WO2022111411A1 - 离心压缩机 - Google Patents

离心压缩机 Download PDF

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Publication number
WO2022111411A1
WO2022111411A1 PCT/CN2021/132077 CN2021132077W WO2022111411A1 WO 2022111411 A1 WO2022111411 A1 WO 2022111411A1 CN 2021132077 W CN2021132077 W CN 2021132077W WO 2022111411 A1 WO2022111411 A1 WO 2022111411A1
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WO
WIPO (PCT)
Prior art keywords
volute
centrifugal
flow channel
outlet
centrifugal compressor
Prior art date
Application number
PCT/CN2021/132077
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English (en)
French (fr)
Inventor
俞国新
桂幸民
李思茹
陈锦践
朱万朋
韩聪
殷纪强
常云雪
毛守博
宋强
魏伟
Original Assignee
青岛海尔智能技术研发有限公司
海尔智家股份有限公司
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Application filed by 青岛海尔智能技术研发有限公司, 海尔智家股份有限公司 filed Critical 青岛海尔智能技术研发有限公司
Publication of WO2022111411A1 publication Critical patent/WO2022111411A1/zh

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D17/00—Radial-flow pumps, e.g. centrifugal pumps; Helico-centrifugal pumps
    • F04D17/08—Centrifugal pumps
    • F04D17/10—Centrifugal pumps for compressing or evacuating
    • F04D17/12—Multi-stage pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D25/00—Pumping installations or systems
    • F04D25/02—Units comprising pumps and their driving means
    • F04D25/06—Units comprising pumps and their driving means the pump being electrically driven
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D25/00—Pumping installations or systems
    • F04D25/02—Units comprising pumps and their driving means
    • F04D25/06—Units comprising pumps and their driving means the pump being electrically driven
    • F04D25/0606—Units comprising pumps and their driving means the pump being electrically driven the electric motor being specially adapted for integration in the pump
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/05—Shafts or bearings, or assemblies thereof, specially adapted for elastic fluid pumps
    • F04D29/056—Bearings
    • F04D29/058—Bearings magnetic; electromagnetic
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/26—Rotors specially for elastic fluids
    • F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/26—Rotors specially for elastic fluids
    • F04D29/28—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps
    • F04D29/284—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors
    • F04D29/286—Rotors specially for elastic fluids for centrifugal or helico-centrifugal pumps for radial-flow or helico-centrifugal pumps for compressors multi-stage rotors
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/40—Casings; Connections of working fluid
    • F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
    • F04D29/44—Fluid-guiding means, e.g. diffusers
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/40—Casings; Connections of working fluid
    • F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
    • F04D29/44—Fluid-guiding means, e.g. diffusers
    • F04D29/441—Fluid-guiding means, e.g. diffusers especially adapted for elastic fluid pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/60—Mounting; Assembling; Disassembling
    • F04D29/62—Mounting; Assembling; Disassembling of radial or helico-centrifugal pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/60—Mounting; Assembling; Disassembling
    • F04D29/62—Mounting; Assembling; Disassembling of radial or helico-centrifugal pumps
    • F04D29/624—Mounting; Assembling; Disassembling of radial or helico-centrifugal pumps especially adapted for elastic fluid pumps

Definitions

  • the invention relates to the technical field of compressors, in particular to a centrifugal compressor.
  • Centrifugal compressors have the significant advantages of energy saving, high efficiency, stable operation and long service life.
  • centrifugal compressors are suitable for large-flow, low-pressure work situations, and it is difficult to achieve high-efficiency, small-flow, high-pressure operation. Therefore, centrifugal compressors are used in large cooling capacity chillers.
  • Small and medium-sized refrigeration systems use more screw compressors, scroll compressors (such as small central air conditioners, including multi-line compressors) and rolling rotor compressors. These kinds of compressors operate far less efficiently than centrifugal compressors.
  • most of these types of compressors are lubricated with lubricating oil. It is very easy to cause problems such as the accumulation of lubricating oil in the heat exchanger, which leads to unfavorable oil return to the compressor, poor lubrication of compression-related components, and large heat exchange thermal resistance of the heat exchanger.
  • centrifugal compressors can be miniaturized and applied to small and medium-sized refrigeration systems to replace screw compressors, scroll compressors and even rolling rotor compressors, these refrigeration systems will be more energy efficient and will be beneficial to the refrigeration industry. have profound impact.
  • An object of the present invention is to solve or at least partially solve the above problems existing in the prior art, and to provide a centrifugal compressor that can maintain high efficiency on the basis of realizing miniaturization.
  • a further object of the present invention is to avoid the problem of compressor oil return in small and medium refrigeration systems.
  • a further object of the present invention is to reduce the diffuser losses of the compressed gas stream in a centrifugal compressor.
  • the present invention provides a centrifugal compressor comprising:
  • each compression unit comprising a volute mounted to the casing and a centrifugal impeller disposed in the volute, the centrifugal impeller configured to be rotated by a motor to compress the airflow entering the volute and pass it through the volute.
  • the outlet of the volute discharges.
  • the volute defines an intake runner, a volute-shaped runner and an outlet runner which are connected in sequence along the airflow direction;
  • the inlet runner extends along the axial direction of the centrifugal impeller;
  • the volute-shaped runner is a thickness direction parallel to the axis of the centrifugal impeller.
  • the direction of the flat shape; and the outlet flow channel gradually transitions from a flat shape to a cylindrical shape from the connection with the scroll flow channel to the outlet of the volute; the inlet of the centrifugal impeller faces the intake flow channel, and the outlet faces the scroll flow channel.
  • the thickness of the spiral flow channel is greater than the outlet width of the centrifugal impeller.
  • the ratio of the thickness of the volute channel to the outlet width of the centrifugal impeller is between 1.5 and 2.
  • the intake runner includes a tapered section whose cross section becomes gradually smaller along the airflow direction.
