if the rotary shaft 10 thermally expands due to the temperature rise of the rotary shaft 10 during the operation of the turbo compressor 1a, the second bearing 42 can cope with the thermal expansion of the rotary shaft 10. The second bearing 42 may be a rolling bearing such as a ball bearing or a magnetic bearing.

When the first bearing 41 is a slide bearing, the fluid that lubricates the first bearing 41 may be lubricating oil or fluid that is compatible with the refrigerant. The fluid that lubricates the first bearing 41 may desirably be a fluid having the same composition as the refrigerant.

As shown in FIG. 1, the turbo compressor 1 a further includes, for example, a motor 70. The motor 70 includes, for example, a rotor 71 and a stator 72. The rotor 71 is fixed to the rotary shaft 10 between the first bearing 41 and the second bearing 42 in the axial direction of the rotary shaft 10, for example.

The stator 72 is fixed to the casing 60, for example. When electric power is supplied to the motor 70, a rotating magnetic field is generated by the stator 72, and a rotating torque is generated in the rotor 71. As a result, the first impeller 21 and the second impeller 22 fixed to the rotating shaft 10 rotate. When the turbo compressor 1a is operated for a long period of time, the rotor 71 generates heat and the temperature of the rotating shaft 10 rises.

The turbo compressor 1a accelerates and compresses a refrigerant having a saturated vapor pressure equal to or lower than atmospheric pressure at room temperature, for example. In the turbo compressor 1a, the refrigerant is typically in the gas phase. Water can be preferably used as such a refrigerant. Since such a refrigerant easily flows through the return passage with a large specific volume, it is highly necessary to increase the cross-sectional area of the return passage. In some cases, the turbo compressor 1a may accelerate and compress a refrigerant having a saturated vapor pressure exceeding atmospheric pressure at room temperature.

Since the first bearing 41 is arranged between the first impeller 21 and the second impeller 22 in the axial direction of the rotary shaft 10, the return passage 82a is suppressed while suppressing the extension of the rotary shaft 10 in the axial direction. It is easy to increase the cross-sectional area of. As a result, the natural frequency based on the bending rigidity of the rotating shaft 10 does not easily decrease. If the cross-sectional area of the return passage 82a is large, the refrigerant easily flows through the return passage 82a with a large specific volume. For example, when the turbo compressor 1a compresses a refrigerant having a saturated vapor pressure equal to or lower than the atmospheric pressure at room temperature, it is easy to flow the refrigerant with a large specific volume in the return flow passage 82a.

The rotary shaft 10 desirably has a highly symmetrical axisymmetric shape in order to minimize unbalance during rotation. The rotor 71 is fixed to the rotary shaft 10 by shrink fitting, for example. A sufficiently large shrink fit margin is secured in the shrink fit of the rotor 71 so that the shrink fit margin remains even if the rotary shaft 10 rotates at a high speed and the diameter of the rotor 71 is slightly increased by centrifugal force. .. The rotation speed of the rotary shaft 10 is, for example, 5000 rpm to 100000 rpm. The material of the rotating shaft 10 is not limited to a particular material. The material of the rotating shaft 10 may be, for example, a high-strength iron-based material such as SNCM630.

The refrigerant sucked from the first suction port 21a is accelerated by the first impeller 21 and discharged from the first discharge port 21b. After that, the refrigerant passes through the return flow path 82a, is sucked into the second suction port 22a, is accelerated by the second impeller 22, and is discharged from the second discharge port 22b. As a result, the refrigerant is compressed. A flow path 83 is connected to the turbo compressor 1a, and the refrigerant discharged from the second discharge port 22b is supplied to the outside of the turbo compressor 1a through the flow path 83.

An example of the operation and action of the turbo compressor 1a will be described. In the space in front of the first suction port 21a of the first impeller 21, for example, a vapor-phase refrigerant having a pressure of 1 kPa exists. When the rotating shaft 10 rotates due to the operation of the motor 70, the first impeller 21 and the second impeller 22 rotate together with the rotating shaft 10. As a result, the refrigerant in the space in front of the first suction port 21a passes through the first suction port 21a, is accelerated by the first impeller 21, and is discharged from the first discharge port 21b. As a result, the refrigerant is compressed. The pressure ratio of the refrigerant in the first impeller 21 may be 2, for example, and the pressure of the refrigerant discharged from the first discharge port 21b may be 2 kPa.

