WO2014199479A1

WO2014199479A1 - Heat pump device

Info

Publication number: WO2014199479A1
Application number: PCT/JP2013/066313
Authority: WO
Inventors: 啓輔高山; 森下　国博; 徹小出
Original assignee: 三菱電機株式会社
Priority date: 2013-06-13
Filing date: 2013-06-13
Publication date: 2014-12-18
Also published as: JP6075451B2; EP3009767A4; EP3009767B1; JPWO2014199479A1; EP3009767A1

Abstract

The purpose of the present invention is to improve COP for a heat pump device that comprises a compressor having a first discharge path and a second discharge path and for which the mass flow rate of refrigerating machine oil discharged along with a refrigerant from the first discharge path is greater than the mass flow rate of the refrigerating machine oil discharged along with a refrigerant from the second discharge path. This heat pump device comprises: a compressor for which the mass flow rate of a refrigerating machine oil discharged from a first discharge path is greater than the mass flow rate of a refrigerating machine oil discharged from a second discharge path; a first heat exchanger having a first refrigerant heat transfer channel through which the refrigerant and the refrigerating machine oil discharged from the first discharge path pass and a first liquid heat transfer channel through which a liquid passes; and a second heat exchanger having a second refrigerant heat transfer channel through which the refrigerant and the refrigerating machine oil discharged from the second discharge path pass and a second liquid heat transfer channel through which a liquid passes. The total cross-sectional area of the first refrigerant heat transfer channel is larger than the total cross-sectional area of the second refrigerant heat transfer channel.

Description

Heat pump equipment

The present invention relates to a heat pump device.

Patent Document 1 discloses a hot water supply having a gas cooler having a high temperature side refrigerant pipe, a low temperature side refrigerant pipe and a water pipe, and a sealed container, a compression element, an electric element, a suction pipe, a discharge pipe, a refrigerant reintroduction pipe and a refrigerant redischarge pipe. A hot water supply cycle device including a compressor is disclosed. In this apparatus, the suction pipe guides the low-pressure refrigerant directly to the compression element, and the high-pressure refrigerant compressed by the compression element is discharged directly from the discharge pipe to the outside of the sealed container without being discharged into the sealed container. The refrigerant after heat exchange through the pipe is guided into the sealed container from the refrigerant reintroduction pipe, and the refrigerant after passing through the electric element in the sealed container is re-discharged out of the sealed container from the refrigerant re-discharge pipe. Send to refrigerant piping.

Japanese Unexamined Patent Publication No. 2006-132427 Japanese Unexamined Patent Publication No. 2004-108616 Japanese Unexamined Patent Publication No. 2008-309361 Japanese Unexamined Patent Publication No. 2009-168383

In the conventional apparatus described above, refrigerating machine oil is supplied into the compression chamber of the compression element in order to lubricate and seal the sliding portion and reduce friction and gap leakage. For this reason, a large amount of refrigerating machine oil is discharged from the discharge pipe of the compressor together with the compressed refrigerant gas to the outside of the compressor and circulates to the high temperature side refrigerant pipe. On the other hand, the amount of refrigerating machine oil contained in the refrigerant discharged from the refrigerant re-discharge pipe of the compressor is significantly smaller than that of the discharge pipe.

Refrigerator oil has a very high viscosity compared to refrigerant. For this reason, in the conventional apparatus described above, a large amount of refrigerating machine oil circulates in the high-temperature side refrigerant pipe together with the refrigerant, so that the pressure loss of the refrigerant increases. As a result, the discharge pressure of the compressor increases and the input of the compressor increases, so that COP (Coefficient Of Performance) decreases.

The present invention has been made to solve the above-described problems, and includes a compressor having a first discharge passage and a second discharge passage, and a mass flow rate of refrigerating machine oil discharged together with refrigerant from the first discharge passage. An object of the present invention is to improve COP in a heat pump device in which the amount of refrigerant is larger than the mass flow rate of refrigerating machine oil discharged together with refrigerant from the second discharge passage.

