Centrifugal compressor and refrigeration device

Abstract

In a centrifugal compressor (10), a casing (20) has a first wall portion (24) opposite to the back side of an impeller (90). A back side gap (92) is formed between the back side (91) of the impeller (90) and the first wall portion (24). The specific speed of the impeller (90) is set to be less than 0.1. When the width of the back side gap (92) in the axial direction of the shaft (62) is set to an axial width s, and the radius of the impeller (90) is set to an impeller radius r, the ratio of the axial width s of the back side gap (92) to the impeller radius r satisfies the relationship of 0.008≤s / r≤0.5.

Description

Technical Field

[0001] The present invention relates to a centrifugal compressor and a refrigeration device. Background Art

[0002] Heretofore, there is known a centrifugal compressor that compresses a fluid by utilizing the centrifugal force generated by the rotation of an impeller. In a centrifugal compressor, a portion of the fluid pumped by the impeller leaks from the back side of the impeller to a space where a motor is disposed, thereby generating leakage loss. In order to reduce this leakage loss, various sealing structures have been proposed for centrifugal compressors. Patent Document 1 discloses an example of such a sealing structure.

[0003] Prior art literature

[0004] Patent Literature

[0005] Patent Document 1: Japanese Patent Publication No. 2014-231827 Summary of the invention

[0006] -Technical problem to be solved by the invention-

[0007] In existing centrifugal compressors, since the impeller is designed to have a high specific speed, the proportion of windage loss caused by friction with the gas on the back of the impeller is small in the mechanical loss. Therefore, no measures have been taken to improve the windage loss generated on the back of the impeller. However, if the impeller is designed to have a lower specific speed, the proportion of windage loss on the back of the impeller in the mechanical loss will increase. Therefore, the effect of windage loss on the back of the impeller on the compression efficiency cannot be ignored.

[0008] An object of the present disclosure is to improve compression efficiency in a centrifugal compressor designed for a low specific speed.

[0009] -Technical solutions to solve technical problems-

[0010] The first aspect of the present disclosure is directed to a centrifugal compressor 10. The centrifugal compressor 10 of the first aspect includes a housing 20, an impeller 90 housed in the housing 20, and a shaft 62 connected to the impeller 90. The housing 20 has a wall portion 24 opposite to the back side 91 of the impeller 90. The shaft 62 is inserted into a through hole 34 formed on the wall portion 24. A back gap 92 is formed between the back side 91 of the impeller 90 and the wall portion 24. The specific speed of the impeller 90 is set to be less than 0.1. When the width of the back gap 92 in the axial direction of the shaft 62 is set to the axial width s, and the radius of the impeller 90 is set to the impeller radius r, the ratio of the axial width s of the back gap 92 to the impeller radius r satisfies the relationship of 0.008≤s / r≤0.5.

[0011] In this first aspect, the specific speed of the impeller 90 is set to be less than 0.1. This impeller 90 designed with a low specific speed is conducive to achieving small capacity and high head performance. In addition, in the centrifugal compressor 10, the ratio of the axial width s of the back clearance 92 to the impeller radius r (s / r) is greater than 0.008 and less than 0.5. In this way, the back clearance 92 facing the back side 91 of the impeller 90 can have an appropriate axial width s according to the impeller radius r. As a result, the windage loss of the back side 91 of the impeller 90 can be reduced. As a result, in the centrifugal compressor 10 designed with a low specific speed, the compression efficiency can be improved.

[0012] According to a second aspect of the present disclosure, based on the centrifugal compressor 10 according to the first aspect, a first gap 94 in the back gap 92 corresponding to the inner peripheral side of the back side 91 of the impeller 90 constitutes a sealing portion 95 that seals between the back side 91 of the impeller 90 and the wall portion 24. The axial width s of a second gap 96 in the back gap 92 corresponding to the outer peripheral side of the back side of the impeller 90 is larger than the axial width of the first gap 94, and a ratio (s / r) of the axial width s of the second gap 96 to the impeller radius r satisfies the relationship.

[0013] In the second aspect, the first gap 94 corresponding to the inner peripheral side of the back surface 91 of the impeller 90 constitutes a sealing portion 95. The sealing portion 95 seals between the back surface 91 of the impeller 90 and the wall portion 24 of the casing 20. In this way, it is possible to suppress the gas processed in the centrifugal compressor 10 from leaking from the inner peripheral surface of the through hole 34 provided in the wall portion 24 and the shaft 62 through the back surface gap 92 located on the back side of the impeller 90. Therefore, the loss (leakage loss) caused by gas leakage in the centrifugal compressor 10 can be reduced.

[0014] In addition, in the second aspect, the axial width s of the second gap 96 corresponding to the outer peripheral side of the back surface 91 of the impeller 90 is larger than the axial width s of the first gap 94. And the ratio (s / r) of the axial width s of the second gap 96 to the impeller radius r is 0.008 or more and 0.5 or less. The rotation speed of the outer peripheral side of the impeller 90 is higher than the rotation speed of the inner peripheral side, and the windage loss of the back surface 91 of the impeller 90 is large. Therefore, the windage loss of the back surface 91 of the impeller 90 can be effectively reduced.

[0015] A third aspect of the present disclosure is the centrifugal compressor 10 according to the second aspect, wherein the first gap 94 expands in a planar manner in a direction orthogonal to the axis X of the shaft 62 .

[0016] In the third aspect, the first gap 94 expands in a planar manner in a direction orthogonal to the axis X of the shaft 62. The back surface 91 of the impeller 90 and the wall portion 24 of the housing 20 that form such a first gap 94 are opposite to each other in the axial direction extending along the axis X of the shaft 62. In the transient state of the centrifugal compressor 10, there is sometimes a situation where a radial force acts on the shaft 62 and the impeller 90 is displaced in the radial direction together with the shaft 62. If the back surface 91 of the impeller 90 and the wall portion 24 of the housing 20 are opposite to each other in the axial direction extending along the axis X of the shaft 62, even if the impeller 90 is displaced in the radial direction together with the shaft 62, the back surface 91 of the impeller 90 and the wall portion 24 that form the first gap 94 will not contact each other, and the first gap 94 will not be narrowed. Therefore, the reliability of the centrifugal compressor 10 can be improved, and the increase in windage loss of the back surface 91 of the impeller 90 can be suppressed even under transient conditions.