The refrigerant discharged from the first discharge port 21b is sucked into the second suction port 22a through the return flow passage 82a. Further, the refrigerant is accelerated by the second impeller 22 and discharged from the second discharge port 22b. Thereby, the refrigerant is further compressed. The pressure ratio of the refrigerant in the second impeller 22 may be 2, for example, and the pressure of the refrigerant discharged from the second discharge port 22b may be 4 kPa. The refrigerant compressed in this way is guided to the outside of the turbo compressor 1a through the flow path 83.

A pressure difference of, for example, 1 kPa may occur between the front surface of the first impeller 21 including the first suction port 21a and the back surface opposite to the front surface of the first impeller 21. In addition, a pressure difference of, for example, 2 kPa may occur between the front surface of the second impeller 22 including the second suction port 22a and the back surface of the second impeller 22 opposite to the front surface. Due to these pressure differences, a thrust force is generated on the rotary shaft 10 from right to left in FIG.

When the turbo compressor 1a is operated for a long period of time, the temperature of the rotating shaft 10 rises due to the heat generation of the rotor 71, and the rotating shaft 10 thermally expands to increase its total length. At this time, the positional relationship between the first bearing 41 that receives the thrust force of the rotating shaft 10 and the rotating shaft 10 hardly changes. On the other hand, the second bearing 42 allows a predetermined dimensional change in the axial direction of the rotary shaft 10, and the positional relationship between the second bearing 42 and the rotary shaft 10 changes. Thereby, the thermal expansion of the rotating shaft 10 can be dealt with.

For example, a refrigeration cycle device can be provided using the turbo compressor 1a. As shown in FIG. 2, the refrigeration cycle device 100 includes an evaporator 30, a turbo compressor 1 a, and a condenser 50. The refrigerant evaporates in the evaporator 30. The turbo compressor 1a compresses the refrigerant evaporated in the evaporator 30. The condenser 50 condenses the refrigerant compressed by the turbo compressor 1a. Since the refrigeration cycle device 100 includes the turbo compressor 1a, the refrigerant evaporated in the evaporator 30 can be compressed under a high pressure ratio or a high adiabatic efficiency. Therefore, the refrigeration cycle apparatus is likely to exhibit a high coefficient of performance (COP).

The refrigeration cycle apparatus 100 can function as an air conditioner, for example.

The first circulation circuit 31 has a first pump 35 and a first heat exchanger 33. In the first circulation circuit 31, the liquid refrigerant stored in the evaporator 30 is supplied to the first heat exchanger 33 by the first pump 35, and the refrigerant absorbed in the first heat exchanger 33 returns to the evaporator 30. Is configured.

In the first circulation circuit 31, the evaporator 30 and the inlet of the first pump 35 are connected by a flow path 31a. The outlet of the first pump 35 and the inlet of the first heat exchanger 33 are connected by the flow path 31b. The outlet of the first heat exchanger 33 and the evaporator 30 are connected by the flow path 31c.

The second circulation circuit 51 has a second pump 55 and a second heat exchanger 53. In the second circulation circuit 51, the liquid refrigerant stored in the condenser 50 is supplied to the second heat exchanger 53 by the second pump 55, and the refrigerant radiated by the second heat exchanger 53 returns to the condenser 50. Is configured.

In the second circulation circuit 51, the condenser 50 and the inlet of the second pump 55 are connected by the flow path 51a. The outlet of the second pump 55 and the inlet of the second heat exchanger 53 are connected by the flow path 51b. The outlet of the second heat exchanger 53 and the condenser 50 are connected by the flow path 51c.

In the refrigeration cycle apparatus 100, for example, the liquid-phase refrigerant stored in the evaporator 30 is used as a lubricant to be supplied to the first bearing 41 and the second bearing 42. In the refrigeration cycle apparatus 100, for example, the liquid-phase refrigerant stored in the condenser 50 may be used as a lubricant to be supplied to the first bearing 41 and the second bearing 42. The liquid-phase refrigerant supplied to the first bearing 41 and the second bearing 42 is returned to the evaporator 30 or the condenser 50.