The heat pump device according to the present invention has a first discharge passage for discharging refrigerant and refrigeration oil, and a second discharge passage for discharging refrigerant and refrigeration oil, and the mass flow rate of the refrigeration oil discharged from the first discharge passage. Is larger than the mass flow rate of the refrigerating machine oil discharged from the second discharge passage, one or a plurality of first refrigerant heat transfer passages through which the refrigerant discharged from the first discharge passage and the refrigerating machine oil pass, and the liquid A first heat exchanger having one or a plurality of first liquid heat transfer channels passing therethrough and exchanging heat between the first refrigerant heat transfer channel and the first liquid heat transfer channel, and discharged from the second discharge passage. The second refrigerant heat transfer channel and the second liquid heat transfer channel have one or more second refrigerant heat transfer channels through which the refrigerant and refrigerating machine oil pass, and one or more second liquid heat transfer channels through which the liquid passes. A second heat exchanger for exchanging heat with the first refrigerant heat transfer flow The total cross-sectional area of the is larger than the total cross-sectional area of the second refrigerant heat transfer passages.

According to the heat pump device of the present invention, it is possible to reliably suppress the pressure loss of the refrigerant discharged from the first discharge passage where the discharge amount of the refrigerating machine oil is large and the refrigerant of the first heat exchanger through which the refrigerating machine oil circulates. . For this reason, it becomes possible to reduce the input of a compressor and to improve COP.

It is a block diagram which shows the heat pump apparatus of Embodiment 1 of this invention. It is a block diagram which shows the hot water storage type hot-water supply system provided with the heat pump apparatus shown in FIG. It is a perspective view which shows the principal part of the 1st gas cooler with which the heat pump apparatus of Embodiment 1 of this invention is provided. It is sectional drawing which shows the principal part of the 1st gas cooler with which the heat pump apparatus of Embodiment 1 of this invention is provided. It is sectional drawing which expands and shows the principal part of the 1st gas cooler with which the heat pump apparatus of Embodiment 1 of this invention is equipped, and a 2nd gas cooler. It is a figure which shows the temperature change of the refrigerant | coolant and water in the 1st gas cooler and the whole 2nd gas cooler, and the division | segmentation position of a 1st gas cooler and a 2nd gas cooler. It is a figure which shows the density change of the refrigerant | coolant in the 1st gas cooler and the 2nd gas cooler whole. It is a figure which shows the ratio of the refrigerant | coolant pressure loss of a 1st gas cooler and a 2nd gas cooler at the time of making the shape of a 1st gas cooler and a 2nd gas cooler the same except flow path length. It is a block diagram of the conventional heat pump apparatus. It is a figure which shows the relationship between the ratio of the twist pitch p of the 1st torsion pipe | tube, and the internal diameter SRi, and the water side heat transfer coefficient. It is a figure which shows the relationship between the ratio of the twist pitch p of the 1st twisted pipe, and the internal diameter SRi, and the required length of a 1st twisted pipe. It is a figure which shows the relationship between the ratio of the twist pitch p of the 1st twisted tube, and the internal diameter SRi, and the required length of a 1st refrigerant | coolant heat exchanger tube. It is a figure which shows the relationship between the refrigerant | coolant pressure loss of a 1st gas cooler, the ratio of the twist pitch p of a 1st twisted tube, and the internal diameter SRi, and the internal diameter di1 of a 1st refrigerant | coolant heat exchanger tube. It is a figure which shows the relationship between the ratio of the twist pitch p of the 1st torsion pipe | tube of the 1st gas cooler, and the internal diameter SRi, and the length of a 1st torsion pipe | tube in each case shown in FIG. When the value of p / SRi of the first torsion pipe is 1.8, the refrigerant pressure loss of the first gas cooler when the inner diameter ratio di1 / di2 of the first refrigerant heat transfer pipe and the second refrigerant heat transfer pipe is changed. It is a figure which shows a change. It is a figure which shows the change of the heat conductivity of the water side when the torsion pitch of a 1st torsion pipe and the torsion pitch of a 2nd torsion pipe are equal, and the internal diameter SRi of a 1st torsion pipe and a 2nd torsion pipe is equal. is there.

