TURBO REFRIGERATOR

A turbo refrigerator includes a centrifugal compressor configured to compress a refrigerant by rotation of an impeller having a plurality of blades, a condenser configured to cool the compressed refrigerant, a first expansion valve and a second expansion valve connected in series and configured to decompress the refrigerant from the condenser to form two phases of a gas and a liquid, an evaporator configured to evaporate the refrigerant from the second expansion valve, an economizer disposed between the first expansion valve and the second expansion valve and configured to separate the refrigerant into two phases of a gas and a liquid, and an introduction path configured to allow the gas phase separated from the refrigerant in the economizer to flow between a front edge and a rear edge of a blade in a main flow channel between the neighboring blades of the impeller.

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Description
TECHNICAL FIELD

The present invention relates to a turbo refrigerator using a centrifugal compressor.

BACKGROUND ART

Among refrigerators, a turbo refrigerator using a centrifugal compressor is known. Turbo refrigerators are used in a wide range of fields such as large air-conditioners of buildings, cooling facilities in chemical plants, or the like.

In recent years, according to a rise in awareness of environmental problems, high performance through improvements in refrigerating capacity is required in turbo refrigerators as well.

In addition, while high performance is required, in view of lower costs, the number of stages in a compressor should be reduced. Accordingly, even when the number of stages of the compressor is reduced to reduce costs, the refrigerating capacity should be maintained, i.e., a requirement of further improvement of the refrigerating capacity is increased.

Here, in a CO2 refrigerating cycle apparatus disclosed in Patent Literature 1, a gas-liquid separator is disposed between two decompression apparatuses (an expansion valve and a capillary tube) that are connected in series, a gas phase and a liquid phase are separated from a refrigerant passing through the first decompression apparatus, and then only the liquid phase is introduced into the second decompression apparatus to perform compression.

As a result, improvement of a refrigerating capacity R serving as an enthalpy difference of the refrigerant before and after the evaporator is accomplished.

CITATION LIST Patent Literature

[Patent Literature 1] Japanese Unexamined Patent Application, First Publication No. 2006-292229

SUMMARY OF INVENTION Technical Problem

However, the structure disclosed in Patent Literature 1 is limited to a screw compressor, and an example applied to a centrifugal compressor including an impeller is not shown.

Here, in a turbo refrigerator in which a multi-stage centrifugal compressor having a plurality of impellers is applied to a compressor, as the gas phase of the refrigerant separated by the gas-liquid separator is suctioned into a flow channel disposed between the impellers disposed between stages of the compressor, improvement of the refrigerating capacity by the gas-liquid separator has been attempted. For this reason, a number of the installed gas-liquid separators is one less than the number of stages of the compressor, and further improvement of the refrigerating capacity using the gas-liquid separators cannot be expected.

Further, as described above, since the gas phase of the refrigerant from the gas-liquid separator is suctioned into the flow channel between the impellers, for example, when a single stage centrifugal compressor configured to perform compression by one impeller is adapted to a compressor, the gas phase of the refrigerant separated by the gas-liquid separator cannot be suctioned. Accordingly, the gas-liquid separator cannot be easily applied to the single stage centrifugal compressor. For this reason, in the refrigerator using the single stage centrifugal compressor, it is difficult to attempt improvement of the refrigerating capacity using the gas-liquid separator.

In this way, when the gas-liquid separator is used, the number that can be installed is limited to the number of stages of the compressor, and improvement of the refrigerating capacity while reducing the number of stages of the compressor is difficult.

In consideration of the above-mentioned circumstances, the present invention provides a turbo refrigerator capable of improving refrigerating capacity while suppressing an increase in cost and improving performance.