  • the tapered section is in the shape of a truncated cone as a whole, and its generatrix is an arc with the concave side facing the direction of the central axis of the intake runner.
  • the centrifugal impeller is a strongly backward curved closed impeller.
  • the centrifugal compressor is a two-stage compression type, the number of at least one compression unit is two, and the outlet of the volute of the compression unit of the low pressure stage is communicated with the inlet of the volute of the compression unit of the high pressure stage through a connecting pipe. .
  • the compression unit of the low pressure stage and the compression unit of the high pressure stage are located on two axial sides of the motor, respectively.
  • the centrifugal compressor further includes: at least one radial magnetic suspension bearing and/or at least one axial magnetic suspension bearing installed in the casing to support the rotor of the motor.
  • the centrifugal compressor of the present invention omits the diffuser, and the centrifugal impeller is directly installed in the volute, so as to avoid the relatively large expansion caused by the large rotation of the airflow in the diffuser.
  • the pressure loss can improve the efficiency of the compressor and make the centrifugal compressor more compact. Therefore, this structure is beneficial to realize the miniaturization of the centrifugal compressor, and keep it high efficiency, so that it is suitable for application in small chillers or multi-connected small central air conditioners.
  • the centrifugal compressor of the present invention can use radial magnetic suspension bearings and axial magnetic suspension bearings, and the magnetic suspension bearings are oil-free bearings, so there is no need to add lubricating oil into the centrifugal compressor, thereby completely avoiding the need for small and medium refrigeration systems.
  • Compressor oil return problem (the traditional screw compressors, scroll compressors and rolling rotor compressors are basically lubricated by oil), which improves the heat exchange efficiency of the heat exchanger; and the mechanical wear is small and the energy consumption is low , The noise is small, and the stability of the whole machine is enhanced, and the service life is longer.
  • the centrifugal compressor of the present invention makes the volute-shaped flow passage defined by the volute to be flat with the thickness direction parallel to the axial direction of the centrifugal impeller, and the flat volute-shaped flow passage flattens the volute as a whole, which is beneficial to realize the compressor. miniaturization. More importantly, the outlet flow channel is gradually transformed from a flat shape to a cylindrical shape from the connection with the volute flow channel to the outlet of the volute. In this way, the air flow can have a very good diffusion effect in the process of entering the cylindrical and wider outlet air passage from the thinner, flat swirl flow passage.
  • the outlet air passage gradually transitions from a flat shape to a cylindrical shape from the connection with the volute flow passage to the exit of the volute, the transition is very smooth, and the unnecessary resistance loss of the air flow is also reduced.
  • the cylindrical shape is also suitable for Connect to downstream piping.
  • the present invention makes the thickness of the volute-shaped flow channel larger than the outlet width of the centrifugal impeller, so that the airflow enters the volute-shaped flow channel of the volute case and then diffuses and decelerates, so as to reduce the Mach number and reduce the centrifugal effect. The result is that the uniformity of the flow field at the outlet of the volute is significantly increased, and the efficiency of the centrifugal compressor is finally improved.
  • the centrifugal impeller of the present invention is a strongly backward-curved closed impeller, so that more work done by the centrifugal impeller on the airflow is converted into a static pressure increase, and less is converted into a speed increase. Due to the large absolute airflow angle at the outlet of the strong backward curved centrifugal impeller, if the traditional diffuser is used, the airflow curl will be larger and the diffusion loss will be larger.
  • the present invention adopts the above-mentioned specially designed volute to directly connect the centrifugal impeller, which can effectively avoid this problem.
  • the present invention comprehensively installs the centrifugal impeller directly in the volute, specially designs the volute flow channel, and adopts the strong backward curved centrifugal impeller, which not only obtains various structural improvements.
  • the beneficial effects are also avoided, and the respective adverse effects are avoided, so that the overall efficiency of the centrifugal compressor is higher, and the structure is more compact, which is conducive to realizing its miniaturization.
  • FIG. 1 is a schematic diagram of a complete machine structure of a centrifugal compressor according to an embodiment of the present invention
  • Fig. 2 is a schematic cross-sectional view obtained by cutting the centrifugal compressor shown in Fig. 1 along the axial direction of the centrifugal impeller;
  • Fig. 3 is the enlarged view of A place of Fig. 2;
  • Fig. 4 is the structural representation of a compression unit in Fig. 1;
  • FIG. 5 is a schematic diagram of another viewing angle of the compression unit shown in FIG. 4;
  • Fig. 6 is the exploded schematic diagram of the compression unit shown in Fig. 4;
  • Fig. 7 is the structural representation of the centrifugal impeller in the compression unit shown in Fig. 6;
  • Figure 8 is a schematic diagram of the blade profile of the centrifugal impeller shown in Figure 7;
  • FIG 9 is an exploded schematic view of the centrifugal impeller shown in Figure 7;
  • FIG. 10 is a schematic structural diagram of the second impeller body in FIG. 9 .
  • the centrifugal compressor according to the embodiment of the present invention will be described below with reference to FIGS. 1 to 10 .
  • the x-axis is used to represent the axial direction of the centrifugal impeller 200, which is also the axial direction of the motor 40 and its stator 41 and rotor 42; solid arrows are used to represent the airflow direction.
  • FIG. 1 is a schematic diagram of the whole machine structure of a centrifugal compressor according to an embodiment of the present invention
  • FIG. 2 is a schematic cross-sectional view obtained by cutting the centrifugal compressor shown in FIG. 1 along the axial direction of the centrifugal impeller 200; Enlarged view of point A of 2.
  • the centrifugal compressor of the embodiment of the present invention may generally include a casing 10 , a motor 40 and at least one compression unit 20 , 30 .
  • the casing 10 defines an accommodating space, and the motor 40 is installed in the casing 10 .
  • the motor 40 includes a stator 41 and a rotor 42 , the stator 41 is fixed to the casing 10 , and the rotor 42 is rotatable relative to the stator 41 .
  • the number of compression units 20, 30 may be one or more.