Embodiments of the present invention will be described below with reference to the drawings. In addition, the same code | symbol is attached | subjected to the element which is common in each figure, and the overlapping description is abbreviate | omitted. In the following description, for the sake of simplicity, the channel length may be simply referred to as “length”.
Embodiment 1 FIG.
FIG. 1 is a configuration diagram illustrating a heat pump apparatus according to Embodiment 1 of the present invention. FIG. 2 is a configuration diagram illustrating a hot water storage type hot water supply system including the heat pump device illustrated in FIG. 1. As shown in FIG. 1, the heat pump device 1 of the first embodiment includes a compressor 3, a first gas cooler 4 as a first heat exchanger, a second gas cooler 5 as a second heat exchanger, and expansion. The refrigerant circuit which connected the expansion valve 6 and the evaporator 7 as a means by refrigerant | coolant piping is provided. The first gas cooler 4 has a first refrigerant heat transfer channel and a first liquid heat transfer channel, and exchanges heat between the first refrigerant heat transfer channel and the first liquid heat transfer channel. The second gas cooler 5 has a second refrigerant heat transfer channel and a second liquid heat transfer channel, and exchanges heat between the second refrigerant heat transfer channel and the second liquid heat transfer channel. The heat pump device 1 circulates a liquid serving as a heat medium or an object to be heated through the first liquid heat transfer channel of the first gas cooler 4 and the second liquid heat transfer channel of the second gas cooler 5, and heats the liquid. In the heat pump device of the first embodiment, the liquid to be heated is water. The evaporator 7 in this Embodiment 1 is comprised with the air refrigerant | coolant heat exchanger which performs heat exchange with air and a refrigerant | coolant. The heat pump device 1 according to the first embodiment further includes a blower 8 that blows air to the evaporator 7 and a high-low pressure heat exchanger 9 that performs heat exchange between the high-pressure refrigerant and the low-pressure refrigerant. The heat pump device 1 operates a heat pump cycle (refrigeration cycle) by operating the compressor 3 during a heating operation for heating water.

As shown in FIG. 2, the heat pump device 1 of the first embodiment can be used as a hot water storage type hot water supply system by combining with the tank unit 2. In the tank unit 2, a hot water storage tank 2a for storing hot water and a water pump 2b are installed. The heat pump device 1 and the tank unit 2 are connected via a pipe 11 and a pipe 12 through which water flows and an electric wiring (not shown). One end of the tube 11 is connected to the water inlet 1 a of the heat pump device 1. The other end of the pipe 11 is connected to the lower part of the hot water storage tank 2 a in the tank unit 2. A water pump 2 b is installed in the middle of the pipe 11 in the tank unit 2. One end of the pipe 12 is connected to the water outlet 1 b of the heat pump device 1. The other end of the pipe 12 is connected to the upper part of the hot water storage tank 2 a in the tank unit 2. Instead of the illustrated configuration, the water pump 2b may be disposed in the heat pump device 1.