[0017] A fourth aspect of the present disclosure is the centrifugal compressor 10 of any one of the first to third aspects, and further includes a radial bearing 80 supporting the shaft 62 at the outer periphery of the shaft 62. The radial bearing 80 is a foil bearing or a magnetic bearing.

[0018] In the fourth aspect, the radial bearing 80 that supports the shaft 62 on the outer periphery of the shaft 62 is a foil bearing or a magnetic bearing. The foil bearing forms an air film GF between the shaft 62 and the shaft 62, and uses the air film GF to float the shaft 62, thereby supporting the shaft 62 in a non-contact manner, and the shaft 62 is rotatable. The magnetic bearing floats the shaft 62 by magnetic force, and supports the shaft 62 in a non-contact manner, and the shaft 62 is rotatable. The above-mentioned foil bearings and magnetic bearings are beneficial for reducing the frictional heat generated by the rotation of the shaft 62 and the wear of the radial bearing 80. In the centrifugal compressor 10 using the foil bearing or the magnetic bearing as the radial bearing 80, the radial displacement of the shaft 62 generated under transient conditions is large. Therefore, in such a centrifugal compressor 10, the structure involved in the above-mentioned third aspect is effective.

[0019] According to a fifth aspect of the present disclosure, in the centrifugal compressor 10 according to any one of the first to fourth aspects, a maximum rotation speed of the shaft 62 is greater than or equal to 30,000 rpm.

[0020] In the fifth aspect, the maximum speed of the shaft 62 is 30,000 rpm or more, which is relatively high. When the centrifugal compressor 10 has a predetermined specific speed, the higher the maximum speed of the shaft 62, the smaller the capacity of the centrifugal compressor 10 can be, or the higher the head can be. Therefore, the higher the maximum speed of the shaft 62, the better it is to achieve a small capacity and high head performance in the centrifugal compressor 10.

[0021] According to a sixth aspect of the present disclosure, in the centrifugal compressor 10 of any one of the first to fifth aspects, the impeller 90 presses and delivers a refrigerant, which is a HFC (hydrofluorocarbon) refrigerant, a HFO (hydrofluoroolefin) refrigerant, a natural refrigerant, or a mixed refrigerant thereof.

[0022] In the sixth aspect, the refrigerant pumped by the impeller 90 is HFC refrigerant, HFO refrigerant, natural refrigerant or a mixed refrigerant thereof. These refrigerants have a relatively high gas density. The windage loss of the back surface 91 of the impeller 90 increases in proportion to the gas density of the refrigerant handled in the centrifugal compressor 10. Therefore, the technology disclosed in the present invention is effective for a centrifugal compressor that handles HFC refrigerant, HFO refrigerant, natural refrigerant or a mixed refrigerant thereof.

[0023] The seventh aspect of the present disclosure is directed to a refrigeration device 1. The refrigeration device 1 of the seventh aspect includes a refrigerant circuit 2 that performs a refrigeration cycle. The refrigerant circuit 2 includes the centrifugal compressor 10 of any one of the first to sixth aspects.

[0024] In the seventh aspect, the centrifugal compressor 10 described above is used in the refrigerant circuit 2. This contributes to improving the efficiency of the refrigeration cycle performed in the refrigeration device 1. BRIEF DESCRIPTION OF THE DRAWINGS

[0025] Figure 1 is a simplified structural diagram of a refrigerant circuit included in a refrigeration device according to an embodiment;

[0026] Figure 2 is a cross-sectional view showing a schematic structure of a centrifugal compressor according to an exemplary embodiment;

[0027] Figure 3 is a cross-sectional view of a main portion of a centrifugal compressor according to an example embodiment;

[0028] Figure 4 is a cross-sectional view showing a schematic structure of a bearing used in a centrifugal compressor according to an exemplary embodiment;

[0029] Figure 5 is a graph schematically showing the relationship between the efficiency and specific speed of centrifugal, diagonal flow and axial flow impellers;

[0030] Figure 6 is a graph schematically showing the relationship between the specific speed of the impeller and the windage loss ratio;

[0031] Figure 7 is a graph schematically showing the relationship between the specific speed of the impeller and the leakage loss ratio;

[0032] Figure 8is a graph showing the relationship between the ratio of the gap to the circular plate radius and the friction loss coefficient;

[0033] Fig. 9 is a cross-sectional view of a main part of a centrifugal compressor according to a first modified example;

[0034] Fig.10 is a cross-sectional view of a main part of a centrifugal compressor according to a second modified example;

[0035] Fig.11 is a cross-sectional view illustrating a schematic structure of a centrifugal compressor according to a fourth modified example;

[0036] Fig.12 It is a cross-sectional view of a main part of a centrifugal compressor according to another embodiment. DETAILED DESCRIPTION

[0037] Below, exemplary embodiments are described in detail with reference to the accompanying drawings. In the following embodiments, the compressor involved in the technology of the present disclosure is applied to a refrigeration device as an example. It should be noted that the accompanying drawings are used to conceptually illustrate the technology of the present disclosure. Therefore, in the accompanying drawings, in order to facilitate the understanding of the technology of the present disclosure, the size, proportion or quantity is sometimes exaggerated or simplified as needed.

[0038] (Implementation Method)

[0039] The compressor 10 of this embodiment is provided in the refrigeration device 1 .

[0040] - Refrigeration equipment -

[0041] like Figure 1 As shown, the refrigeration device 1 includes a refrigerant circuit 2. The refrigerant circuit 2 is filled with a refrigerant. The fluid compressed in the compressor 10 of this example is a refrigerant. The refrigerant is, for example, an HFC (Hydro Fluoro Carbon) refrigerant such as R32, an HFO (Hydro Fluoro Olefin) refrigerant such as R1234yf, a natural refrigerant, or a refrigerant obtained by mixing the above refrigerants, such as R454C (a mixed refrigerant of R32 and R1234yf). As a natural refrigerant, for example, a natural refrigerant containing HC such as propane can be cited.