  • the centrifugal compressor can be a single-stage compression type with only one compression unit.
  • the centrifugal compressor can also be of a multi-stage compression type, which provides a plurality of compression units 20 , 30 .
  • Each compression unit 20 , 30 includes a volute 100 mounted to the casing 10 and a centrifugal impeller 200 disposed within the volute 100 .
  • the centrifugal impeller 200 is configured to be rotated by the motor 40 to compress the airflow entering the volute 100 and discharge it through the outlet of the volute 100 .
  • a diffuser In traditional centrifugal compressors, a diffuser is basically installed downstream of the centrifugal impeller of each stage. The centrifugal impeller discharges the airflow into the diffuser, and the airflow is diffused by the diffuser and then enters the volute.
  • the centrifugal compressor of the present invention omits the diffuser, and the centrifugal impeller 200 is directly installed in the volute 100, so as to avoid the relatively large expansion caused by the large rotation of the airflow in the diffuser.
  • the pressure loss improves the overall efficiency of the centrifugal compressor, and also makes the structure of the centrifugal compressor more compact. Therefore, this structure is beneficial to realize the miniaturization of the centrifugal compressor, and keep it high efficiency, so that it is suitable for application in small chillers or multi-connected small central air conditioners.
  • the centrifugal compressor may be a two-stage compression type, and the number of compression units is two. It can be seen that one of the two compression units 20 and 30 must be a low-pressure stage and the other is a high-pressure stage. As shown in FIG. 1 and FIG. 2 , the compression unit 20 on the left side of the drawing is the low-pressure stage, and the compression unit on the right side is the low-pressure stage. 30 is the high pressure level. The outlet of the volute 100 of the compression unit 20 of the low pressure stage communicates with the inlet of the volute 100 of the compression unit 30 of the high pressure stage through the connecting pipe 90 .
  • connection pipe 90 is provided with a flange 91 to connect with the flange 130 of the outlet of the volute 100 of the compression unit 20 of the low pressure stage, and the outlet end of the connection pipe 90 is provided with a flange 92 to be connected with the flange 92 of the compression unit 30 of the high pressure stage.
  • the volute 100 is connected.
  • the compression unit 20 of the low pressure stage and the compression unit 20 of the high pressure stage are located on both sides of the motor 40 in the axial direction, so that the centrifugal impellers 200 of the two compression units 20 and 30 are directly connected to the motor 40 respectively, and it is beneficial to make the two centrifugal The axial force of the impeller 200 is partially offset.
  • the centrifugal compressor further includes at least one radial magnetic suspension bearing 60 and/or at least one axial magnetic suspension bearing 80 mounted within the casing 10 to support the rotor 42 of the motor 40 .
  • the centrifugal compressor includes two radial magnetic bearings 60 to support the rotor 42 in the radial direction.
  • the centrifugal compressor also includes an axial magnetic bearing 80 to counteract the axial force on the rotor 42 caused by the movement of the centrifugal impeller 200 .
  • the magnetic suspension bearing is made of the principle of magnetic suspension and is an oil-free bearing.
  • a common radial bearing 70 can also be installed at the axial end of the rotor 42 to support the end of the rotor 42 more and more. Stable and improve the operational reliability of centrifugal compressors.
  • FIG. 4 is a schematic structural diagram of a compression unit 20 shown in FIG. 1 ;
  • FIG. 5 is a schematic diagram of another viewing angle of the compression unit 20 shown in FIG. 4 ;
  • FIG. 6 is an exploded schematic diagram of the compression unit 20 shown in FIG. 4 .
  • the airflow enters the flow channel defined by the volute 100 from the inlet of the volute 100, then enters the centrifugal impeller 200 in the flow channel of the volute 100, and finally flows out from the outlet of the volute 100 and enters the next stage of compression Unit 30 or discharge compressor.
  • the volute 100 defines an intake flow channel 101 , a scroll-shaped flow channel 102 and an outlet flow channel 103 that are connected in sequence along the airflow direction, that is, the flow channel of the volute 100 is divided into three sections.
  • the inlet of the intake runner 101 constitutes the inlet of the volute 100 described herein, and the outlet of the outlet runner 103 constitutes the outlet of the volute 100 .
  • the intake flow passage 101 extends in the axial direction (x-axis direction) of the centrifugal impeller 200 .
  • the scroll channel 102 has a flat shape whose thickness direction is parallel to the axial direction of the centrifugal impeller 200 .
  • the outlet flow channel 103 gradually transitions from a flat shape to a cylindrical shape from the connection with the scroll-shaped flow channel 102 to the outlet of the volute 100.
  • the inlet 201 (refer to FIG. 8 ) of the centrifugal impeller 200 faces the intake runner 101
  • the outlet 202 faces the volute runner 102 , so as to take in air from the inlet runner 101 , compress the airflow and discharge it to the volute runner 102 .
  • the flat volute-shaped flow channel 102 flattens the whole of the volute 100 , which is beneficial to reduce the axial dimension of the centrifugal compressor and realize the miniaturization of the compressor. More importantly, since the outlet flow channel 103 gradually transitions from a flat shape to a cylindrical shape from the connection with the scroll-shaped flow channel 102 to the exit of the volute 100, the air flow changes from a thinner, flat scroll-shaped flow channel to a cylindrical shape. When the 102 enters the cylindrical and relatively spacious outlet air passage 103, a very good diffusion effect can be obtained.
  • outlet flow channel 103 gradually transitions from a flat shape to a cylindrical shape from the connection with the scroll-shaped flow channel 102 to the outlet of the volute 100, the transition is very smooth, and the unnecessary resistance loss of the air flow is also reduced.
  • the shape is also suitable for connection with downstream pipelines.
  • the intake runner 101 can include a tapered section 1011 with a gradually decreasing cross section along the airflow direction, so as to improve the suction efficiency.
  • the tapered section 1011 can be a truncated cone as a whole, and its generatrix is an arc with the concave side facing the direction of the central axis of the intake runner 101 .