As shown in FIG. 1, the compressor 3 of the heat pump apparatus 1 includes a sealed container 31, a compression element 32 and an electric element 33 provided in the sealed container 31, a first suction passage 34, and a first discharge passage. 35, a second suction passage 36, and a second discharge passage 37. The low-pressure refrigerant sucked from the first suction passage 34 flows directly into the compression element 32 without being discharged into the internal space 38 of the sealed container 31. The compression element 32 is driven by the electric element 33 and compresses the low-pressure refrigerant into a high-pressure refrigerant. The high-pressure refrigerant compressed by the compression element 32 is discharged directly outside the sealed container 31 through the first discharge passage 35 without being discharged into the internal space 38 of the sealed container 31. The high-pressure refrigerant discharged from the first discharge passage 35 passes through the pipe 10 and flows into the first gas cooler 4. The high-pressure refrigerant that has passed through the first gas cooler 4 passes through the pipe 17 and reaches the second suction passage 36 of the compressor 3. The high-pressure refrigerant sucked into the compressor 3 from the second suction passage 36 is discharged into the internal space 38 of the sealed container 31. In the first embodiment, the compression element 32 is disposed under the electric element 33. The outlet of the second suction passage 36 opens at a height between the electric element 33 and the compression element 32 in the internal space 38 of the sealed container 31. The inlet of the second discharge passage 37 opens at a height above the electric element 33 in the internal space 38 of the sealed container 31. The high-pressure refrigerant released from the outlet of the second suction passage 36 into the internal space 38 of the hermetic container 31 passes through the gap between the rotor 331 and the stator 332 of the electric element 33 and reaches the electric element 33. It is discharged out of the sealed container 31 through the two discharge passages 37. The high-pressure refrigerant discharged from the second discharge passage 37 passes through the pipe 18 and flows into the second gas cooler 5. The high-pressure refrigerant that has passed through the second gas cooler 5 passes through the pipe 19 and reaches the expansion valve 6. The high-pressure refrigerant becomes a low-pressure refrigerant by passing through the expansion valve 6. This low-pressure refrigerant flows into the evaporator 7 through the pipe 20. The low-pressure refrigerant that has passed through the evaporator 7 reaches the first suction passage 34 of the compressor 3 through the pipe 21 and is sucked into the compressor 3. The high / low pressure heat exchanger 9 exchanges heat between the high-pressure refrigerant passing through the pipe 19 and the low-pressure refrigerant passing through the pipe 21. The high-pressure refrigerant discharged from the first discharge passage 35 decreases due to pressure loss while returning to the second suction passage 36 via the first gas cooler 4. For this reason, the pressure PH2 of the high-pressure refrigerant in the internal space 38 of the sealed container 31 is lower than the pressure PH1 of the high-pressure refrigerant discharged from the first discharge passage 35. That is, the discharge pressure PH1 of the first discharge passage 35 is higher than the discharge pressure PH2 of the second discharge passage 37.

The heat pump device 1 includes a water flow path 23 that guides water flowing from the water inlet 1a to the water inlet of the second gas cooler 5, and a water flow path 26 that guides water (hot water) flowing out of the water outlet of the first gas cooler 4 to the water outlet 1b. Are further provided. The water outlet of the second gas cooler 5 is connected to the water inlet of the first gas cooler 4. During the heating operation, the water flowing in from the water inlet 1 a flows into the second gas cooler 5 through the water flow path 23 and is heated by the heat of the refrigerant in the second gas cooler 5. Hot water generated by being heated in the second gas cooler 5 flows into the first gas cooler 4 and is further heated by the heat of the refrigerant in the first gas cooler 4. Hot water that has been heated further by being further heated in the first gas cooler 4 reaches the water outlet 1 b through the water flow path 26, and is sent to the tank unit 2 through the pipe 12.

As the refrigerant, a refrigerant capable of producing high temperature hot water, for example, a refrigerant such as carbon dioxide, R410A, propane, propylene or the like is suitable, but is not particularly limited thereto.

The high-temperature and high-pressure refrigerant gas discharged from the first discharge passage 35 of the compressor 3 decreases in temperature while dissipating heat while passing through the first gas cooler 4. In the first embodiment, the refrigerant whose temperature has decreased while passing through the first gas cooler 4 is sucked into the internal space 38 of the sealed container 31 from the second suction passage 36 to cool the electric element 33. Thereby, the temperature of the electric element 33 and the surface temperature of the sealed container 31 can be lowered. As a result, the motor efficiency of the electric element 33 can be improved, and the heat dissipation loss from the surface of the sealed container 31 can be reduced. The refrigerant gas sucked into the internal space 38 of the hermetic container 31 rises in temperature by removing heat from the electric element 33, and then is discharged from the second discharge passage 37 and flows into the second gas cooler 5. The temperature drops while passing through the heat. The high-pressure refrigerant whose temperature has been lowered passes through the expansion valve 6 after heating the low-pressure refrigerant while passing through the high-low pressure heat exchanger 9. By passing through the expansion valve 6, the refrigerant is decompressed to a low-pressure gas-liquid two-phase state. The refrigerant that has passed through the expansion valve 6 absorbs heat from the outside air while passing through the evaporator 7 and is evaporated into gas. The low-pressure refrigerant exiting the evaporator 7 is heated by the high-low pressure heat exchanger 9 and then sucked into the compressor 3 from the first suction passage 34.

