CN114542507A

CN114542507A - Centrifugal compressor and impeller thereof

Info

Publication number: CN114542507A
Application number: CN202011329547.8A
Authority: CN
Inventors: 俞国新; 桂幸民; 殷纪强; 陈锦践; 李思茹; 朱万朋; 韩聪; 常云雪; 魏伟
Original assignee: Qingdao Haier Smart Technology R&D Co Ltd; Haier Smart Home Co Ltd
Current assignee: Qingdao Haier Smart Technology R&D Co Ltd; Haier Smart Home Co Ltd
Priority date: 2020-11-24
Filing date: 2020-11-24
Publication date: 2022-05-27

Abstract

The invention provides a centrifugal compressor and an impeller thereof, wherein the impeller comprises a plurality of blades which are arranged along the circumferential direction of the impeller, and a flow channel is formed between every two adjacent blades; each blade is of a backward bending type which is gradually bent from the inlet end to the outlet end of the blade in the direction departing from the rotation direction of the impeller; and the thickness of each blade is gradually increased and then gradually decreased in the direction from the inlet end to the outlet end. The impeller of the invention has higher pneumatic efficiency and is beneficial to realizing the miniaturization of the centrifugal compressor.

Description

Centrifugal compressor and impeller thereof

Technical Field

The invention relates to the technical field of compressors, in particular to a centrifugal compressor and an impeller thereof.

Background

The centrifugal compressor has the remarkable advantages of energy conservation, high efficiency, stable operation and long service life. However, in the refrigeration field, the centrifugal compressor is suitable for large-flow and low-pressure-ratio working occasions, and the efficient low-flow and high-pressure-ratio operation is difficult to realize. Therefore, the centrifugal compressor is applied to a large-cooling-capacity water chilling unit. Medium and small-sized refrigeration systems are increasingly using screw compressors, scroll compressors (such as small central air conditioners including multi-split air conditioners) and rolling rotor compressors. These types of compressors, however, operate far less efficiently than centrifugal compressors. Most of these compressors are lubricated with lubricating oil. The problems that the lubricating oil is accumulated in the heat exchanger, the oil is disadvantageously returned to the compressor, the lubrication of related compression parts is poor, the heat exchange resistance of the heat exchanger is increased and the like are easily caused.

Therefore, how to solve various problems caused by the miniaturization of the centrifugal compressor can make the centrifugal compressor applied to small and medium-sized refrigeration systems to replace screw compressors, scroll compressors and even rolling rotor compressors, so that the energy efficiency of the refrigeration systems is higher, and the refrigeration industry is influenced profoundly.

Disclosure of Invention

An object of the present invention is to provide a centrifugal compressor and an impeller thereof, which solve or at least partially solve the above problems of the prior art.

The invention aims to provide a centrifugal compressor and an impeller thereof, which improve the pneumatic efficiency of the impeller.

In one aspect, the present invention provides an impeller for a centrifugal compressor, comprising a plurality of blades arranged along a circumferential direction thereof, wherein a flow passage is formed between every two adjacent blades;

each blade is of a backward bending type which is gradually bent from the inlet end to the outlet end of the blade in the direction departing from the rotation direction of the impeller; and is

The thickness of each vane gradually increases and then gradually decreases in a direction from the inlet end toward the outlet end.

Optionally, a side surface of each blade facing the rotation direction of the impeller is a pressure surface, and a side surface of each blade facing away from the rotation direction is a suction surface;

in the direction from the inlet end to the outlet end, the pressure surface of each blade comprises a first concave section and a first convex section which are smoothly connected in sequence, and each suction surface comprises a second convex section and a second concave section which are smoothly connected in sequence.

Optionally, in a direction from the inlet end to the outlet end, the installation angle of the pressure surface and the suction surface of each blade gradually increases and then gradually decreases;

the installation angle is an included angle between a tangent line of any point of the pressure surface or the suction surface and the tangential direction of the impeller of the point.

Optionally, the mounting angle of the pressure face at the outlet end is between 0 ° and 10 °;

the suction surface has an angle of incidence at the outlet end of between 20 ° and 40 °.

Optionally, the mounting angle of the pressure surface at the inlet end is between 5 ° and 15 °;

the mounting angle of the suction surface at the inlet end is smaller than the mounting angle of the pressure surface at the inlet end and is between 0 ° and 10 °.