Solution to Problem

(I) According to a first aspect of the present invention, a turbo refrigerator includes a centrifugal compressor, a condenser, a plurality of decompressors, an evaporator, a gas-liquid separator and an introduction path. The centrifugal compressor compresses a refrigerant by rotation of an impeller having a plurality of blades. The condenser cools the compressed refrigerant. The decompressors decompress the refrigerant from the condenser to form two phases of a gas and a liquid and are connected to each other in series in a number greater than the number of stages of the centrifugal compressor. The evaporator evaporates the refrigerant passing through the plurality of decompressors. The gas-liquid separators are disposed between the decompressors and are configured to separate the refrigerant into the two phases of the gas and the liquid. The introduction path is configured to cause the gas phase separated from the refrigerant to flow between a front edge and a rear edge between the neighboring blades in at least one of the gas-liquid separators.

According to the above-mentioned configuration, in the at least one gas-liquid separator, the gas phase separated from the refrigerant is suctioned from the introduction path between the front edge and the rear edge of the blade. For this reason, there is no need to suction the gas phase separated from the refrigerant by the gas-liquid separator between the impellers that are between the stages of the centrifugal compressor. Further, whether the number of stages of the centrifugal compressor is a single stage or multiple stages, the gas-liquid separator can be securely installed regardless of the number of stages of the centrifugal compressor.

Since the gas-liquid separator can cause the refrigerant to be in only the liquid phase state, compression can be performed by the decompressor again. That is, for example, the refrigerating cycle serving as a single stage compression and single stage expansion cycle can become a single stage compression and two-stage expansion cycle. Accordingly, in comparison with the case in which the gas phase is not separated from the refrigerant by the gas-liquid separator, an enthalpy difference of the refrigerant before and after passing the evaporator can be increased, and the refrigerating capacity can be improved. Further, as the gas phase separated from the refrigerant by the gas-liquid separator, is suctioned into the centrifugal compressor, a temperature of the refrigerant in the compressor can be reduced, and the compression efficiency can also be improved.

(2) In the turbo refrigerator of the above-mentioned aspect (1), the introduction path may allow the gas phase to flow closer to the front edge side than an intermediate section between the front edge and the rear edge of the blade.

According to the above-mentioned configuration, as the gas phase is introduced as described above, in particular, the introduction path can speed up a stall region generated at the front edge side around the blade of the impeller, and an operation range of the centrifugal compressor is expanded by improving a surge suppressing effect. Accordingly, the performance can be further improved.

(3) In the turbo refrigerator of the above-mentioned aspect (1) or (2), the introduction path may allow the gas phase to flow in a flowing direction of the refrigerant on a meridional plane of the impeller.

According to the above-mentioned configuration, as the introduction path allows the gas phase to flow as described above, smoothness of a flow of a main stream when the gas phase is mixed with the main stream of the refrigerant flowing through the impeller is not interfered with. Accordingly, the mixing loss is decreased, and the performance of the impeller can be further improved.

(4) In the turbo refrigerator of any one of the above-mentioned aspects (1) to (3), the introduction path may have a guide vane formed parallel to the blade on an inner circumferential surface of the introduction path.

According to the above-mentioned configuration, the gas phase from the gas-liquid separator is suctioned through the introduction path by the above-mentioned guide vane, and flows in the same direction as the circumferential direction along a direction of the flow of the main stream when the gas phase is mixed with the main stream of the refrigerant in the impeller. Accordingly, the performance of the impeller can be improved by reducing the mixing loss without interference with smoothness of the flow of the main stream.

(5) In the turbo refrigerator of any one of the above-mentioned aspects (1) to (4), an end section of the blade side of the introduction path may have a diameter that gradually increases downstream.

According to the above-mentioned configuration, as the diameter of the introduction path increases at the blade side, in a state in which a flow velocity of the gas phase is reduced, the gas phase can be suctioned into the impeller. Accordingly, when the gas phase is mixed with the main stream in the impeller, a decrease in performance of the impeller due to a decrease in mixing loss can be prevented without interference with smoothness of a flow of the main stream.

Advantageous Effects of Invention

According to the above-mentioned turbo refrigerator, as the introduction path is installed between the front edge and the rear edge between the neighboring blades, the gas-liquid separators can be installed without restriction on the number that can be installed from the number of stages of the centrifugal compressor. Accordingly, the refrigerating capacity can be improved while reducing the number of stages of the centrifugal compressor to suppress an increase in cost, and improvement of performance becomes possible.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 is a general system diagram showing a turbo refrigerator according to a first embodiment of the present invention.