If the high-pressure refrigerant pressure is equal to or higher than the critical pressure, the refrigerant in the first gas cooler 4 and the second gas cooler 5 is radiated by lowering the temperature without undergoing a gas-liquid phase transition in a supercritical state. Further, when the high-pressure refrigerant pressure is lower than the critical pressure, the refrigerant dissipates heat while liquefying. In the first embodiment, it is preferable to set the high-pressure refrigerant pressure to a critical pressure or higher by using carbon dioxide or the like as the refrigerant. When the high-pressure refrigerant pressure is equal to or higher than the critical pressure, the liquefied refrigerant can be reliably prevented from flowing into the internal space 38 of the sealed container 31 from the second suction passage 36. For this reason, it can prevent reliably that the liquefied refrigerant | coolant adheres to the electric element 33, and the rotational resistance of the electric element 33 can be reduced. Further, since the liquefied refrigerant does not flow into the internal space 38 of the sealed container 31 from the second suction passage 36, there is an advantage that the refrigerating machine oil is prevented from being diluted by the refrigerant.

As shown in FIG. 2, a water supply pipe 13 is further connected to the lower part of the hot water storage tank 2 a of the tank unit 2. Water supplied from an external water source such as water supply flows through the water supply pipe 13 into the hot water storage tank 2a and is stored. The hot water storage tank 2a is always maintained in a full water state when water flows in from the water supply pipe 13. In the tank unit 2, a hot water supply mixing valve 2c is further provided. The hot water supply mixing valve 2 c is connected to the upper part of the hot water storage tank 2 a through the hot water discharge pipe 14. In addition, a water supply branch pipe 15 branched from the water supply pipe 13 is connected to the hot water supply mixing valve 2c. One end of a hot water supply pipe 16 is further connected to the hot water supply mixing valve 2c. Although not shown, the other end of the hot water supply pipe 16 is connected to a hot water supply terminal such as a faucet, a shower, or a bathtub.

During the heating operation for boiling the water stored in the hot water storage tank 2a, the water stored in the hot water storage tank 2a is sent to the heat pump device 1 through the pipe 11 by the water pump 2b and heated in the heat pump device 1. It becomes hot water. The hot water generated in the heat pump device 1 returns to the tank unit 2 through the pipe 12, and flows into the hot water storage tank 2a from above. By such heating operation, hot water is stored in the hot water storage tank 2a by forming a temperature stratification in which the upper side is high temperature and the lower side is low temperature.

When hot water is supplied from the hot water supply pipe 16 to the hot water supply terminal, hot water in the hot water storage tank 2 a is supplied to the hot water supply mixing valve 2 c through the hot water supply pipe 14, and low temperature water is supplied to the hot water supply pipe through the water supply branch pipe 15. It is supplied to the mixing valve 2c. The hot water and the low temperature water are mixed by the hot water supply mixing valve 2 c and then supplied to the hot water supply terminal through the hot water supply pipe 16. The hot water supply mixing valve 2c has a function of adjusting the mixing ratio of the hot water and the low temperature water so that the hot water temperature set by the user is obtained.

The heat pump device 1 includes a control unit 50. The control unit 50 is electrically connected to actuators and sensors (not shown) provided in the heat pump device 1 and the tank unit 2 and a user interface device (not shown), respectively, and operates the hot water storage hot water supply system. It functions as a control means for controlling. In FIG. 2, the control unit 50 is installed in the heat pump device 1, but the installation location of the control unit 50 is not limited to the heat pump device 1. The control unit 50 may be installed in the tank unit 2. Moreover, you may make it the structure which distribute | arranges and arrange | positions the control part 50 in the heat pump apparatus 1 and the tank unit 2, and is connected so that communication is mutually possible.