[0042] The refrigerant circuit 2 includes a compressor 10, a radiator (condenser) 3, a decompression mechanism 4, and an evaporator 5. The compressor 10, the radiator 3, the decompression mechanism 4, and the evaporator 5 are connected in series through a pipe. The decompression mechanism 4 is, for example, an expansion valve. The refrigerant circuit 2 circulates the refrigerant to perform a vapor compression refrigeration cycle.

[0043] In the refrigeration cycle, the refrigerant compressed by the compressor 10 releases heat to the air in the radiator 3. At this time, the refrigerant is liquefied. The refrigerant after heat release is decompressed by the decompression mechanism 4. The decompressed refrigerant evaporates in the evaporator 5. The evaporated refrigerant is sucked into the compressor 10. The compressor 10 compresses the sucked refrigerant.

[0044] The refrigeration device 1 is, for example, an air conditioning device. The air conditioning device may also be a combined cooling and heating device that switches between cooling and heating. In this case, the air conditioning device has a switching mechanism that switches the circulation direction of the refrigerant. The switching mechanism is, for example, a four-way reversing valve. The air conditioning device may also be a dedicated cooling device or a dedicated heating device.

[0045] In addition, the refrigeration device 1 may also be a water heater, a cooling unit, a cooling device for cooling the air in a storage, etc. The cooling device is a device for cooling the air inside a refrigerator, a freezer, a container, etc.

[0046] -compressor-

[0047] The compressor 10 is a centrifugal compressor 10. The compressor 10 sucks in a low-pressure refrigerant and compresses the refrigerant. The compressor 10 ejects the compressed high-pressure refrigerant. It should be noted that in the following description, the direction extending along the axis X of the shaft 62 of the compressor 10 is referred to as the "axial direction", the direction perpendicular to the axial direction is referred to as the "radial direction", and the direction extending around the shaft 62 is referred to as the "circumferential direction".

[0048] In the compressor 10, the maximum rotation speed of the shaft 62 is above 30,000 rpm. The maximum rotation speed specifies the maximum value of the rotation speed of the motor 60. In the compressor 10, increasing the maximum rotation speed of the shaft 62 is beneficial to increasing the circulation amount of the refrigerant in the refrigerant circuit 2 and ensuring the maximum circulation amount of the refrigerant. This is beneficial to improving the cooling capacity during cooling operation and the heating capacity during heating operation.

[0049] like Figure 2 As shown, the compressor 10 includes a housing 20, a motor 60, a bearing 70, and an impeller 90. The motor 60, the bearing 70, and the impeller 90 are accommodated in the housing 20.

[0050] <Housing>

[0051] The shell 20 is a cylindrical sealed container with both ends closed. The shell 20 is configured such that its center line substantially extends in the horizontal direction. The shell 20 extends in the axial direction. The shell 20 has an internal space 22. The shell 20 has a first wall portion 24 and a second wall portion 26. The first wall portion 24 divides the internal space 22 on one side in the axial direction. The second wall portion 26 divides the internal space 22 on the other side in the axial direction.

[0052] In the internal space 22, one side of the space divided by the first wall portion 24 on the axially outer side constitutes an impeller chamber 28. In the internal space 22, the other side of the space divided by the second wall portion 26 on the axially outer side constitutes a thrust bearing chamber 30. In the internal space 22, the intermediate space divided by the first wall portion 24 and the second wall portion 26 on the axially inner side constitutes a motor chamber 32.

[0053] Through holes 34 and 36 are formed in the first wall portion 24 and the second wall portion 26, respectively. The shaft 62 of the motor 60 is inserted into the two through holes 34 and 36. A gap 38 is provided between the inner peripheral surface of the through hole 34 of the first wall portion 24 and the shaft 62. The gap 38 constitutes a radial seal portion 40. The radial seal portion 40 seals between the first wall portion 24 and the shaft 62 in the axial direction.

[0054] In the housing 20, a diffuser 42 and a vortex flow path 44 are provided on the outer periphery of the impeller chamber 28. The diffuser 42 is formed in an annular shape between the impeller chamber 28 and the vortex flow path 44. The diffuser 42 is defined by a pair of side surfaces that are opposite to each other in the axial direction of the housing 20. The diffuser 42 communicates between the impeller chamber 28 and the vortex flow path 44. The vortex flow path 44 is formed in a vortex shape around the diffuser 42.

[0055] The housing 20 is provided with a suction port 46 and a discharge port 48. The suction port 46 opens at one end of the housing 20 on the impeller chamber 28 side in the axial direction. The suction port 46 is communicated with the central portion of the impeller chamber 28. A suction pipe 50 is connected to the suction port 46. The discharge port 48 is formed at the outer end of the vortex flow path 44. The discharge port 48 is communicated with the vortex flow path 44. A discharge pipe 52 is connected to the discharge port 48.

[0056] <Motor>

[0057] The motor 60 is a driving source for the impeller 90. The motor 60 is housed in the motor chamber 32. The motor 60 is, for example, a permanent magnet synchronous motor. The motor 60 includes a shaft 62, a rotor 64, and a stator 66. The motor 60 is provided in a form in which the axis X direction (axial direction) of the shaft 62 extends in the horizontal direction.

[0058] The shaft 62 is a rod-shaped member that drives the impeller 90. The shaft 62 extends in the inner space 22 in the direction in which the center line of the housing 20 extends. The shaft 62 is inserted into the through holes 34 and 36 formed in the first wall portion 24 and the second wall portion 26. One end of the shaft 62 is located in the impeller chamber 28. The other end of the shaft 62 is located in the thrust bearing chamber 30.

[0059] The rotor 64 is formed in a substantially cylindrical shape. The shaft 62 is inserted into the rotor 64. The rotor 64 is provided at a midway portion of the shaft 62. The rotor 64 is fixed to the shaft 62. The rotor 64 is arranged substantially coaxially with the shaft 62. A plurality of permanent magnets are provided on the rotor 64. The rotor 64 rotates integrally with the shaft 62.