FIG. 2 is a cross-sectional view related to a centrifugal compressor in the turbo refrigerator according to the first embodiment of the present invention and showing a periphery of an impeller.

FIG. 3 is a general perspective view of the impeller related to the centrifugal compressor in the turbo refrigerator according to the first embodiment of the present invention.

FIG. 4 is a view related to the turbo refrigerator according to the first embodiment of the present invention and schematically showing a refrigerating cycle.

FIG. 5 is a cross-sectional view showing the periphery of the impeller related to the centrifugal compressor in the turbo refrigerator according to the first embodiment of the present invention, and showing the case in which the impeller is a closed type.

FIG. 6 is a general system diagram showing a first modified example of the turbo refrigerator according to the first embodiment of the present invention.

FIG. 7 is a general system diagram showing a second modified example of the turbo refrigerator according to the first embodiment of the present invention.

FIG. 8 is a general system diagram showing a third modified example of the turbo refrigerator according to the first embodiment of the present invention.

FIG. 9 is a cross-sectional view showing the periphery of the impeller related to the centrifugal compressor in the third modified example of the turbo refrigerator according to the first embodiment of the present invention.

FIG. 10 is a cross-sectional view showing a periphery of an impeller related to a centrifugal compressor in a turbo refrigerator according to a second embodiment of the present invention.

FIG. 11 is a view related to the centrifugal compressor in the turbo refrigerator according to the second embodiment of the present invention when an introduction path is seen from the outside in a radial direction, showing a cross section taken along line A-A of FIG. 10.

FIG. 12 is a cross-sectional view showing a periphery of an impeller related to a centrifugal compressor in a turbo refrigerator according to a third embodiment of the present invention.

FIG. 13 is a cross-sectional view showing the periphery of the impeller related to a centrifugal compressor in a turbo refrigerator according to a first modified example of the turbo refrigerator according to the third embodiment of the present invention.

FIG. 14 is a cross-sectional view showing the periphery of the impeller related to a centrifugal compressor in a turbo refrigerator according to a second modified example of the turbo refrigerator according to the third embodiment of the present invention.

 DESCRIPTION OF EMBODIMENTS

Hereinafter, a turbo refrigerator 1A according to a first embodiment of the present invention will be described.

The turbo refrigerator 1A is a cooling apparatus using a turbo type compressor such as a centrifugal compressor or the like, which is used for an air-conditioning apparatus in a large facility such as an office building or the like.

As shown in FIG. 1, the turbo refrigerator 1A includes a centrifugal compressor 10 configured to compress a refrigerant W, a condenser 11 configured to cool the compressed refrigerant W, a first expansion valve (a decompressor) 12 configured to decompress the refrigerant W from the condenser 11, and an economizer (a gas-liquid separator) 14 configured to separate the refrigerant W from the first expansion valve 12 into two phases of a gas and a liquid.

Further, the turbo refrigerator 1A includes an introduction path 16 configured to introduce a gas phase W1 from the economizer 14 into the centrifugal compressor 10, a second expansion valve (a decompressor) 13 configured to decompress a liquid phase from the economizer 14 again, and an evaporator 15 configured to evaporate the refrigerant W from the second expansion valve 13.

Here, for example, R134a (a hydrofluorocarbon) of alternative Freon or the like is used as the refrigerant W.

As shown in FIG. 2, the centrifugal compressor 10 is attached to a rotary shaft 5 that is rotatable about an axis P. The centrifugal compressor 10 includes an impeller 18 that is rotatable about the axis P together with the rotary shaft 5, and a casing 17 configured to cover the impeller 18 from the outside in the radial direction of the axis P.

The rotary shaft 5 is axially coupled to an electric motor or the like (not shown), and is rotatable about the axis P.