As shown in FIG. 4, in the turbo compressor 1b, the first bearing 41 is a slide bearing having a bearing surface 41b. The bearing surface 41b forms a projection surface when projected on a plane perpendicular to the axis of the rotating shaft 10. The rotating shaft 10 has a first flow path 12 and a second flow path 14. The first flow passage 12 extends along the axis inside the rotary shaft 10, and the lubricant flows through the first flow passage 12. The second flow path 14 connects the space in contact with the bearing surface 41b with the first flow path 12. For example, at least a part of the second flow path 14 extends in the radial direction of the rotating shaft 10.

The lubricant flowing through the first flow path 12 may be a lubricating oil, a fluid having compatibility with the refrigerant, or a fluid containing the same component as the refrigerant.

As shown in FIG. 4, for example, in the radial gap R between the radial support surface 41a of the first bearing 41 and the outer surface of the rotary shaft 10, the lubricant supplied from the second flow path 14 exists. There is. Further, in the axial gap T between the rotary shaft 10 and the bearing surface 41b of the first bearing 41, the lubricant that has passed through the radial gap R is present. The lubricant present in the axial gap T flows in the radial direction due to the pressure in the radial gap R and the rotation of the rotary shaft 10. At this time, for example, the pressure loss of the lubricant flow in the flow path of the axial gap T is smaller than the pressure loss of the lubricant flow in the flow path of the radial gap R when the axial gap T is sufficiently wide. .. In this case, the pressure of the lubricant in the axial gap T is almost the same as the pressure around the axial gap T. When the rotating shaft 10 moves to the left due to the thrust force received by the rotating shaft 10, the axial gap T decreases, and the pressure loss of the flow of the lubricant in the flow path of the axial gap T increases. On the other hand, the size of the radial gap R or the pressure loss of the lubricant flow in the flow path of the radial gap R hardly changes.

When the axial gap T is sufficiently small, the pressure loss of the lubricant flow in the flow passage of the axial gap T exceeds the pressure loss of the lubricant flow in the flow passage of the radial gap R. Therefore, the pressure of the lubricant existing in the axial gap T increases. As a result, a reaction force that pushes back the rotating shaft 10 that has moved to the right is generated. In other words, the thrust load capacity is generated on the rotary shaft 10. Furthermore, as the axial clearance T decreases, thrust support rigidity that increases the reaction force in the axial direction is developed.

When the turbo compressor 1b is operated under a high pressure ratio condition, it is necessary to rotate the rotary shaft 10 at high speed. The higher the pressure ratio under the operating conditions of the turbo compressor 1b, the greater the thrust force acting on the first impeller 21 and the second impeller 22. According to the turbo compressor 1b, the second flow path 14 communicates with the first flow path 12 and the radial gap R, for example. Therefore, the lubricant introduced into the first flow path 12 from the outside of the first flow path 12 can be supplied to the radial gap R. When the rotating shaft 10 rotates, the lubricant flowing through the second flow path 14 is accelerated and pressurized by the centrifugal force, and the lubricant can be supplied to the radial gap R and the axial gap T at a high pressure of 1 MPa, for example. As a result, the thrust support rigidity of the first bearing 41 increases. Therefore, even when the turbo compressor 1b is operated under the condition that the pressure ratio is high and the thrust force received by the rotating shaft 10 is large, the rotational speed of the rotating shaft is used to move the rotating shaft 10 in the axial direction. The amount can be suppressed. As a result, the design value of the clearance to be considered for preventing contact between the first impeller and the member arranged near the first impeller can be set more reliably. This is advantageous from the viewpoint of increasing the pressure ratio or increasing the adiabatic efficiency in the operation of the turbo compressor 1b.

The turbo compressor according to the present disclosure can be applied to applications such as an air conditioner, a chiller, and a turbo compressor for a heat cycle power generation system.

The turbo compressor according to the present disclosure can prevent resonance due to a decrease in natural frequency based on the bending rigidity of the rotating shaft, and can be applied to an air conditioner, a chiller, and a turbo compressor for a heat cycle power generation system.