[0060] The stator 66 is formed in a substantially cylindrical shape. The stator 66 is arranged to surround the outer periphery of the rotor 64. The stator 66 is fixed to the inner wall of the housing 20. A coil is wound around the stator 66. The inner peripheral surface of the stator 66 faces the outer peripheral surface of the rotor 64 with a predetermined gap (air gap) in the radial direction.

[0061] The motor 60 rotates the shaft 62 by the interaction of magnetic flux and current between the rotor 64 and the stator 66. A disc portion 68 is provided at the end of the shaft 62 located in the thrust bearing chamber 30. The disc portion 68 is formed in an annular shape extending toward the radial outer side of the shaft 62. The disc portion 68 is arranged substantially coaxially with the shaft 62. The disc portion 68 is a component of the thrust bearing 74.

[0062] <Bearing>

[0063] The compressor 10 includes a pair of bottom bearings 72 , a thrust bearing 74 , and a pair of radial bearings 80 as the bearings 70 .

[0064] The bottom bearing 72 is a rolling bearing. When the motor 60 is de-energized, the bottom bearing 72 supports the shaft 62 so that the shaft 62 can rotate. The bottom bearing 72 receives a radial load acting on the radially outer side of the shaft 62. The bottom bearing 72 is mounted on the inner wall of the housing 20.

[0065] The touchdown bearings 72 are respectively provided on the inner circumferential surface of the through hole 34 of the first wall portion 24 and the inner circumferential surface of the through hole 36 of the second wall portion 26. One touchdown bearing 72 is arranged to surround the outer circumference of the portion of the shaft 62 close to the impeller chamber 28. The other touchdown bearing 72 is arranged to surround the outer circumference of the portion of the shaft 62 close to the thrust bearing chamber 30.

[0066] The thrust bearing 74 is a magnetic bearing. The thrust bearing 74 floats the disc portion 68 of the shaft 62 by electromagnetic force, and supports the disc portion 68 of the shaft 62 in a non-contact manner, and the disc portion 68 is rotatable. The thrust bearing 74 bears the thrust load acting in the axial direction of the shaft 62. The thrust bearing 74 is mounted on the inner wall of the housing 20.

[0067] The thrust bearing 74 is arranged in the thrust bearing chamber 30. The thrust bearing 74 has a pair of electromagnets 76. The pair of electromagnets 76 are each formed in an annular shape. The pair of electromagnets 76 are arranged to face each other across the outer circumference of the disk portion 68 of the shaft 62. Each electromagnet 76 is provided with a gap between the disk portion 68.

[0068] The radial bearing 80 is arranged on both sides of the rotor 64 and the stator 66 in the motor chamber 32. The rotor 64 and the stator 66 divide the motor chamber 32 into a first space 32a and a second space 32b. The first space 32a is a space on the side of the first wall portion 24. The second space 32b is a space on the side of the second wall portion 26. The radial bearing 80 is provided in each of the first space 32a and the second space 32b.

[0069] The radial bearing 80 is held by a holding member 82. The holding member 82 is formed in an approximately disk shape. The outer peripheral surface of the holding member 82 is fixed to the inner wall of the housing 20. A cylindrical portion 83 is provided at the central portion of the holding member 82. A through hole 84 is formed in the cylindrical portion 83 to penetrate the holding member 82. The shaft 62 is inserted into the through hole 84. The radial bearing 80 is accommodated inside the through hole 84.

[0070] The radial bearing 80 supports the shaft 62 at its outer periphery. Figure 4 As shown, the radial bearing 80 is a foil bearing. The radial bearing 80 forms an air film GF between the radial bearing 80 and the shaft 62, and the shaft 62 is floated by the air film GF, thereby supporting the shaft 62 in a non-contact manner, and the shaft 62 is rotatable. The radial bearing 80 receives a radial load acting on the radial outer side of the shaft 62. The radial bearing 80 includes a bearing seat 86, a top foil 88, and a back foil 89.

[0071] The bearing seat 86 is formed in a cylindrical shape. The bearing seat 86 has a through hole 87. The shaft 62 is inserted into the central portion of the through hole 87. A top foil 88 and a back foil 89 are accommodated on the outer periphery of the shaft 62 in the through hole 87. The back foil 89 is located on the inner peripheral surface side of the through hole 87. The top foil 88 is located on the central side of the through hole 87 (the shaft 62 side).

[0072] The top foil 88 is formed in a cylindrical shape. The inner circumferential surface of the top foil 88 faces the outer circumferential surface of the shaft 62 and constitutes a bearing surface. The top foil 88 is made of a thin metal plate and has flexibility. One end portion of the top foil 88 in the circumferential direction is bent toward the outer circumferential side and joined to the back foil 89. In this way, the top foil 88 is fixed to the back foil 89.

[0073] The back foil 89 is formed into a cylindrical shape. The back foil 89 is arranged between the bearing seat 86 and the top foil 88. The back foil 89 is fixed to the bearing seat 86. The back foil 89 elastically supports the top foil 88. The back foil 89 is, for example, a corrugated foil. The back foil 89 may also be a mesh foil.

[0074] A predetermined gap G is set between the top foil 88 and the shaft 62. When the shaft 62 rotates, a gap G is formed between the inner peripheral surface of the top foil 88 and the shaft 62, and gas is introduced into the gap G between the top foil 88 and the shaft 62 to generate a gas film GF. The gas film GF floats the shaft 62 from the top foil 88. In this way, the radial bearing 80 supports the shaft 62 in a non-contact manner.

[0075] <Impeller>

[0076] The impeller 90 is housed in the impeller chamber 28. One end of the shaft 62 is connected to the impeller 90. The impeller 90 is formed in a substantially conical shape. The impeller 90 has a plurality of blades. The first wall portion 24 faces the back surface 91 of the impeller 90. A back surface gap 92 (see FIG. 2 ) is formed between the back surface 91 of the impeller 90 and the first wall portion 24. Figure 3 The back gap 92 is formed in an annular shape. The impeller 90 and the shaft 62 together constitute the rotating unit 100. The impeller 90 rotates integrally with the shaft 62 to pressurize and deliver the refrigerant.