As shown in FIG. 3, the impeller 18 has a disk 20 having a curved surface with a diameter that gradually increases from the inside of the axis P to the outside in the radial direction while being directed downstream from an upstream side, as an upstream surface into which the refrigerant W is introduced, serving as one side in a direction of the axis P (an upper side of FIG. 3), and a plurality of (in the embodiment, 17) blades 21 having a fan shape installed to stand up from the curved surface.

In addition, in the embodiment, the impeller 18 is an open type having no shroud.

A space between the neighboring blades 21 is a main flow channel FC through which the refrigerant W can flow downstream from the upstream side.

The casing 17 is a member configured to cover the impeller 18 from the outside in the radial direction in a state in which a gap is provided between the casing 17 and the impeller 18.

Here, in the embodiment, the centrifugal compressor 10 is a single stage compressor configured to perform insulation compression of the refrigerant W using the one impeller 18.

The condenser 11 cools the refrigerant W compressed by the centrifugal compressor 10 through heat exchange using cooling water or the like such that the refrigerant W becomes a liquid.

The first expansion valve 12 decompresses the liquid refrigerant W from the condenser 11 through insulation expansion such that some of the liquid is evaporated and the refrigerant W is in the gas-liquid two-phase state.

The economizer 14 separates the refrigerant W in the gas-liquid two-phase state in the first expansion valve 12 into the gas phase W1 and the liquid phase.

The introduction path 16 can cause the gas phase W1 separated from the refrigerant W in the gas-liquid two-phase state by the economizer 14 to flow into the main flow channel FC in the impeller 18 of the centrifugal compressor 10. Specifically, the introduction path 16 is installed at the casing 17 of the centrifugal compressor 10 between a front edge 21a serving as an upstream end section of the blade 21 and a rear edge 21b serving as a downstream end section. The introduction path 16 has an introduction port 22 opened at a surface directed toward the impeller 18, and an introduction pipe 23 connecting the introduction port 22 and the economizer 14.

The introduction port 22 is formed to pass through the inside and the outside of the casing 17. An aperture position of the introduction port 22 may be formed closer to the front edge 21a than an intermediate section between the front edge 21a and the rear edge 21b of the blade 21.

Similar to the first expansion valve 12, in the second expansion valve 13, the gas phase W1 is separated by the economizer 14, and only the refrigerant W in the liquid phase is compressed through insulation expansion.

The evaporator 15 evaporates the refrigerant W from the second expansion valve 13 through heat exchange with water or the like such that the refrigerant W is in a saturated vapor state.

In the above-mentioned turbo refrigerator 1A, according to a p-h diagram shown in FIG. 4, as shown by a solid line, the gas refrigerant W starts at a point A and reaches a point B in an isentropic state through insulation compression by the centrifugal compressor 10. After that, the gas refrigerant W is cooled by the condenser 11 to become a saturated liquid at a point C on a saturation curve, and further, the liquid refrigerant W is insulation-expanded by the first expansion valve 12 to reach the gas-liquid two-phase state at a point D.

Here, the gas phase W1 is separated from the refrigerant W passing through the first expansion valve 12 by the economizer 14, and the gas phase W1 is suctioned into the main flow channel FC of the impeller 18 in the centrifugal compressor 10 from the introduction port 22 of the introduction path 16. Accordingly, as only the liquid phase of the refrigerant W remains, the refrigerant W is guided into the second expansion valve 13 with the refrigerant W in the saturated liquid state. That is, the refrigerant W reaches a point E on the saturation curve from the point D of FIG. 4.

Only the refrigerant W in the liquid phase due to the second expansion valve 13, i.e., the liquid refrigerant W, is insulation-expanded again from the point E to reach a point F. Then, the liquid refrigerant W is evaporated by the evaporator 15 from the point F to become the saturated vapor, and reaches the point A on the saturation curve.

In this way, since the gas phase W1 of the refrigerant W can be introduced from the introduction port 22 formed at the casing 17 of the centrifugal compressor 10 into the main flow channel FC of the impeller 18 through the introduction pipe 23 in the introduction path 16, even when the single stage centrifugal compressor is used, the economizer 14 can be installed. That is, an extent of an isobaric change at the point E from the point D in FIG. 4 can be added to a refrigerating cycle.