  • the generatrix of the tapered section 1011 can also be a straight line or a combination of various shapes.
  • the intake flow channel 101 can be made to be tapered as a whole, or a part of the section can be a tapered section, and a part of the section can be a straight section whose cross section does not change with the change of the axial position.
  • the volute 100 may be a split structure, which includes a volute body 110 and a cover plate 120 that are assembled along the axis direction of the intake flow channel 101 .
  • the volute body 110 defines the aforementioned intake runner 101 , the first half of the scroll runner 102 and the outlet runner 103 , wherein the scroll runner 102 is open to one side of the cover plate 120 .
  • the cover plate 120 is covered on the axial side of the volute body 110 to cover the open side of the first half of the scroll flow channel 102, and defines the second half of the scroll flow channel 102, the scroll flow
  • the first half of the channel 102 and the second half of the scroll channel 102 relatively constitute a complete scroll channel 102 .
  • the rotating shaft 214 of the centrifugal impeller 200 is connected to the rotor 42 of the motor 40 through the central hole of the cover plate 120 .
  • the volute 100 as a split structure, the volute body 110 and the cover plate 120 are processed separately, so as to form the intake flow channel 101 , the scroll-shaped flow channel 102 and the outlet flow channel 103 by machining.
  • the surfaces of the intake runner 101 , the volute-shaped runner 102 and the outlet runner 103 are smoother, which can better satisfy the uniformity of the internal flow field. Reduce the flow loss caused by the rough surface of the flow channel, and improve the operating efficiency of the centrifugal compressor.
  • a rounded transition (R angle in FIG. 3 ) between the two planes in the thickness direction of the scroll channel 102 and the circumferential scroll side surface can be used to increase the volume of the scroll case. strength, relieve the local stress concentration, eliminate the angular vortex, and ensure the uniformity of the flow field.
  • the size of R can be selected according to the thickness of the snail-shaped flow channel 102 .
  • the split structure of the volute 100 facilitates the machining of the above-mentioned rounded corners.
  • FIG. 7 is a schematic structural diagram of the centrifugal impeller 200 in the compression unit 20 shown in FIG. 6 ;
  • FIG. 8 is a schematic diagram of a blade profile of the centrifugal impeller 200 shown in FIG. 7 .
  • the thickness of the scroll channel 102 is made greater than the outlet width of the centrifugal impeller 200 .
  • the thickness of the volute-shaped flow channel 102 refers to its size in the direction of the axis (x-axis) of the centrifugal impeller 200
  • the outlet width B of the centrifugal impeller 200 refers to the outlet 202 of the centrifugal impeller 200 in the direction of the axis of the centrifugal impeller 200 .
  • the dimensions are specifically marked in Figure 7. Specifically, it has been confirmed by the inventor that the ratio of the thickness of the snail-shaped flow channel 102 to the outlet width of the centrifugal impeller 200 is set between 1.5 and 2, and the optimal effect can be obtained.
  • the thickness of the scroll-shaped flow channel 102 is particularly larger than the outlet width B of the centrifugal impeller 200, so that the airflow enters the scroll casing 100 (the scroll-shaped flow channel 102 of the scroll-shaped flow channel 102) and then diffuses and reduces the speed. , the Mach number is reduced, the centrifugal effect is reduced, and finally the uniformity of the flow field at the outlet of the volute 100 is significantly increased, and the compressor efficiency is finally improved.
  • the centrifugal impeller 200 is a strongly backward curved closed impeller. As shown in FIG. 8 , the centrifugal impeller 200 has a plurality of blades 203 arranged along its circumferential direction, a flow channel 212 is formed between every two adjacent blades 203 , and the airflow enters each flow channel through the inlet 201 of the centrifugal impeller 200 The radially inner side of the 212 is rotated by the centrifugal impeller 200 , so that the airflow flows in each flow channel to its radially outer side, so as to flow out of the centrifugal impeller 200 and flow to the volute-shaped flow channel 102 of the volute 100 .
  • FIG. 8 shows the direction of rotation of the centrifugal impeller 200 with arrows.
  • Each blade 203 of the centrifugal impeller 200 is a back-bending structure, and the tip of the blade (the end adjacent to the radially outer edge of the centrifugal impeller 200 ) is bent backward compared to the rest of the sections, so that each blade of the centrifugal impeller 200 forms a strong back-bending structure, as shown in Figure 8.
  • the centrifugal impeller 200 is strongly backward curved, so that more work done by the centrifugal impeller 200 on the airflow is converted into a static pressure increase, and less is converted into a speed increase. Due to the large absolute airflow angle at the outlet of the strong backward curved centrifugal impeller, if the traditional diffuser form is used, the airflow curl will be larger and the diffusion loss will be larger. In the embodiment of the present invention, the above-mentioned specially designed volute 100 is directly connected to the centrifugal impeller 200, which can effectively avoid this problem. It can be seen that the improvement points of the embodiments of the present invention are not isolated from each other, but function in combination.
  • the centrifugal impeller 200 is directly installed in the volute 100, the flow channel of the volute 100 is specially designed, and the strong backward curved centrifugal impeller 200 is used.
  • the beneficial effects of various structural improvements are greatly avoided, and the respective adverse effects are greatly avoided, so that the overall efficiency of the centrifugal compressor is higher, and the structure is more compact, which is conducive to realizing miniaturization.
  • FIG. 9 is an exploded schematic diagram of the centrifugal impeller 200 shown in FIG. 7 ;
  • FIG. 10 is a schematic structural diagram of the second impeller body 220 in FIG. 9 .
  • the centrifugal impeller 200 may be of a split type. Specifically, the centrifugal impeller 200 includes a first impeller body 210 and a second impeller body 220, the first impeller body 210 and the second impeller body 220 are both disc-shaped and connected to each other, and the opposite surfaces of the two are respectively formed with blade halves, The respective blade halves of the two are joined to form a complete blade 203 .