[0077] When the impeller 90 rotates, the refrigerant sucked into the impeller chamber 28 is compressed by the centrifugal force. The refrigerant compressed in the impeller chamber 28 flows in the vortex flow path 44 via the diffuser 42. The refrigerant flowing through the vortex flow path 44 is ejected from the ejection port 48 to the ejection pipe 52. The impeller chamber 28 is pressurized by the rotation of the impeller 90. At this time, the air pressure in the impeller chamber 28 becomes relatively high, and the air pressure in the motor chamber 32 becomes relatively low.

[0078] The specific speed Ns of the impeller 90 is set to be less than 0.1. A dimensionless specific speed is used as the specific speed Ns. The specific speed Ns is calculated based on the following formula (1).

[0079] [Mathematical formula 1]

[0080]

[0081] Where, “N” is the rotation speed [s-1]. “G” is the fluid mass flow rate [kg / s]. “ρ” is the fluid density [kg / m 3 ]. “gH” is the effective specific work [J / kg]. “g” is the acceleration due to gravity [m / s2]. “H” is the head [m].

[0082] like Figure 5 As shown, the value of the specific speed Ns is roughly determined by the structural type of the impeller 90, centrifugal, diagonal flow, or axial flow. In addition, the efficiency value that can be achieved according to the value of the specific speed Ns is also known empirically. Compared with the diagonal flow impeller and the axial flow impeller, the centrifugal impeller 90 has an efficiency advantage at a lower specific speed. The centrifugal impeller 90 has the highest efficiency in the range of the specific speed Ns greater than 0.1.

[0083] As for the centrifugal impeller 90, when the specific speed Ns is less than 0.1, the smaller the specific speed Ns is, the lower the efficiency is. When the compressor 10 is designed to have a small capacity and a high head, the specific speed Ns of the impeller 90 can be increased by increasing the speed N. However, from the perspective of shaft resonance, etc., there is a limit to increasing the speed N. Therefore, in the compressor 10 of this example, the specific speed Ns of the impeller 90 is designed to be relatively low, that is, less than 0.1.

[0084] In such an impeller 90 designed with a low specific speed, in order to increase the head H, the diameter D of the impeller 90 needs to be increased. The height of the head H is proportional to the product of the speed N and the diameter D of the impeller 90 (H∝N×D). In addition, when the compressor 10 is designed with the same capacity and the same head, as the diameter D of the impeller 90 increases, the windage loss W of the impeller 90 increases. Therefore, Figure 6 As shown, the smaller the specific speed Ns is, the greater the proportion of the windage loss W of the impeller 90 in the mechanical loss of the compressor 10 is.

[0085] The windage loss W of the impeller 90 is a loss caused by frictional resistance between the impeller 90 and the gas. The windage loss W of the impeller 90 is defined by, for example, the following equation (2).

[0086] [Mathematical formula 2]

[0087] W=C·ρ·N 3 ·D 5 …(2)

[0088] Wherein, "C" is the windage loss coefficient. The windage loss coefficient C is determined by the material, structure, friction resistance between the impeller 90 and the gas, etc. As the windage loss coefficient C, any one of the design value, analysis value and measured value can be used. "ρ" is the fluid density [kg / m 3 ]. "N" is the rotation speed [s-1]. "D" is the diameter of the impeller 90 [mm].

[0089] The leakage loss of the compressor 10 is the loss caused by the gas leaking from the impeller chamber 28. The leakage loss increases as the amount of gas leakage increases. The amount of gas leakage is determined by the pressure difference between the impeller chamber 28 and its outside. If the compressor 10 is made smaller in capacity under the same head design, the pressure difference between the impeller chamber 28 and its outside will not change, but the mass flow rate of the mainstream will decrease. Therefore, the proportion of the leakage amount increases, and the leakage loss of the compressor 10 increases. Therefore, if Figure 7 As shown, the smaller the specific rotational speed Ns of the impeller 90 is, the greater the proportion of leakage loss in the mechanical loss of the compressor 10 is.

[0090] As can be seen from the above, in the compressor 10 using the impeller 90 designed to have a low specific speed, the ratio of the windage loss W and the leakage loss of the impeller 90 to the mechanical loss is high. In contrast, in the compressor 10 of this example, the back clearance 92 located at the back surface 91 of the impeller 90 is designed to suppress the increase of the windage loss W of the impeller 90 and improve the sealing performance between the rotating part 100 and the first wall part 24.

[0091] When the width of the back gap 92 in the axial direction of the shaft 62 is defined as the axial width s, the back gap 92 is configured such that the axial width s is different between the inner peripheral side and the outer peripheral side. Thus, with respect to the back gap 92, the axial width s is relatively large on the outer peripheral side where the rotation speed of the impeller 90 is high to reduce the windage loss W, and the inner peripheral side is used as a sealing unit.

[0092] Specifically, if Figure 3 As shown, a recess 93 is formed on the surface of the first wall portion 24 facing the impeller chamber 28. The recess 93 is composed of an annular groove that is recessed toward the motor chamber 32 side. The recess 93 is provided at a position corresponding to the outer peripheral side of the impeller 90. As for the back gap 92, the axial width s is different between the portion corresponding to the recess 93 and the portion without the recess 93. The back gap 92 includes a first gap 94 and a second gap 96. The first gap 94 is the back gap 92 at the portion without the recess 93. The second gap 96 is the back gap 92 at the portion corresponding to the recess 93.

[0093] The first gap 94 is provided at a position corresponding to the inner circumference of the impeller 90. The first gap 94 is formed in an annular shape between a portion of the first wall portion 24 that is closer to the inner circumference than the recessed portion 93 and the back surface 91 of the impeller 90. The axial width s1 of the first gap 94 is relatively small in the back surface gap 92. The first gap 94 constitutes an axial seal portion 95. The axial seal portion 95 seals between the back surface 91 of the impeller 90 and the first wall portion 24 in the radial direction. The axial seal portion 95, together with the radial seal portion 40, suppresses leakage of the gas processed in the compressor 10 from the impeller chamber 28 to the motor chamber 32 through the through hole 34 in the first wall portion 24.