Here, as shown by a broken line of FIG. 4, in a case that the economizer 14 is not installed, a line segment of the point E from the point D in FIG. 4 is not present. That is, the point F is disposed at the point F1. Accordingly, it will be appreciated that the point F1 is disposed closer to a high enthalpy side than the point F, and a distance R between the point A and the point F is larger than a distance R1 between the point A and the point F1.

Here, in FIG. 4, the refrigerating cycle shown by a broken line is a single stage compression and single stage expansion cycle, and the refrigerating cycle shown by a solid line is a single stage compression two-stage expansion cycle.

In this way, the single stage compression and single stage expansion cycle in which the gas phase W1 is not separated from the refrigerant W can become the single stage compression two-stage expansion cycle when the economizer 14 is installed. As a result, an enthalpy difference of the refrigerant W before and after passing through the evaporator 15 can be increased. That is, R>R1, and an improvement of the refrigerating capacity becomes possible.

Further, as the gas phase W1 separated from the refrigerant W is suctioned into the centrifugal compressor 10 by the economizer 14, since a temperature of the refrigerant W in the centrifugal compressor 10 can be decreased, the compression efficiency can be improved.

In addition, the aperture position of the introduction port 22 is formed between the front edge 21a and the rear edge 21b of the blade 21, preferably closer to the front edge 21a than the intermediate section between the front edge 21a and the rear edge 21b of the blade 21. Accordingly, a stall region generated at the front edge 21a side around the blade 21 can be accelerated. Accordingly, a surge suppressing effect is improved, and an operation range of the centrifugal compressor 10 is expanded.

According to the turbo refrigerator 1A of the embodiment, as the introduction port 22 of the introduction path 16 is formed between the front edge 21a and the rear edge 21b of the blade 21, preferably at the front edge 21a side, and the gas phase W1 of the refrigerant W from the economizer 14 can flow into the main flow channel FC, the economizer 14 can also be installed at the single stage centrifugal compressor. Accordingly, as the centrifugal compressor 10 has a single stage, i.e., the number of stages is decreased, the refrigerating capacity can be improved while suppressing an increase in cost, and further, since the compression efficiency can also be improved, improvement of performance can be accomplished.

Further, the first expansion valve 12 or the second expansion valve 13 may be a capillary tube or the like constituted by, for example, metal capillary tubes.

In addition, in the embodiment, while the case in which the impeller 18 is an open type has been described, for example, a closed type impeller 18A having a shroud 29 may be provided. In this case, as shown in FIG. 5, the introduction port 22 of the introduction path 16 from the economizer 14 is formed at a diaphragm 28 of the outside of the shroud 29.

Then, in this case, the gas phase W1 is suctioned into a gap between the shroud 29 and the diaphragm 28, and suctioned from the upstream side into the main flow channel FC of the impeller 18A through a seal 24.

Here, for example, as shown in FIG. 6, even when the two-stage centrifugal compressor is applied to the centrifugal compressor 10, the economizer 14 described in the embodiment is installed, and the gas phase W1 of the refrigerant W from the economizer 14 at the casing 17 of the centrifugal compressor 10 can be introduced into the main flow channel FC of the impeller 18.

Specifically, three expansion valves 25, 26 and 27 are connected in series, and two economizers 14 are installed between the expansion valves 25, 26 and 27. Then, the introduction pipe 23 from one of the economizers 14 is connected to the introduction port 22 of the one impeller 18, and the introduction pipe 23 from the other economizer 14 is connected to the introduction port 22 of the other impeller 18.

In this way, since there is no need to connect the introduction pipe 23 from the economizer 14 between the stages that are between the impellers 18, even when the centrifugal compressor 10 is the two-stage centrifugal compressor, the two economizers 14 can be installed. That is, the economizers 14 can be installed regardless of the number of stages, the refrigerating capacity can be improved while suppressing an increase in cost, and improvement of performance can be accomplished.