  • the first impeller body 210 and the second impeller body 220 may be connected and fastened by a plurality of fasteners such as rivets 230 .
  • One or more positioning grooves 2113 can also be formed on the first impeller body 210, and the second impeller body 220 is formed with the same number of positioning protrusions 2214, so that each positioning protrusion 2214 is snapped into a positioning groove 2113, in order to make the position between the first impeller body 210 and the second impeller body 220 more stable, make the alignment between the blade halves more accurate, and prevent the misalignment of the blades 203 from affecting the performance of the centrifugal impeller 200 .
  • the first impeller body 210 is provided with a rotating shaft 214 , and a mounting hole 215 is formed in the center thereof so as to be connected to the rotor 42 by a screw 300 .
  • the second impeller body 220 is provided with the inlet 201 of the centrifugal impeller 200 .
  • Traditional centrifugal impellers are all cast in one piece, and their surface accuracy is not ideal, which will affect their compression efficiency and adverse noise. Especially for a closed impeller, the blade is inside, and the accuracy of the blade surface is more difficult to guarantee.
  • centrifugal impeller 200 as the above-mentioned split type, two impeller bodies can be fabricated separately, so that the blades of each impeller body are exposed so that the surface of the impeller body can be processed to make it smoother, and a more ideal surface accuracy.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Electromagnetism (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)

Abstract

一种离心压缩机，其包括：机壳（10）；电机（40），安装于机壳（10）内；和至少一个压缩单元（23,30），每个压缩单元（23,30）包括安装于机壳（10）的蜗壳（100）和设置在蜗壳（100）内的离心叶轮（200），离心叶轮（200）配置成在电机（40）驱动下转动，以对进入蜗壳（100）的气流进行压缩并将其经蜗壳（100）的出口排出。该离心压缩机能够小型化，且能保持高效率。

Description

离心压缩机 技术领域
本发明涉及压缩机技术领域，特别涉及一种离心压缩机。
背景技术
离心压缩机具有节能高效、运行稳定和寿命长的显著优点。但是在制冷领域，离心压缩机适合于大流量、低压比工作场合，难以实现高效率地小流量、高压比运行。因此，离心压缩机均应用于大冷量冷水机组。而中小型的制冷系统则更多使用螺杆压缩机、涡旋压缩机(如小型中央空调，包括多联机)和滚动转子压缩机。这些种类的压缩机的运行效率远不如离心压缩机。而且，这些种类的压缩机大多采用润滑油润滑。非常容易产生因润滑油积存于换热器中，导致向压缩机回油不利、使压缩相关部件润滑变差、换热器换热热阻变大等问题。
因此，如果能实现离心压缩机的小型化，使其应用于中小型制冷系统，以代替螺杆压缩机、涡旋压缩机甚至滚动转子压缩机，将使这些制冷系统的能效更高，对制冷行业产生深远影响。
发明内容
本发明的一个目的是要解决或至少部分地解决现有技术存在的上述问题，提供一种离心压缩机，在实现小型化的基础上，能保持高效率。
本发明的进一步的目的是要避免中小型制冷系统的压缩机回油问题。
本发明的进一步的目的是要减少离心压缩机中压缩后气流的扩压损失。
特别地，本发明提供了一种离心压缩机，其包括：
机壳；
电机，安装于机壳内；和
至少一个压缩单元，每个压缩单元包括安装于机壳的蜗壳和设置在蜗壳内的离心叶轮，离心叶轮配置成在电机驱动下转动，以对进入蜗壳的气流进行压缩并将其经蜗壳的出口排出。