[0094] Under transient conditions of the compressor 10, radial force may act on the shaft 62, causing the impeller 90 to be displaced in the radial direction together with the shaft 62. At this time, if the back surface 91 of the impeller 90 and the first wall portion 24 are opposite to each other in a direction perpendicular to the axis X of the shaft 62, the two 90 and 24 may contact each other. In addition, even if the back surface 91 of the impeller 90 and the first wall portion 24 do not contact each other, the gap between the two 90 and 24 becomes narrower, so the windage loss W of the back surface 91 of the impeller 90 increases.

[0095] Thus, the first gap 94 expands in a planar manner in a direction perpendicular to the axis X of the shaft 62, that is, in the radial direction. The back surface 91 of the impeller 90 and the first wall portion 24 are axially opposed to each other via the first gap 94. Therefore, under transient conditions of the compressor 10, even if the impeller 90 is displaced in the radial direction together with the shaft 62, the first gap 94 does not become narrow, and contact between the back surface 91 of the impeller 90 and the first wall portion 24 can be avoided.

[0096] The second gap 96 is provided at a position corresponding to the outer peripheral side of the impeller 90. The second gap 96 is formed in an annular shape between the back surface 91 of the impeller 90 and the bottom surface of the recessed portion 93. The axial width s2 of the second gap 96 is relatively large in the back surface gap 96. That is, the axial width s2 of the second gap 96 is larger than the axial width s1 of the first gap 94. And, when the radius of the impeller 90 is set to the impeller radius r, the ratio (s / r) of the axial width s of the back surface gap 92 to the impeller radius r satisfies the relationship of 0.008≤s / r≤0.5 at the second gap 96.

[0097] That is, the ratio (s2 / r) of the axial width s2 of the second gap 96 to the impeller radius r is set to be greater than 0.008 and less than 0.5. If the ratio (s2 / r) is less than 0.008, the windage loss W of the impeller 90 will be relatively large. On the other hand, even if the ratio (s2 / r) is greater than 0.5, it is not expected that the effect of reducing the windage loss W of the impeller 90 will be further increased compared to the case where the ratio (s2 / r) is less than 0.5. According to the above description, the ratio (s2 / r) of the axial width s2 of the second gap 96 to the impeller radius r is set to satisfy the above relationship.

[0098] The ratio (s2 / r) of the axial width s2 of the second gap 96 to the impeller radius r is preferably 0.02 or more. This is advantageous from the viewpoint of reducing the windage loss W of the impeller 90. From the same viewpoint, the ratio s2 / r of the axial width s2 of the second gap 96 to the impeller radius r is more preferably 0.06 or more. In addition, the ratio (s2 / r) is more preferably 0.1 or more. Moreover, from the above viewpoint, the ratio (s2 / r) is preferably 0.125 or more.

[0099] like Figure 8 As shown by the dashed line in the middle, for the case where water is used as the friction object, the relationship between the ratio of the axial gap s of the circular plate to the radius r of the circular plate (hereinafter referred to as the radius gap ratio) (s / r) and the friction loss coefficient when the circular plate is rotated is known (the function representing the curve). The higher the friction loss coefficient of the circular plate, the greater the energy loss caused by the rotation of the circular plate. This relationship is obtained for each Reynolds number Re, as shown in Figure 8 As shown by the dashed line, there is a tendency that the higher the Reynolds number Re, the smaller the radius clearance ratio when the friction loss coefficient of the circular plate takes the minimum value.

[0100] The relationship between the radial clearance ratio of the circular plate and the friction loss coefficient is described in the following prior art document 1. In addition, a test method for obtaining the relationship is described in the following prior art document 2.

[0101] Prior art document 1: Roughness Effect on Frictional Resistance of Enclosed Rotating Disks (Nece, RE, and Daily, JW, 1960, ASME J. Basic Eng., 82, pp. 553-560)

[0102] ·Prior Art Document 2: Disk Friction Experiment (Versuche uber Scheibenreibung) (VonProf.Dr.-Ing.Kurt Pantell, Berlin, Forschung auf dem Gebiete desIngenieurwesens, Band 16Dusseldorf 1949 / 50Nr.4)

[0103] The inventors of the present application have found the relationship between the radius gap ratio (s / r) of a circular plate and the friction loss coefficient (wind resistance loss coefficient) by conducting a new test with refrigerant gas as the friction object. As the test, an indirect measurement method was implemented, that is, the circular plate was accommodated in a cylindrical shell in a freely rotatable manner, and the circular plate was made to idle in a state where the shell was filled with refrigerant gas, and the deceleration of the rotation speed was measured, and the friction loss coefficient (wind resistance loss coefficient) of the circular plate was calculated based on the deceleration.

[0104] The refrigerant gas used in this test is a Reynolds number Re of 10 7 R32. In this test, a circular plate with a diameter of 40 mm was used as the circular plate. In addition, the idling speed of the circular plate was set to 55000 rpm and 75000 rpm, and the deceleration from the above speeds was measured. Moreover, in this test, the axial gap s between the circular plate and the housing was adjusted to three points where the radius gap ratio (s / r) reached 0.008, 0.02, and 0.125, and the friction loss coefficient of the circular plate was calculated at these three points.

[0105] This test shows that when the disk is placed in a refrigerant gas (R32), there is a tendency that the larger the radius gap ratio (s / r) of the disk is than 0.008, the smaller the value of the friction loss coefficient of the disk is. 7When water is used as the friction object, the friction loss coefficient takes a minimum value near the radius gap ratio (s / r) of the circular plate of 0.008, but when the refrigerant gas (R32) is used as the friction object, the minimum value of the friction loss coefficient exists on the side where the radius gap ratio (s / r) of the circular plate is larger than 0.008, and further exists on the side larger than 0.125.