Further, as shown in FIG. 7, when the centrifugal compressor 10 is the two-stage centrifugal compressor, the introduction pipe 23 from one of the economizers 14 may be connected to the introduction port 22 of one of the impellers 18, and the introduction pipe 23 from the other economizer 14 may be connected between the stages between the impellers 18.

Then, as shown in FIG. 8, when the centrifugal compressor 10 is the single stage centrifugal compressor, the centrifugal compressor 10 is not limited to the one economizer 14 and the two expansion valves as described in the embodiment. For example, the two economizers 14 and the three expansion valves 25, 26 and 27 may be installed, the introduction pipes 23 from the two economizers 14 may be connected to the introduction port 22 of the one impeller 18, and the gas phase W1 of the refrigerant W from the economizer 14 may be introduced into the main flow channel FC.

Further, when the two economizers 14 are installed, as shown in FIG. 9, the two or more introduction ports 22 are formed with respect to the one impeller 18 and spaced apart from each other between the front edge 21a and the rear edge 21b of the blade 21. Then, one of the introduction ports 22 may be connected to one of the economizers 14, and the other introduction port 22 may be connected to the other economizer 14.

In addition, three or more economizers 14 and four or more expansion valves may be installed with respect to the one impeller 18. That is, when the number of the expansion valves is set to be one larger than the number of the economizers 14, the number of the economizers 14 installed is set independently of the number of stages of the centrifugal compressor 10. Accordingly, the number of the economizers 14 installed can be selected without limitation from the number of stages of the centrifugal compressor 10, and performance can be further improved through improvement of the refrigerating capacity caused by the economizers 14. The above-mentioned configuration can also be applied to the two-stage centrifugal compressor or the multi-stage centrifugal compressor.

Next, a turbo refrigerator 1B according to a second embodiment of the present invention will be described.

Further, the same components as in the first embodiment are designated by the same reference numerals, and detailed description thereof will be omitted.

In the embodiment, in a centrifugal compressor 30, an introduction path 36 from the economizer 14 to the impeller 18 is different from the introduction path 16 of the first embodiment.

As shown in FIG. 10, the introduction path 36 is constituted by an introduction port 42 formed at the casing 17 of the centrifugal compressor 30, and an introduction pipe 43 connecting the introduction port 42 and the economizer 14. A formation position of the introduction port 42 is, like the first embodiment, between the front edge 21a and the rear edge 21b of the blade 21, preferably closer to the front edge 21a than the intermediate section between the front edge 21a and the rear edge 21b of the blade 21.

Further, as shown in FIG. 11, each of the introduction paths 36 has a guide vane 44 disposed in front of the aperture in the introduction port 42 and extending from an inner circumferential surface 42a thereof throughout the entire height of the introduction port 42. The guide vane 44 is installed in parallel to an extension direction of the blade 21.

In addition, the introduction port 42 is opened in a flowing direction of the refrigerant W on a meridional plane of the impeller 18 in the main flow channel FC. Specifically, as shown in FIG. 10, in order to cause the gas phase W1 to flow in the flowing direction of the refrigerant W, an aperture section of the introduction port 42 is formed in the flowing direction of the refrigerant W. In this case, the introduction port 42 may be smoothly turned in front of the aperture section (see FIG. 10), or may be turned in the middle of the introduction pipe 43.

In the above-mentioned turbo refrigerator 1B, the gas phase W1 of the refrigerant W from the economizer 14 is suctioned into the main flow channel FC in the impeller 18 through the introduction path 36. Then, the gas phase W1 of the refrigerant W from the economizer 14 is mixed with the refrigerant W flowing through the main flow channel FC. Here, the gas phase W1 of the refrigerant W from the economizer 14 flows in the flowing direction of the refrigerant W in the main flow channel FC on the meridional plane of the impeller 18. Further, the gas phase W1 flows in the flowing direction of the refrigerant W in the main flow channel FC and in a circumferential direction by the guide vane 44. Accordingly, mixing loss of the refrigerant W in the main flow channel FC can be reduced without interference with smoothness of a flow of the refrigerant W in the main flow channel FC.