Optionally, the ratio of the length of the first convex segment to the first concave segment is between 3 and 5.

Optionally, the ratio of the length of the second concave-inward segment to the second convex-outward segment is between 1 and 2.

Optionally, the inlet end and the outlet end of each vane each form a tip structure.

Optionally, the pressure surface and the suction surface are both parallel to the axial direction of the impeller.

In another aspect, the present invention provides a centrifugal compressor comprising an impeller, the impeller being an impeller as described in any one of the above.

The invention designs the blade shape of the impeller of the centrifugal compressor, for example, each blade is in a backward bending type which is gradually bent from the inlet end to the outlet end to deviate from the rotation direction of the impeller, and the thickness of each blade is gradually increased and then gradually decreased in the direction from the inlet end to the outlet end, thereby increasing the reaction degree of the blade, reducing the absolute Mach number of the airflow at the outlet of the blade, reducing the flow loss, obviously improving the pneumatic efficiency of the impeller, enabling the impeller to be more suitable for the working conditions of low flow and high pressure ratio, being suitable for small-sized water chilling units, multi-split air conditioners and other small-sized central air conditioners, and realizing the miniaturization of the centrifugal compressor.

Further, the impeller of the present invention is designed specifically for the structural details of the pressure surface and the suction surface, such as the shape and the installation angle thereof, for example, in the direction from the inlet end to the outlet end, such that the pressure surface of each blade includes a first concave section and a first convex section which are smoothly connected in sequence, and each suction surface includes a second convex section and a second concave section which are smoothly connected in sequence. And further, the installation angles of the pressure surface and the suction surface are gradually increased and then gradually reduced. And the ranges of the installation angles of the pressure surface and the suction surface at the inlet end and the outlet end are particularly limited. The special designs all have obvious effect on improving the pneumatic efficiency of the impeller, so that the impeller is more suitable for working conditions with low flow and high pressure ratio.

The above and other objects, advantages and features of the present invention will become more apparent to those skilled in the art from the following detailed description of specific embodiments thereof, taken in conjunction with the accompanying drawings.

Drawings

Some specific embodiments of the invention will be described in detail hereinafter, by way of illustration and not limitation, with reference to the accompanying drawings. The same reference numbers in the drawings identify the same or similar elements or components. Those skilled in the art will appreciate that the drawings are not necessarily drawn to scale. In the drawings:

FIG. 1 is a profile schematic of an impeller according to an embodiment of the present invention;

FIG. 2 is a schematic view of an angle of incidence marking of a pressure surface of one of the blades in the impeller of FIG. 1;

FIG. 3 is a schematic view of a mounting angle index for the suction side of one of the blades in the impeller of FIG. 1;

FIG. 4 is a schematic structural view of an impeller according to an embodiment of the present invention;

FIG. 5 is an exploded schematic view of the impeller shown in FIG. 4;

FIG. 6 is a schematic structural view of the second impeller body in FIG. 5;

FIG. 7 is a schematic diagram of the overall structure of a centrifugal compressor according to an embodiment of the present invention;

fig. 8 is a schematic sectional view of the centrifugal compressor shown in fig. 7, taken along the axial direction of the impeller;

FIG. 9 is an enlarged view at A of FIG. 8;

FIG. 10 is a schematic diagram of the structure of one of the compression units of FIG. 7;

FIG. 11 is another angular schematic view of the compression unit shown in FIG. 10;

fig. 12 is an exploded schematic view of the compression unit shown in fig. 10.

Detailed Description

A centrifugal compressor and its impeller 200 according to an embodiment of the present invention will be described with reference to fig. 1 to 12. In the partial figures, the axis direction of the impeller 200 is represented by the x axis, and is also the axis direction of the motor 40 and the stator 41 and the rotor 42 thereof; the direction of airflow is indicated by the solid arrows.

FIG. 1 is a profile schematic of an impeller according to an embodiment of the present invention; FIG. 2 is a schematic representation of the mounting angle of the pressure face 241 of one of the blades 240 of the impeller 200 shown in FIG. 1; FIG. 3 is a schematic view of the mounting angle of the suction surface 242 of one of the blades 240 of the impeller of FIG. 1.