[0106] - Features of the embodiment -

[0107] In the compressor 10 of this embodiment, the specific speed Ns of the impeller 90 is set to be less than 0.1. This impeller 90 designed with a low specific speed is conducive to achieving small capacity and high head performance. In addition, in the compressor 10, the ratio (s / r) of the axial width s of the second gap 96 in the back gap 92 to the impeller radius r is greater than 0.008 and less than 0.5 at the second gap 96. In this way, the second gap 96 facing the back side 91 of the impeller 90 can have an appropriate axial width s according to the impeller radius r. As a result, the windage loss W of the back side 91 of the impeller 90 can be reduced. As a result, in the compressor 10 designed with a low specific speed Ns, the compression efficiency can be improved.

[0108] In the compressor 10 of this embodiment, the first gap 94 corresponding to the inner peripheral side of the back surface 91 of the impeller 90 constitutes an axial seal portion 95. The axial seal portion 95 seals between the back surface 91 of the impeller 90 and the first wall portion 24. In this way, it is possible to suppress the gas processed in the compressor 10 from leaking from the inner peripheral surface of the through hole 34 provided in the first wall portion 24 and the shaft 62 through the back gap 92 located on the back side of the impeller 90. Therefore, the loss (leakage loss) caused by gas leakage in the compressor 10 can be reduced.

[0109] In the compressor 10 of this embodiment, the axial width s2 of the second gap 96 corresponding to the outer peripheral side of the back side 91 of the impeller 90 is larger than the axial width s1 of the first gap 94. In addition, the ratio (s2 / r) of the axial width s2 of the second gap 96 to the impeller radius r is 0.008 or more and 0.5 or less. The rotation speed of the outer peripheral side of the impeller 90 is higher than the rotation speed of the inner peripheral side, and the windage loss W of the back side 91 of the impeller 90 is large. Therefore, the windage loss W of the back side 91 of the impeller 90 can be effectively reduced.

[0110] In the compressor 10 of this embodiment, the first gap 94 expands in a planar manner in a direction orthogonal to the axis X of the shaft 62. The back surface 91 of the impeller 90 and the first wall portion 24 that form such a first gap 94 are axially opposed to each other. Therefore, even if the impeller 90 is displaced in the radial direction together with the shaft 62, the back surface 91 of the impeller 90 does not contact the first wall portion 24, and the first gap 94 does not narrow. Therefore, the reliability of the compressor 10 can be improved, and the increase in the windage loss W of the back surface 91 of the impeller 90 can be suppressed even under transient conditions.

[0111] In the compressor 10 of the present embodiment, the radial bearing 80 supporting the shaft 62 is a foil bearing. Such a radial bearing 80 forms an air film GF between the radial bearing 80 and the shaft 62, and the shaft 62 is floated by the air film GF, thereby supporting the shaft 62 in a non-contact manner, and the shaft 62 is rotatable. The radial bearing 80 is advantageous in reducing the frictional heat generated by the rotation of the shaft 62 and the amount of wear of the bearing 80.

[0112] In the compressor 10 using the foil bearing as the radial bearing 80, the radial displacement of the shaft 62 generated under the transient condition is large. Therefore, in such a compressor 10 in particular, as described above, the first gap 94 can be prevented from narrowing even if the impeller 90 is displaced in the radial direction, and therefore, a structure in which the first gap 94 expands in a planar manner in the radial direction is effective.

[0113] In the compressor 10 of this embodiment, the maximum speed of the shaft 62 is 30000 rpm or higher, which is relatively high. When the compressor 10 has a predetermined specific speed Ns, the higher the maximum speed of the shaft 62, the smaller the capacity of the compressor 10 can be, or the higher the head can be. Therefore, the higher the maximum speed of the shaft 62, the more advantageous it is to achieve a small capacity and high head performance in the compressor 10.

[0114] In the compressor 10 of this embodiment, the refrigerant pumped by the impeller 90 is HFC refrigerant, HFO refrigerant, natural refrigerant or a mixed refrigerant thereof. These refrigerants have a relatively high gas density. The windage loss W of the back surface 91 of the impeller 90 increases in proportion to the gas density of the refrigerant handled in the compressor 10. Therefore, the technology disclosed in the present invention is effective for the compressor 10 that handles HFC refrigerant, HFO refrigerant, natural refrigerant or a mixed refrigerant thereof.

[0115] In the refrigeration device 1 of this embodiment, the compressor 10 is used in the refrigerant circuit 2. This contributes to the improvement of the efficiency of the refrigeration cycle performed in the refrigeration device 1.

[0116] - First Modification -

[0117] like Fig. 9As shown, in the compressor 10 of the first modified example, two first gaps 94 and two second gaps 96 are provided as the back gaps 92 .

[0118] A first recess 93a and a second recess 93b are formed as recesses 93 on the first wall portion 24 of the housing 20. The first recess 93a and the second recess 93b are arranged at intervals from each other in the radial direction. The first recess 93a is provided at a position corresponding to a portion including the outer peripheral end of the impeller 90. The second recess 93b is formed in an annular shape with a smaller diameter than the first recess 93a. The second recess 93b is provided on the inner peripheral side of the first wall portion 24 relative to the first recess 93a.

[0119] The first wall portion 24 is provided with a partition 98 between the first recess 93a and the second recess 93b. The partition 98 is formed in an annular shape. The partition 98 partitions the first recess 93a and the second recess 93b. The axial width s of the back gap 92 between the back side 91 of the impeller 90 and the partition 98 is the same as the axial width s of the back gap 92 between the portion of the first wall portion 24 that is closer to the inner peripheral side than the second recess 93b and the back side 91 of the impeller 90.

[0120] The first gap 94 is formed between a portion of the first wall portion 24 that is closer to the inner peripheral side than the second recessed portion 93 and the back surface 91 of the impeller 90. The first gap 94 is also formed between the back surface 91 of the impeller 90 and the partition 98. The second gap 96 is formed between the back surface 91 of the impeller 90 and the bottom surface of the first recessed portion 93a. The second gap 96 is also formed between the back surface 91 of the impeller 90 and the bottom surface of the second recessed portion 93b.

[0121] - Second variant -

[0122] like Fig.10 As shown, in the compressor 10 of the second modified example, one first gap 94 and two second gaps 96 are provided as the back gap 92 .