1a, 1b Turbo compressor 10 Rotating shaft 11 One end 21 First impeller 21a First suction port 21b First discharge port 22 Second impeller 22a Second suction port 22b Second discharge port 23 Third impeller 30 Evaporator 41 First bearing 42 Second bearing 50 Condenser 70 Motor 71 Rotor 82a Return flow path 100 Refrigeration cycle device

Claims

A rotation axis,
A first impeller fixed to the rotating shaft, having a first suction port for sucking a working fluid and a first discharge port for discharging the working fluid;
A second impeller fixed to the rotating shaft, having a second suction port for sucking the working fluid and a second discharge port for discharging the working fluid,
A first bearing that is arranged between the first impeller and the second impeller in the axial direction of the rotating shaft and rotatably supports the rotating shaft,
A first flow path for guiding the working fluid from the first discharge port to the second suction port,
The first suction port and the second suction port are opened in the same direction,
Turbo compressor.
The first flow path is a return flow path that is arranged so as to approach the rotation shaft from the first discharge port and is connected to the second suction port.
The turbo compressor according to claim 1.
In the axial direction, the first region when the first flow path is projected onto the rotation axis and the second region when the first bearing is projected onto the rotation axis overlap.
The turbo compressor according to claim 1.
The width of the first region in the axial direction is larger than the width of the second region,
The turbo compressor according to any one of claims 1 to 3.
The turbo compressor according to any one of claims 1 to 4, which accelerates and compresses the working fluid having a saturated vapor pressure of atmospheric pressure or less at room temperature.
A second bearing that is arranged at a position farther from the first impeller than the first bearing in the axial direction, and rotatably supports the rotating shaft,
Further comprising a motor having a rotor fixed to the rotating shaft between the first bearing and the second bearing in the axial direction,
The turbo compressor according to any one of claims 1 to 5.
It has a third suction port that opens in the same direction as the first suction port and the second suction port, is arranged at a position apart from the first bearing in the axial direction, and is fixed to the rotary shaft. Further equipped with a third impeller,
The second impeller is arranged between the first impeller and the third impeller,
The turbo compressor according to any one of claims 1 to 6, wherein a diameter of the rotating shaft at the fixed position of the third impeller is equal to or larger than a diameter of the rotating shaft at the fixed position of the second impeller.
A second bearing that is arranged at a position farther from the first impeller than the first bearing in the axial direction, and rotatably supports the rotating shaft,
Further comprising a motor having a rotor fixed to the rotating shaft between the first bearing and the second bearing in the axial direction,
The turbo compressor according to claim 7, wherein the third impeller is fixed between the second impeller and the rotor in the axial direction.
The first bearing has a load capacity of thrust load of the rotating shaft,
The turbo compressor according to any one of claims 1 to 8.
The first bearing is a slide bearing having a bearing surface that forms a projection surface when projected onto a plane perpendicular to the axis of the rotating shaft,
The rotating shaft extends along the axis inside the rotating shaft, a first flow path through which a lubricant flows, and a second flow path that connects the space in contact with the bearing surface and the first flow path. Have,
The turbo compressor according to any one of claims 1 to 9.
The first bearing has a load capacity of a radial load of the rotating shaft,
The turbo compressor according to any one of claims 1 to 10.
An evaporator that produces a vapor phase refrigerant,
The turbo compressor according to any one of claims 1 to 11, wherein the gas-phase refrigerant generated in the evaporator is compressed as the working fluid.
A condenser for condensing the gas-phase refrigerant compressed by the turbo compressor,
Refrigeration cycle device.
The evaporator stores a liquid phase refrigerant therein,
The condenser stores the liquid phase refrigerant therein,
At least one of the liquid-phase refrigerant stored in the evaporator and the liquid-phase refrigerant stored in the condenser is supplied to the first bearing,
The refrigeration cycle apparatus according to claim 12.
A first impeller fixed to the rotating shaft and having a first suction port;
A second impeller fixed to the rotating shaft and opening in the same direction as the first impeller,
A first bearing arranged between the first impeller and the second impeller,
A first flow path that connects the first impeller and the second impeller, and is held by the first bearing,
A second flow path connected to the second impeller, comprising: a compression method of a turbo compressor,
The first impeller sucks a working fluid from the first suction port of the first impeller, compresses the working fluid, and discharges the working fluid to the first flow path,
The displacement of the first flow path due to the discharge of the working fluid is suppressed by being held by the first bearing,
The second impeller sucks the working fluid discharged from the first flow path by the second impeller, compresses the working fluid, and discharges the working fluid to the second flow path.
Working fluid compression method.