可选地，蜗壳限定出沿气流方向依次相连的进气流道、蜗形流道和出气流道；进气流道沿离心叶轮的轴线方向延伸；蜗形流道为厚度方向平行于离心叶轮轴线方向的扁平状；且出气流道从与蜗形流道相接处至蜗壳的出口处 逐渐从扁平状过渡为圆柱状；离心叶轮的进口朝向进气流道，出口朝向蜗形流道。
可选地，蜗形流道的厚度大于离心叶轮的出口宽度。
可选地，蜗形流道的厚度与离心叶轮的出口宽度之比在1.5至2之间。
可选地，进气流道包括沿气流方向截面逐渐变小的渐缩段。
可选地，渐缩段整体为截锥形，其母线为凹侧朝向进气流道中心轴线方向的弧形。
可选地，离心叶轮为强后弯式闭式叶轮。
可选地，离心压缩机为双级压缩式，至少一个压缩单元的数量为两个，且其中低压级的压缩单元的蜗壳的出口通过连接管与高压级的压缩单元的蜗壳的进口连通。
可选地，低压级的压缩单元与高压级的压缩单元分别位于电机的轴向两侧。
可选地，离心压缩机还包括：至少一个径向磁悬浮轴承和/或至少一个轴向磁悬浮轴承，其安装于机壳内，以支撑电机的转子。
本发明的离心压缩机相比于传统的离心压缩机而言，省略了扩压器，将离心叶轮直接安装于蜗壳内，以避免气流在扩压器内旋度较大引发比较大的扩压损失，使压缩机整机效率得以提升，同时也使离心压缩机的结构更加紧凑。因此，这种结构有利于实现离心压缩机的小型化，且使其保持较高效率，以适于应用于小型的冷水机组或多联机等小型中央空调。
进一步地，本发明的离心压缩机可采用径向磁悬浮轴承和轴向磁悬浮轴承，磁悬浮轴承为无油化轴承，因此无需再在离心压缩机内加入润滑油，从而彻底避免了中小型制冷系统的压缩机回油问题(传统惯常采用的螺杆式压缩机、涡旋压缩机和滚动转子压缩机基本均为有油润滑)，提升了换热器的换热效率；而且机械磨损小、能耗低、噪声小，也使整机的稳定性增强，寿命更长。
进一步地，本发明的离心压缩机使蜗壳限定出的蜗形流道为厚度方向平行于离心叶轮轴线方向的扁平状，扁平状的蜗形流道使得蜗壳整体扁平化，利于实现压缩机小型化。更重要的是，使出气流道从与蜗形流道相接处至蜗壳的出口处逐渐从扁平状过渡为圆柱状。如此一来，气流从较薄的、扁平状的蜗形流道进入圆柱状、较宽敞的出气流道的过程中，能够有非常好的扩压 效果。而且由于出气流道从与蜗形流道相接处至蜗壳的出口处逐渐从扁平状过渡为圆柱状，过渡非常平顺，也减少了气流的不必要的阻力损失，同时圆柱状也适于与下游管道进行连接。
进一步地，发明人认识到，由离心叶轮直接向蜗壳排气将导致气流马赫数增加、气流离心效应大，使气流向径向外侧聚积，导致流场不均匀，引起较大的流动损失。为消除或至少缓解上述不利影响，本发明特别使蜗形流道的厚度大于离心叶轮的出口宽度，使得气流进入蜗壳的蜗形流道后扩压降速，使其马赫数下降，离心效应降低，最终使蜗壳出口的流场均匀性显著增加，最终提升了离心压缩机的效率。
进一步地，本发明的离心叶轮为强后弯式闭式叶轮，以使离心叶轮对气流做功更多转化为静压提升，更少转化为速度增加。由于强后弯式离心叶轮的出口绝对气流角度较大，若采用传统的扩压器将导致气流旋度更大，扩压损失更大。本发明采用上述特别设计的蜗壳直接连接离心叶轮，可有效避免这一问题。由此可见，本发明综合性地把离心叶轮直接安装于蜗壳内，对蜗壳流道进行特别设计，以及采用强后弯式离心叶轮这些改进结合在一起，不仅获得了各项结构改进的有益效果，而且还避免了各自的不利影响，使得离心压缩机的整体效率更高，而且结构更加紧凑，利于实现其小型化。
根据下文结合附图对本发明具体实施例的详细描述，本领域技术人员将会更加明了本发明的上述以及其他目的、优点和特征。
附图说明
后文将参照附图以示例性而非限制性的方式详细描述本发明的一些具体实施例。附图中相同的附图标记标示了相同或类似的部件或部分。本领域技术人员应该理解，这些附图未必是按比例绘制的。附图中：
图1是根据本发明一个实施例的离心压缩机的整机结构示意图；
图2是对图1所示离心压缩机沿离心叶轮的轴线方向剖切后得到的示意性剖视图；
图3是图2的A处放大图；
图4是图1中的一个压缩单元的结构示意图；
图5是图4所示压缩单元的另一视角的示意图；
图6是图4所示压缩单元的分解示意图；
图7是图6所示压缩单元中的离心叶轮的结构示意图；
图8是图7所示离心叶轮的叶片型线示意图；
图9是图7所示离心叶轮的分解示意图；
图10是图9中的第二叶轮体的结构示意图。
具体实施方式
下面参照图1至图10来描述本发明实施例的离心压缩机。部分图中用x轴表示离心叶轮200的轴线方向，同时也是电机40及其定子41和转子42的轴线方向；用实心箭头表示气流方向。
图1是根据本发明一个实施例的离心压缩机的整机结构示意图；图2是对图1所示离心压缩机沿离心叶轮200的轴线方向剖切后得到的示意性剖视图；图3是图2的A处放大图。
如图1至图3所示，本发明实施例的离心压缩机一般性地可包括机壳10、电机40和至少一个压缩单元20,30。
机壳10限定有容纳空间，电机40安装于机壳10内。电机40包括定子41和转子42，定子41固定于机壳10，转子42可相对定子41转动。压缩单元20,30的数量可为一个或多个。例如，可使离心压缩机为单级压缩式，仅设置一个压缩单元。也可使离心压缩机为多级压缩式，其设置多个压缩单元20,30。每个压缩单元20,30包括安装于机壳10的蜗壳100和设置在蜗壳100内的离心叶轮200。离心叶轮200配置成在电机40驱动下转动，以对进入蜗壳100的气流进行压缩并将其经蜗壳100的出口排出。
传统的离心压缩机基本在每一级的离心叶轮的下游设置一扩压器，离心叶轮将气流排入扩压器，气流被扩压器扩压后再进入蜗壳。
本发明的离心压缩机相比于传统的离心压缩机，省略了扩压器，将离心叶轮200直接安装于蜗壳100内，以避免气流在扩压器内旋度较大引发比较大的扩压损失，使离心压缩机的整机效率得以提升，并且也使离心压缩机的结构更加紧凑。因此，这种结构有利于实现离心压缩机的小型化，且使其保持较高效率，以适于应用于小型的冷水机组或多联机等小型中央空调。