According to the turbo refrigerator 1B of the embodiment, like the first embodiment, improvement of performance can be accomplished while suppressing an increase in cost by reducing the number of stages of the centrifugal compressor 30.

In addition, since the mixing loss when the gas phase W1 of the refrigerant W from the economizer 14 flows into the main flow channel FC can be reduced by the formation direction of the introduction port 42 of the introduction path 36 and the guide vane 44, performance of the impeller 18 can be further improved.

Further, the guide vane 44 may not be formed as long as the formation direction of the introduction port 42 is oriented in the flowing direction of the refrigerant W. In addition, when the guide vane 44 is installed, the formation direction of the introduction port 42 may not be oriented in the flowing direction of the refrigerant W.

Next, a turbo refrigerator 1C according to a third embodiment of the present invention will be described.

Further, the same components as in the first embodiment and the second embodiment are designated by the same reference numerals, and detailed description thereof will be omitted.

In the embodiment, in a centrifugal compressor 50, an introduction path 56 from the economizer 14 to the impeller 18 is different from the introduction path 16 in the first embodiment and the introduction path 36 in the second embodiment.

As shown in FIG. 12, the introduction path 56 is constituted by an introduction port 62 formed at the casing 17 of the centrifugal compressor 50, and an introduction pipe 63 configured to connect the introduction port 62 and the economizer 14. A formation position of the introduction port 62 is, like the first embodiment and the second embodiment, between the front edge 21a and the rear edge 21b of the blade 21, preferably closer to the front edge 21a than the intermediate section between the front edge 21a and the rear edge 21b of the blade 21.

Further, a diameter of the introduction path 56 increases at the aperture side serving as an end section of the blade 21 side of the introduction port 62. That is, the introduction path 56 has an expanded diameter section 64 recessed in a concave shape with a dimension larger than the introduction port 62 to an intermediate position of the introduction port from the aperture toward the inside of the casing 17 when seen in a circumferential direction.

In the above-mentioned turbo refrigerator 1C, the gas phase W1 of the refrigerant W from the economizer 14 is suctioned into the main flow channel FC in the impeller 18 through the introduction path 56, and the gas phase W1 of the refrigerant W from the economizer 14 is mixed with the refrigerant W flowing through the main flow channel FC. Here, as the introduction path 56 has the expanded diameter section 64, a cross-sectional area of the introduction port 62 increases at the aperture side, and the gas phase W1 of the refrigerant W from the economizer 14 flows thereinto in a state in which a flow velocity is decreased. Accordingly, the mixing loss to the refrigerant W in the main flow channel FC can be reduced without interference with smoothness of a flow of the refrigerant W in the main flow channel FC.

According to the turbo refrigerator 1C of the embodiment, like the first embodiment and the second embodiment, improvement of performance can be accomplished while suppressing an increase in cost by reducing the number of stages of the centrifugal compressor 50.

In addition, since the mixing loss when the gas phase W1 of the refrigerant W from the economizer 14 flows into the main flow channel FC can be reduced by the expanded diameter section 64 of the introduction path 56, performance of the impeller 18 can be further improved.

First Modified Example of Third Embodiment

Here, as a first modified example of the above-mentioned third embodiment, as shown in FIG. 13, the expanded diameter section 64 does not have a concave shape but may be formed by a curved surface in which a diameter of an inner circumferential surface 62a of the introduction port 62 gradually increases toward the aperture. In this case, the diameter of the inner circumferential surface 62a smoothly increases without a rapid increase in a cross-sectional area of the introduction port 62. Accordingly, the gas phase W1 of the refrigerant W blown out of the introduction port 62 can flow into the main flow channel FC while exfoliation or the like is suppressed in a state in which the flow velocity is more smoothly reduced.