The impeller 200 according to the embodiment of the present invention is used in a centrifugal compressor, and as shown in fig. 1, the impeller 200 includes a plurality of blades 240 arranged along a circumferential direction thereof, and a flow passage 250 is formed between each adjacent two of the blades 240. Each vane 240 is of a backward curved type which is gradually curved away from the direction of rotation of the impeller from the inlet end (a end) to the outlet end (B end), the direction of rotation of the impeller being indicated by an arrow in the figure. The inlet end (a-end) refers to the radially inner end of the vane 240, which is also the end adjacent to the inlet side of the flow channel 250. The outlet end (B-end) refers to the radially outer end of the vane 240, which is also the end adjacent to the outlet side of the flow passage 250. In a direction from the inlet end (a end) toward the outlet end (B end), the thickness of each vane 240 gradually increases and then gradually decreases. For example, as shown in fig. 1, the inlet end (a-end) and the outlet end (B-end) of each vane 240 are each formed into a tip structure. In this way, the reaction degree of the blades 240 can be increased, the absolute mach number of the airflow at the outlet of the blades 240 can be reduced, the flow loss can be reduced, the aerodynamic efficiency of the impeller 200 can be remarkably improved, and the impeller is more suitable for the working conditions of low flow and high pressure ratio, so that the centrifugal compressor is suitable for being applied to small central air conditioners such as small water chilling units and multi-split air conditioners, and the centrifugal compressor is miniaturized.

As shown in fig. 1 to 3, the side of each blade 240 facing the rotation direction of the impeller 200 is a pressure surface 241, and the side facing away from the rotation direction is a suction surface 242. The pressure surface 241 and the suction surface 242 may be parallel to the axial direction of the impeller 200.

In the direction from the inlet end (a end) to the outlet end (B end), the pressure surface 241 of each vane 240 includes a first inner concave section AG and a first outer convex section GB which are smoothly connected in sequence, and each suction surface 242 includes a second outer convex section AH and a second inner concave section HB which are smoothly connected in sequence. Specifically, the installation angle of the pressure surface 241 and the suction surface 242 of each blade 240 gradually increases and then gradually decreases in the direction from the inlet end to the outlet end. The installation angle is an angle between a tangent of any one of the pressure surface 241 or the suction surface 242 and a tangent direction of the impeller 200 at the point. The larger the setting angle of a point on the pressure surface 241 or the suction surface 242, the closer the tangent of the point is to the radial direction.

The ratio of the length of the first convex section GB to the length of the first concave section AG is preferably set to 3 to 5, and more preferably set to 3.5 to 4.5. The ratio of the length of the second concave section HB to the length of the second convex section AH is preferably set to 1 to 2, and more preferably set to 1.2 to 1.8.

As shown in fig. 2 and 3, the installation angle α 1 of the pressure surface 241 at the outlet end (the included angle between the tangent BD1 of the pressure surface 241 at the B end and the impeller tangent BC at the B end) is made smaller to achieve the effect of strong backward bending, and preferably, α 1 is made to be between 0 ° and 10 °. The mounting angle α 2 of the suction surface 242 at the outlet end (the angle between tangent lines BD2 and BC of the suction surface 242 at end B) is greater than α 1, preferably such that α 2 is between 20 ° and 40 °. The installation angle β 1 of the pressure surface 241 at the inlet end (the included angle between the tangent AF1 of the pressure surface 241 at the a end and the impeller tangent AE at the a end) can be set to be between 5 ° and 15 °, and the installation angle β 2 of the suction surface 242 at the inlet end (the included angle between the tangent AF2 of the suction surface 242 at the a end and the impeller tangent AE at the a end) is smaller than the installation angle β 1 of the pressure surface 241 at the inlet end thereof and is set to be between 0 ° and 10 °.

The embodiment of the invention specially designs the shapes and the installation angles of the pressure surface 241 and the suction surface 242 of the impeller blade 240, and the heavy spot size is specially limited, so that the impeller 200 has a remarkable effect of improving the pneumatic efficiency of the impeller 200, and the impeller 200 is more suitable for the working conditions of low flow and high pressure ratio.

FIG. 4 is a schematic structural view of an impeller according to an embodiment of the present invention; FIG. 5 is an exploded schematic view of the impeller shown in FIG. 4; fig. 6 is a schematic structural view of the second impeller body in fig. 5.