在一些实施例中，如图1和图2所示，离心压缩机可为双级压缩式，压缩单元的数量为两个。可知，两个压缩单元20,30中必然有一个为低压级，另一个为高压级，如图1和图2所示，位于图面左侧的压缩单元20为低压 级，右侧的压缩单元30为高压级。低压级的压缩单元20的蜗壳100的出口通过连接管90与高压级的压缩单元30的蜗壳100的进口连通。具体地，连接管90的进口端设置法兰91以与低压级的压缩单元20的蜗壳100出口的法兰130相接，连接管90出口端设置法兰92以与高压级压缩单元30的蜗壳100连接。优选使低压级的压缩单元20与高压级的压缩单元20分别位于电机40的轴向两侧，以便两个压缩单元20,30的离心叶轮200分别直接连接于电机40，且利于使两个离心叶轮200的轴向力进行部分抵消。
在一些实施例中，离心压缩机还包括至少一个径向磁悬浮轴承60和/或至少一个轴向磁悬浮轴承80，其安装于机壳10内，以支撑电机40的转子42。如图2所示，离心压缩机包括两个径向磁悬浮轴承60，以在径向方向支撑转子42。离心压缩机还包括一个轴向磁悬浮轴承80，以抵消离心叶轮200运动给转子42产生的轴向力。磁悬浮轴承采用磁悬浮原理制成，为无油化轴承。因此无需再在离心压缩机内加入润滑油，从而彻底避免了中小型制冷系统的压缩机回油问题(传统惯常采用的螺杆式压缩机、涡旋压缩机和滚动转子压缩机基本均为有油润滑)，提升了换热器的换热效率。而且采用磁悬浮轴承使得离心压缩机的机械磨损小、能耗低、噪声小、稳定性增强、寿命更长。
进一步地，如图3所示，在设置上述磁悬浮轴承的基础上，还可在转子42的轴向端部另设普通的径向轴承70，以对转子42端部进行重点支撑，使其更加稳定，提升离心压缩机的运行可靠性。
图4是图1中的一个压缩单元20的结构示意图；图5是图4所示压缩单元20的另一视角的示意图；图6是图4所示压缩单元20的分解示意图。
在本发明实施例中，气流从蜗壳100的进口进入蜗壳100限定的流道，然后进入处于蜗壳100流道内的离心叶轮200，最后从蜗壳100的出口流出，进入下一级压缩单元30或排出压缩机。
在一些实施例中，如图4至图6所示，蜗壳100限定出沿气流方向依次相连的进气流道101、蜗形流道102和出气流道103，即蜗壳100流道分为三个区段。进气流道101的进口即构成本文所述的蜗壳100的进口，出气流道103的出口构成蜗壳100的出口。进气流道101沿离心叶轮200的轴线方向(x轴方向)延伸。蜗形流道102为厚度方向平行于离心叶轮200的轴线方向的扁平状。出气流道103从与蜗形流道102相接处至蜗壳100的出口处 逐渐从扁平状过渡为圆柱状。离心叶轮200的进口201(参考图8)朝向进气流道101，出口202(参考图8)朝向蜗形流道102，以从进气流道101进气，将气流压缩后将其排向蜗形流道102。
本实施例中，扁平状的蜗形流道102使得蜗壳100的整体扁平化，利于减小离心压缩机的轴向尺寸，实现压缩机小型化。更重要的是，由于出气流道103从与蜗形流道102相接处至蜗壳100的出口处逐渐从扁平状过渡为圆柱状，使气流从较薄的、扁平状的蜗形流道102进入圆柱状、较宽敞的出气流道103的过程中，能够有非常好的扩压效果。而且，由于出气流道103从与蜗形流道102相接处至蜗壳100的出口处逐渐从扁平状过渡为圆柱状，过渡非常平顺，也减少了气流的不必要的阻力损失，同时圆柱状也适于与下游管道进行连接。
在一些实施例中，如图2所示，可使进气流道101包括沿气流方向截面逐渐变小的渐缩段1011，以提升吸气效率。具体地，可使渐缩段1011整体为截锥形，其母线为凹侧朝向进气流道101中心轴线方向的弧形。此外，也可使渐缩段1011的母线为直线或多种形状的组合。可使进气流道101整体为渐缩状，也可使部分区段为渐缩段，部分区段为截面不随轴向位置改变而改变的平直段。
在一些实施例中，如图6所示，可使蜗壳100为分体式结构，其包括沿进气流道101的轴线方向拼合而成的蜗壳本体110和盖板120。蜗壳本体110限定有前述的进气流道101、蜗形流道102的第一半部和出气流道103，其中蜗形流道102朝向盖板120的一侧敞开。盖板120盖设在蜗壳本体110的轴向一侧，以封盖蜗形流道102的第一半部的敞开侧，且限定有蜗形流道102的第二半部，蜗形流道102的第一半部和蜗形流道102的第二半部相对构成完整的蜗形流道102。离心叶轮200的转轴214穿过盖板120的中心孔连接至电机40的转子42。本实施例通过将蜗壳100设置为分体结构，使蜗壳本体110和盖板120分别加工，以便通过机加工方式形成进气流道101、蜗形流道102和出气流道103。相比于现有的一体铸造式的蜗壳，本实施例中，进气流道101、蜗形流道102和出气流道103的表面更加光滑，能更好地满足内部流场的均匀性，减少因流道表面过于粗糙带来的流动损失，提升离心压缩机的运行效率。
进一步地，如图3所示，可使蜗形流道102的厚度方向上的两个平面与 周向蜗形侧面之间以圆角过渡(图3中的R角)，以便增大蜗壳强度，缓解该局部应力集中，消除角涡，保证流场的均匀性。R的大小可根据蜗形流道102的厚度进行选定。蜗壳100的分体式结构方便了上述圆角的加工。
图7是图6所示压缩单元20中的离心叶轮200的结构示意图；图8是图7所示离心叶轮200的叶片型线示意图。
在一些实施例中，使蜗形流道102的厚度大于离心叶轮200的出口宽度。其中，蜗形流道102的厚度指的是其在离心叶轮200轴线(x轴)方向上的尺寸，离心叶轮200的出口宽度B指的是离心叶轮200的出口202在离心叶轮200轴线方向上的尺寸，具体标注在图7中。具体地，经发明人多次试验确认，将蜗形流道102的厚度与离心叶轮200的出口宽度之比设置在1.5至2之间，可取得最优效果。
发明人认识到使离心叶轮200直接向蜗壳100排气将导致气流马赫数增加，气流离心效应大而向径向外侧聚积，导致流场不均匀，引起较大的流动损失。为消除或至少缓解上述不利影响，本发明实施例特别使蜗形流道102的厚度大于离心叶轮200的出口宽度B，使得气流进入蜗壳100(的蜗形流道102)后扩压降速，使其马赫数下降，离心效应降低，最终使蜗壳100出口的流场均匀性显著增加，最终提升了压缩机效率。