Second Modified Example of Third Embodiment

In addition, as a second modified example of the above-mentioned third embodiment, as shown in FIG. 14, the expanded diameter section 64 does not have a concave shape but may be formed by a curved surface in which a diameter of the inner circumferential surface 62a gradually increases toward the introduction port 22 only at a rear edge side of the impeller 18. In this case, in a state in which the flow velocity is smoothly reduced, the gas phase W1 of the refrigerant W blown out of the introduction port 62 can flow into the main flow channel FC. In addition, the gas phase W1 can be blown out of the introduction port 62 in the flowing direction of the refrigerant W flowing through the main flow channel FC.

Hereinabove, while embodiments of the present invention have been described in detail, some design changes may be made without departing from the technical scope of the present invention.

For example, the guide vane 44 according to the second embodiment may be applied to the introduction paths 16 and 56 of the first embodiment and the third embodiment.

INDUSTRIAL APPLICABILITY

According to the above-mentioned turbo refrigerator, as the introduction path is installed between the front edge and the rear edge between the neighboring blades, the gas-liquid separators can be installed without limitation on the number that can be installed from the number of stages of the centrifugal compressor. Accordingly, the refrigerating capacity can be improved while suppressing an increase in cost by reducing the number of stages of the centrifugal compressor, and improvement of performance becomes possible.

REFERENCE SIGNS LIST

  • 1A turbo refrigerator
  • 5 rotary shaft
  • 10 centrifugal compressor
  • 11 condenser
  • 12 first expansion valve (decompressor)
  • 13 second expansion valve (decompressor)
  • 14 economizer (gas-liquid separator)
  • 15 evaporator
  • 16 introduction path
  • 17 casing
  • 18 impeller
  • 18A impeller
  • 20 disk
  • 21 blade
  • 21a front edge
  • 21b rear edge
  • 22 introduction port
  • 23 introduction pipe
  • 24 seal
  • 25, 26, 27 expansion valve
  • 28 diaphragm
  • 29 shroud
  • W refrigerant
  • W1 gas phase
  • P axis
  • FC main flow channel
  • 1B turbo refrigerator
  • 30 centrifugal compressor
  • 36 introduction path
  • 42 . . . introduction port
  • 42a inner circumferential surface
  • 43 introduction pipe
  • 44 guide vane
  • 1C turbo refrigerator
  • 50 centrifugal compressor
  • 56 introduction path
  • 62 introduction port
  • 62a inner circumferential surface
  • 63 introduction pipe
  • 64 expanded diameter section

Claims

1. A turbo refrigerator comprising:

a centrifugal compressor configured to compress a refrigerant by rotation of an impeller having a plurality of blades;
a condenser configured to cool the compressed refrigerant;
a plurality of decompressors configured to decompress the refrigerant from the condenser to form two phases of a gas and a liquid and connected to each other in series in a number greater than the number of stages of the centrifugal compressor;
an evaporator configured to evaporate the refrigerant passing through the plurality of decompressors;
gas-liquid separators disposed between the decompressors and configured to separate the refrigerant into the two phases of the gas and the liquid; and
an introduction path configured to cause the gas phase separated from the refrigerant to flow between a front edge and a rear edge between the neighboring blades in at least one of the gas-liquid separators.

2. The turbo refrigerator according to claim 1, wherein the introduction path allows the gas phase to flow closer to the front edge side than an intermediate section between the front edge and the rear edge of the blade.

3. The turbo refrigerator according to claim 1, wherein the introduction path allows the gas phase to flow in a flowing direction of the refrigerant on a meridional plane of the impeller.

4. The turbo refrigerator according to claim 1, wherein the introduction path comprises a guide vane formed parallel to the blade on an inner circumferential surface of the introduction path.

5. The turbo refrigerator according to claim 1, wherein an end section of the blade side of the introduction path has a diameter that gradually increases downstream.

Patent History
Publication number: 20160123639
Type: Application
Filed: Jun 24, 2013
Publication Date: May 5, 2016
Applicant: MITSUBISHI HEAVY INDUSTRIES, LTD. (Tokyo)
Inventors: Jun KOGA (Tokyo), Kenji UEDA (Tokyo)
Application Number: 14/893,405
Classifications
International Classification: F25B 43/00 (20060101); F25B 1/053 (20060101);