As shown in fig. 4 to 6, the impeller 200 according to the embodiment of the present invention may be a split type, and specifically, the impeller 200 includes a first impeller body 210 and a second impeller body 220. The second impeller body 220 is provided with an inlet 201 of the impeller 200. The first impeller body 210 and the second impeller body 220 are both disc-shaped and are connected in an abutting joint. The first impeller body 210 is formed with a plurality of first sub-blades 2401, the second impeller body 220 is formed with a plurality of second sub-blades 2402, and the plurality of first sub-blades 2401 and the plurality of second sub-blades 2402 are relatively spliced into a plurality of complete blades 240.

Each of the first sub-blades 2401 may be formed with a screw hole 214, and each of the second sub-blades 2402 may be formed with a screw hole 224, so that the first impeller body 210 and the second impeller body 220 may be coupled and fastened by a plurality of screws 230.

Each first sub-vane 2401 may be provided with a positioning groove 213, each second sub-vane 2402 may be provided with a positioning protrusion 223, and each positioning protrusion 223 may be clamped in one positioning groove 213, so that the position between the first impeller body 210 and the second impeller body 220 may be more stable, and the alignment between each first sub-vane 2401 and each second sub-vane 2402 may be more accurate.

The traditional impeller is of an integral casting type, the surface precision of the traditional impeller is not ideal, and the compression efficiency and the adverse noise of the traditional impeller are affected. Especially for a closed impeller, the blades are inside, and the precision of the blade surface is more difficult to guarantee. In the present embodiment, the impeller 200 is configured as a split type as described above, so that the two impeller bodies are separately manufactured, and the sub-blades of each impeller body are exposed to the outside so as to treat the surfaces of the sub-blades, so that the sub-blades are smoother.

In another aspect, an embodiment of the present invention provides a centrifugal compressor, which includes an impeller, which is the impeller 200 according to any of the above embodiments.

FIG. 7 is a schematic diagram of the overall structure of a centrifugal compressor according to an embodiment of the present invention; fig. 8 is a schematic sectional view of the centrifugal compressor shown in fig. 7, taken along the axial direction of the impeller 200; fig. 9 is an enlarged view of fig. 8 at a.

As shown in fig. 7 to 9, a centrifugal compressor of an embodiment of the present invention may generally include a casing 10, a motor 40, and at least one

compression unit

20, 30.

The cabinet 10 defines an accommodation space, and the motor 40 is installed in the cabinet 10. The motor 40 includes a stator 41 and a rotor 42, the stator 41 is fixed to the housing 10, and the rotor 42 is rotatable relative to the stator 41. The number of the

compression units

20, 30 may be one or more. For example, the centrifugal compressor may be of a single-stage compression type, and only one compression unit may be provided. The centrifugal compressor may be of a multistage compression type in which a plurality of

compression units

20 and 30 are provided. Each

compression unit

20, 30 includes a volute 100 mounted to the casing 10 and an impeller 200 disposed within the volute 100. The impeller 200 is configured to rotate under the drive of the motor 40 to compress the airflow entering the volute 100 and discharge it through the outlet of the volute 100.

In a conventional centrifugal compressor, a diffuser is disposed downstream of an impeller of each stage, the impeller discharges an air flow into the diffuser, and the air flow is diffused by the diffuser and then enters a volute.

Compared with the traditional centrifugal compressor, the centrifugal compressor of the invention omits a diffuser, and the impeller 200 is directly arranged in the volute 100, so that the larger diffusion loss caused by the larger rotation degree of the airflow in the diffuser is avoided, the overall efficiency of the centrifugal compressor is improved, and the structure of the centrifugal compressor is more compact. Therefore, this structure is advantageous for the miniaturization of the centrifugal compressor and for the application to a small central air conditioner such as a small chiller or a multi-split air conditioner while maintaining high efficiency.

In some embodiments, such as shown in fig. 7 and 8, the centrifugal compressor may be of the two-stage compression type, with two compression units. It can be seen that one of the two

compression units

20, 30 is necessarily a low pressure stage, and the other is a high pressure stage, as shown in fig. 7 and 8, the compression unit 20 on the left side of the drawing is a low pressure stage, and the compression unit 30 on the right side is a high pressure stage. 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 a connection pipe 90. Specifically, an inlet end of the connection pipe 90 is provided with a flange 91 to be connected with a flange 130 of an outlet of the volute 100 of the compression unit 20 of the low pressure stage, and an outlet end of the connection pipe 90 is provided with a flange 92 to be connected with the volute 100 of the compression unit 30 of the high pressure stage. It is preferable that the compression unit 20 of the low pressure stage and the compression unit 20 of the high pressure stage are respectively located at both axial sides of the motor 40, so that the impellers 200 of the two

compression units

20 and 30 are respectively directly connected to the motor 40, and it is advantageous to partially cancel the axial forces of the two impellers 200.