在一些实施例中，离心叶轮200为强后弯式闭式叶轮。如图8所示，离心叶轮200具有多个沿其周向排列的多个叶片203，每相邻两个叶片203之间形成一个流道212，气流经离心叶轮200的进口201进入各流道的212的径向内侧，经离心叶轮200转动，使气流在各流道内向其径向外侧流动，以流出离心叶轮200，流向蜗壳100的蜗形流道102。在此期间各叶片203对气流做功使气流压力升高。图8用箭头示意了离心叶轮200的转动方向。离心叶轮200的各叶片203为后弯式结构，且叶片末端(邻近离心叶轮200径向外边缘的一端)相比其余区段向后弯折，以使离心叶轮200的各叶片形成强后弯式结构，如图8所示。
本发明实施例使离心叶轮200为强后弯式，以使离心叶轮200对气流做功更多转化为静压提升，更少转化为速度增加。由于强后弯式离心叶轮的出口绝对气流角度较大，若采用传统的扩压器形式将导致气流旋度更大，扩压损失更大。本发明实施例采用上述特别设计的蜗壳100直接连接离心叶轮200，可有效避免这一问题。由此可见，本发明实施例各改进点并非相互孤 立，而是相结合地发挥作用。具体地，本发明实施例综合性地把离心叶轮200直接安装于蜗壳100内，对蜗壳100流道进行特别设计，以及采用强后弯式离心叶轮200这些改进结合在一起，不仅获得了各项结构改进的有益效果，而且还极大避免了各自的不利影响，使得离心压缩机整体的效率较高，而且结构更加紧凑，利于实现小型化。
图9是图7所示离心叶轮200的分解示意图；图10是图9中的第二叶轮体220的结构示意图。
在一些实施例中，如图8至图10所示，离心叶轮200可为分体式。具体地，离心叶轮200包括第一叶轮体210和第二叶轮体220，第一叶轮体210和第二叶轮体220均为盘状且对接相连，两者相对的表面各自形成有叶片半部，两者各自的叶片半部相接构成完整的叶片203。
可使第一叶轮体210和第二叶轮体220通过多个紧固件例如铆钉230连接紧固。还可使第一叶轮体210上形成有一个或多个定位凹槽2113，使第二叶轮体220上形成有相同数量的定位凸起2214，使每个定位凸起2214卡入一个定位凹槽2113中，以使第一叶轮体210和第二叶轮体220之间的位置更加稳固，使各叶片半部之间的对位更加精准，避免各叶片203对位不齐影响离心叶轮200的性能。
第一叶轮体210上设置有转轴214，其中央形成有安装孔215，以便通过一螺钉300连接在转子42上。第二叶轮体上220设置有离心叶轮200的进口201。传统的离心叶轮均为一体铸造式，其表面精度不理想，会影响其压缩效率和不利的噪声。特别是对于封闭式叶轮而言，叶片处于内部，叶片表面的精度更加难以保证。本实施例通过将离心叶轮200设置为上述的分体式，以便对两个叶轮体分别进行制作，使每个叶轮体的叶片显露在外以便对其表面进行处理，使其更加光滑，可获得更加理想的表面精度。
至此，本领域技术人员应认识到，虽然本文已详尽示出和描述了本发明的多个示例性实施例，但是，在不脱离本发明精神和范围的情况下，仍可根据本发明公开的内容直接确定或推导出符合本发明原理的许多其他变型或修改。因此，本发明的范围应被理解和认定为覆盖了所有这些其他变型或修改。

Claims (10)

  1. 一种离心压缩机，包括：
    机壳；
    电机，安装于所述机壳内；和
    至少一个压缩单元，每个所述压缩单元包括安装于所述机壳的蜗壳和设置在所述蜗壳内的离心叶轮，所述离心叶轮配置成在所述电机驱动下转动，以对进入所述蜗壳的气流进行压缩并将其经所述蜗壳的出口排出。
  2. 根据权利要求1所述的离心压缩机，其中，
    所述蜗壳限定出沿气流方向依次相连的进气流道、蜗形流道和出气流道；
    所述进气流道沿所述离心叶轮的轴线方向延伸；
    所述蜗形流道为厚度方向平行于所述离心叶轮轴线方向的扁平状；且
    所述出气流道从与所述蜗形流道相接处至所述蜗壳的出口处逐渐从扁平状过渡为圆柱状；
    所述离心叶轮的进口朝向所述进气流道，出口朝向所述蜗形流道。
  3. 根据权利要求2所述的离心压缩机，其中，
    所述蜗形流道的厚度大于所述离心叶轮的出口宽度。
  4. 根据权利要求3所述的离心压缩机，其特征在于，
    所述蜗形流道的厚度与所述离心叶轮的出口宽度之比在1.5至2之间。
  5. 根据权利要求2所述离心压缩机，其中，
    所述进气流道包括沿气流方向截面逐渐变小的渐缩段。
  6. 根据权利要求5所述的离心压缩机，其中，
    所述渐缩段整体为截锥形，其母线为凹侧朝向进气流道中心轴线方向的弧形。
  7. 根据权利要求1所述的离心压缩机，其中，
    所述离心叶轮为强后弯式闭式叶轮。
  8. 根据权利要求1所述的离心压缩机，其中，
    所述离心压缩机为双级压缩式，所述至少一个压缩单元的数量为两个，且其中低压级的压缩单元的所述蜗壳的出口通过连接管与高压级的压缩单元的所述蜗壳的进口连通。
  9. 根据权利要求8所述的离心压缩机，其中，
    低压级的压缩单元与高压级的压缩单元分别位于所述电机的轴向两侧。
  10. 根据权利要求1所述离心压缩机，还包括：
    至少一个径向磁悬浮轴承和/或至少一个轴向磁悬浮轴承，其安装于所述机壳内，以支撑所述电机的转子。
PCT/CN2021/132077 2020-11-24 2021-11-22 离心压缩机 WO2022111411A1 (zh)

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CN117588424A (zh) * 2024-01-19 2024-02-23 沈阳山图透平技术有限公司 高速磁悬离心式一体化工艺压缩机

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