[0123] A first recess 93a and a second recess 93b are formed as recesses 93 on the first wall 24 of the housing 20. The first recess 93a and the second recess 93b are provided in the radial direction with a partition 98 interposed therebetween, as in the first modification described above. The second recess 93b is provided on the inner peripheral side of the first wall 24 relative to the first recess 93a. The second recess 93b of this example is open toward the shaft 62 side.

[0124] The first gap 94 is formed between the back surface 91 of the impeller 90 and the partition 98. The second gap 96 is formed between the back surface 91 of the impeller 90 and the bottom surface of the first recess 93a. The second gap 96 is also formed between the back surface 91 of the impeller 90 and the bottom surface of the second recess 93b. The axial seal portion 95 formed by the first gap 94 is provided only in the substantially central portion of the radial direction of the impeller 90.

[0125] - Third Modification -

[0126] like Fig.11 As shown, in the compressor 10 of the third modification, the radial bearing 80 is a magnetic bearing. The radial bearing 80 floats the shaft 62 by electromagnetic force and supports the shaft 62 in a non-contact manner so that the shaft 62 can rotate. The radial bearing 80 includes a rotor 110 and a stator 112.

[0127] The rotor 110 is formed into a substantially cylindrical shape. The shaft 62 is inserted into the rotor 110. The rotor 110 is fixed to the shaft 62. The rotor 110 rotates integrally with the shaft 62. The rotor 110 is formed, for example, by stacking ferromagnetic steel plates. The stator 112 is formed into a substantially cylindrical shape. The stator 112 is provided at a predetermined distance from the rotor 110. The stator 112 is held by a holding member 116. The holding member 116 is formed into a ring shape. The holding member 116 is mounted on the inner wall of the housing 20. The stator 112 has an electromagnet 114.

[0128] (Other embodiments)

[0129] In the compressor 10 , for a portion of the back gap 92 or the entire back gap 92 , the ratio (s / r) of the axial width s of the back gap 92 to the impeller radius r only needs to satisfy the relationship of 0.008≤s / r≤0.5.

[0130] For example, Fig.12 As shown, in the compressor 10, the back gap 92 may also be constituted by the second gap 96. In this case, the recess 93 is formed on the first wall portion 24 of the housing 20 over the entire area of ​​the portion corresponding to the back side 91 of the impeller 90. The recess 93 of this example is open toward the shaft 62 side. The ratio of the axial width s of the back gap 92 to the impeller radius r satisfies the relationship of 0.008≤s / r≤0.5 over the entire area of ​​the back gap 92. The axial seal portion 95 is not provided on the back side of the impeller 90.

[0131] The above embodiments and variations are described, but it should be understood that various changes can be made to the embodiments and specific matters without departing from the spirit and scope of the claims. As long as the functions of the objects of the present disclosure are not affected, the above embodiments and variations can also be appropriately combined and replaced.

[0132] - Industrial Applicability -

[0133] In summary, the present disclosure is useful for centrifugal compressors and refrigeration devices.

[0134] - Explanation of symbols -

[0135] 1 Refrigeration unit

[0136] 2 Refrigerant circuit

[0137] 10Compressor (centrifugal compressor)

[0138] 20 Shell

[0139] 24 first wall portion (wall portion)

[0140] 24 Through Holes

[0141] 62 Axis

[0142] 80 Radial bearings

[0143] 90 Impeller

[0144] 91 Back

[0145] 92 Back Clearance

[0146] 94 First Gap

[0147] 95 axial sealing part (sealing part)

[0148] 96 Second Gap

Claims

1. A centrifugal compressor, comprising a casing (20), an impeller (90) housed in the casing (20), and a shaft (62) connected to the impeller (90), wherein the casing (20) has a wall portion (24) opposite to the back side of the impeller (90), the shaft (62) is inserted into a through hole (34) formed on the wall portion (24), and a back side gap (92) is formed between the back side (91) of the impeller (90) and the wall portion (24), characterized in that: A recessed portion (93) is formed on a surface of the wall portion (24) facing the impeller (90). The impeller (90) is located outside the recess (93). The specific speed of the impeller (90) is set to be less than 0.1, When the width of the back gap (92) in the axial direction of the shaft (62) is set to an axial width s, and the radius of the impeller (90) is set to an impeller radius r, the ratio of the axial width s of the back gap (92) to the impeller radius r satisfies the relationship of 0.008≤s / r≤0.

5.

2. The centrifugal compressor according to claim 1, characterized in that: A first gap (94) in the back gap (92) corresponding to the inner peripheral side of the back side (91) of the impeller (90) constitutes a sealing portion (95) for sealing between the back side (91) of the impeller (90) and the wall portion (24). The axial width s of a second gap (96) in the back gap (92) corresponding to the outer peripheral side of the back side (91) of the impeller (90) is larger than the axial width s of the first gap (94), and the ratio of the axial width s of the second gap (96) to the impeller radius r, i.e., s / r, satisfies the relationship.

3. The centrifugal compressor according to claim 2, characterized in that: The first gap (94) expands in a planar manner in a direction orthogonal to the axis (X) of the shaft (62).

4. The centrifugal compressor according to any one of claims 1 to 3, characterized in that: The centrifugal compressor further includes a radial bearing (80) supporting the shaft (62) at the outer periphery of the shaft (62). The radial bearing (80) is a foil bearing or a magnetic bearing.

5. The centrifugal compressor according to any one of claims 1 to 3, characterized in that: The maximum rotation speed of the shaft (62) is above 30,000 rpm.

6. The centrifugal compressor according to any one of claims 1 to 3, characterized in that: The impeller (90) presses and delivers the refrigerant. The refrigerant is a hydrofluorocarbon refrigerant, a hydrofluoroolefin refrigerant, a natural refrigerant or a mixed refrigerant thereof.

7. A refrigeration device comprising a refrigerant circuit (2) for performing a refrigeration cycle, characterized in that: The refrigerant circuit (2) includes the centrifugal compressor (10) according to any one of claims 1 to 3.

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