In some embodiments, the centrifugal compressor further comprises at least one radial magnetic bearing 60 and/or at least one axial magnetic bearing 80 mounted within the casing 10 to support the rotor 42 of the motor 40. As shown in fig. 8, the centrifugal compressor includes two radial magnetic bearings 60 to support the rotor 42 in a radial direction. The centrifugal compressor further comprises an axial magnetic bearing 80 to counteract the axial force imparted to the rotor 42 by the movement of the impeller 200. The magnetic suspension bearing is made by adopting a magnetic suspension principle and is an oilless bearing. Therefore, lubricating oil does not need to be added into the centrifugal compressor, so that the oil return problem of the compressor of a small and medium-sized refrigeration system is thoroughly solved (the conventional screw compressor, the scroll compressor and the rolling rotor compressor which are usually adopted are basically lubricated by oil), and the heat exchange efficiency of the heat exchanger is improved. And the magnetic suspension bearing is adopted, so that the centrifugal compressor has the advantages of small mechanical wear, low energy consumption, small noise, enhanced stability and longer service life.

Further, as shown in fig. 9, on the basis of the magnetic suspension bearing, a common radial bearing 70 may be further disposed at an axial end of the rotor 42 to support the end of the rotor 42 in an important manner, so that the end is more stable and the operational reliability of the centrifugal compressor is improved.

Fig. 10 is a schematic view of a structure of one of the compressing units 20 of fig. 7; FIG. 11 is another angular schematic view of the compression unit 20 of FIG. 10; fig. 12 is an exploded schematic view of the compressing unit 20 shown in fig. 10.

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 volute 100 defines an inlet flow channel 101, a volute flow channel 102 and an outlet flow channel 103 which are connected in sequence along the airflow direction, namely the volute 100 flow channel is divided into three sections. The inlet of inlet flow path 101 forms the inlet of volute 100 as described herein, and the outlet of outlet flow path 103 forms the outlet of volute 100. The intake runner 101 extends in the axial direction (x-axis direction) of the centrifugal impeller 200. The scroll flow path 102 is flat in a thickness direction parallel to the axial direction of the centrifugal impeller 200. The outlet flow channel 103 gradually changes from a flat shape to a cylindrical shape from the junction with the volute flow channel 102 to the outlet of the volute 100. The centrifugal impeller 200 has an inlet 201 facing the inlet flow channel 101 and an outlet facing the volute flow channel 102, so as to intake air from the inlet flow channel 101, compress the air flow and discharge the compressed air flow to the volute flow channel 102.

In this embodiment, the flat volute flow channel 102 flattens the whole volute 100, which is beneficial to reducing the axial size of the centrifugal compressor and realizing the miniaturization of the compressor. More importantly, as the outlet flow channel 103 gradually changes from flat to cylindrical from the junction with the volute flow channel 102 to the outlet of the volute 100, the gas flow can have a very good diffusion effect in the process of entering the cylindrical and wide outlet flow channel 103 from the thin and flat volute flow channel 102. Moreover, because the outlet flow channel 103 gradually transits from the flat shape to the cylindrical shape from the joint with the volute flow channel 102 to the outlet of the volute 100, the transition is very smooth, unnecessary resistance loss of the air flow is reduced, and the cylindrical shape is also suitable for being connected with a downstream pipeline.

In some embodiments, as shown in fig. 12, the volute 100 may be a split structure including a volute body 110 and a cover plate 120 that are split along the axial direction of the inlet flow channel 101. The volute body 110 defines the aforementioned inlet flow passage 101, a first half of the volute flow passage 102 and an outlet flow passage 103, wherein the volute flow passage 102 is open to one side of the cover plate 120. The cover plate 120 covers an axial side of the volute body 110 to cover an open side of the first half of the scroll 102 and define a second half of the scroll 102, the first half of the scroll 102 and the second half of the scroll 102 opposing each other to form the completed scroll 102.

In the present embodiment, the volute casing 100 is provided as a separate structure, and the volute casing body 110 and the cover plate 120 are separately machined to form the inlet flow passage 101, the volute flow passage 102 and the outlet flow passage 103 by machining. Compared with the existing integrally cast volute, in the embodiment, the surfaces of the inlet flow channel 101, the volute flow channel 102 and the outlet flow channel 103 are smoother, the uniformity of an internal flow field can be better met, the flow loss caused by the fact that the surfaces of the flow channels are too rough is reduced, and the operating efficiency of the centrifugal compressor is improved.

Further, as shown in fig. 9, two planes in the thickness direction of the volute flow channel 102 and the circumferential volute side face may be transited by a fillet (R angle in fig. 9) so as to increase the volute strength, relieve the local stress concentration, eliminate the corner vortex and ensure the uniformity of the flow field. The size of R can be selected based on the thickness of the volute 102. The split configuration of the volute 100 facilitates the machining of the aforementioned rounded corners.

The embodiment of the invention makes the impeller 200 in a strong backward bending type, so that the impeller 200 applies more work to the airflow and converts the work into static pressure lifting, and the work is converted into speed increase. Because the absolute airflow angle of the outlet of the strong backward-bending impeller is larger, if a traditional diffuser form is adopted, the airflow rotation degree is larger, and the diffusion loss is larger. The embodiment of the present invention employs the above-mentioned specially designed volute 100 to directly connect with the impeller 200, which can effectively avoid this problem. It follows that the various improvements of the embodiments of the present invention do not act in isolation from each other, but act in combination. Specifically, in the embodiment of the present invention, the impeller 200 is integrally and directly installed in the volute 100, the flow channel of the volute 100 is specially designed, and the improvements of the impeller 200 with the strong backward bending mode are combined together, so that the beneficial effects of various structural improvements are obtained, the respective adverse effects are greatly avoided, the overall efficiency of the centrifugal compressor is high, the structure is more compact, and the miniaturization is facilitated.

Thus, it should be appreciated by those skilled in the art that while a number of exemplary embodiments of the invention have been illustrated and described in detail herein, many other variations or modifications consistent with the principles of the invention may be directly determined or derived from the disclosure of the present invention without departing from the spirit and scope of the invention. Accordingly, the scope of the invention should be understood and interpreted to cover all such other variations or modifications.

Claims

1. An impeller for a centrifugal compressor comprises a plurality of blades which are arranged along the circumferential direction of the impeller, and a flow passage is formed between every two adjacent blades;

The thickness of each of the vanes is gradually increased and then gradually decreased in a direction from the inlet end toward the outlet end.

2. The impeller of claim 1, wherein

The side surface of each blade facing the rotation direction of the impeller is a pressure surface, and the side surface of each blade facing away from the rotation direction is a suction surface;

in the direction from the inlet end to the outlet end, the pressure surface of each blade comprises a first inward concave section and a first outward convex section which are connected in sequence in a smooth mode, and the suction surface of each blade comprises a second outward convex section and a second inward concave section which are connected in sequence in a smooth mode.

3. An impeller according to claim 2, wherein

In the direction from the inlet end to the outlet end, the installation angle of the pressure surface and the suction surface of each blade gradually increases and then gradually decreases;

the installation angle is an included angle between a tangent of any point of the pressure surface or the suction surface and the tangential direction of the impeller of the point.

4. An impeller according to claim 2, wherein

The installation angle of the pressure surface at the outlet end is between 0 and 10 degrees;

5. An impeller according to claim 2, wherein

The installation angle of the pressure surface at the inlet end is between 5 degrees and 15 degrees;

the installation angle of the suction surface at the inlet end is smaller than that of the pressure surface at the inlet end and is between 0 and 10 degrees.

6. An impeller according to claim 2, wherein

The ratio of the length of the first convex segment to the first concave segment is between 3 and 5.

7. An impeller according to claim 2, wherein

The ratio of the length of the second concave-inward segment to the second convex-outward segment is between 1 and 2.

8. An impeller according to claim 2, wherein

The pressure surface and the suction surface are both parallel to the axial direction of the impeller.

9. The impeller of claim 1, wherein

The inlet and outlet ends of each of the vanes each form a tip structure.

10. A centrifugal compressor comprising an impeller according to any one of claims 1 to 9.