The controller 50 controls the temperature of hot water supplied from the heat pump device 1 to the tank unit 2 (hereinafter referred to as “hot water temperature”) at the target hot water temperature during the heating operation. The target hot water temperature is set to 65 ° C. to 90 ° C., for example. In this Embodiment 1, the control part 50 controls the tapping temperature by adjusting the rotation speed of the water pump 2b. The control unit 50 detects the tapping temperature with a temperature sensor (not shown) provided in the water flow path 26, and when the detected tapping temperature is higher than the target tapping temperature, the rotation speed of the water pump 2b is increased. If the tapping temperature is lower than the target tapping temperature, the water pump 2b is corrected so as to decrease the rotational speed. In this way, the control unit 50 can perform control so that the tapping temperature matches the target tapping temperature. However, the temperature of the discharged hot water may be controlled by controlling the temperature of the refrigerant discharged from the first discharge passage 35 of the compressor 3 or the rotational speed of the compressor 3.

In the lower part of the internal space 38 of the sealed container 31 of the compressor 3 shown in FIG. Refrigerating machine oil is supplied to the compression element 32 from this oil reservoir in order to lubricate and seal the sliding portion and reduce friction and gap leakage. The refrigerating machine oil supplied to the compression element 32 is discharged from the first discharge passage 35 together with the compressed high-temperature and high-pressure refrigerant gas. For this reason, a relatively large amount of refrigerating machine oil is discharged from the first discharge passage 35. The refrigerant gas and the refrigerating machine oil discharged from the first discharge passage 35 are in a gas-liquid two-phase flow, reach the second suction passage 36 via the first gas cooler 4, and are sealed from the second suction passage 36 to the sealed container 31. Is released into the internal space 38.

Refrigerating machine oil has a higher density than refrigerant gas. For this reason, the refrigerating machine oil that has flowed into the internal space 38 of the sealed container 31 from the second suction passage 36 falls due to gravity and accumulates in an oil reservoir below the internal space 38 of the sealed container 31. In this way, the refrigerant and the refrigerating machine oil are separated. However, a part of the refrigerating machine oil is atomized and mixed in the refrigerant gas. Further, when refrigerant and refrigerating machine oil are discharged from the outlet of the second suction passage 36 into the internal space 38 of the sealed container 31, a part of the refrigerating machine oil film is wound up and scattered by the flow of the refrigerant gas. There is also. For this reason, a small amount of refrigerating machine oil is also mixed in the refrigerant gas reaching the electric element 33 through the gap between the rotor 331 and the stator 332 of the electric element 33. A part of the mixed refrigerating machine oil is separated from the refrigerant gas by the centrifugal force generated by the rotation of the rotor 331. The remaining refrigerating machine oil is discharged out of the sealed container 31 through the second discharge passage 37 together with the refrigerant gas. As described above, the mass flow rate of the refrigerating machine oil discharged from the first discharge passage 35 is larger than the mass flow rate of the refrigerating machine oil discharged from the second discharge passage 37. On the other hand, the mass flow rate of the refrigerant discharged from the first discharge passage 35 is equal to the mass flow rate of the refrigerant discharged from the second discharge passage 37.

A large amount of refrigerating machine oil circulates in the first refrigerant heat transfer passage of the first gas cooler 4 together with the refrigerant gas. On the other hand, the amount of refrigerating machine oil circulating in the second refrigerant heat transfer passage of the second gas cooler 5 is smaller than that of the first gas cooler 4. Refrigerating machine oil has a very high viscosity compared to refrigerant. For this reason, if refrigerating machine oil circulates through the 1st gas cooler 4 in large quantities, a refrigerant pressure loss will become large easily. When the refrigerant pressure loss of the first gas cooler 4 increases, the discharge pressure of the compressor 3 increases, and the input of the compressor 3 increases, so that COP (Coefficient Of Performance) decreases. In order to solve this problem, in the first embodiment, the entire cross-sectional area of the first refrigerant heat transfer passage of the first gas cooler 4 through which the refrigerant discharged from the first discharge passage 35 and the refrigerating machine oil pass is expressed as the second discharge passage. It is larger than the entire cross-sectional area of the second refrigerant heat transfer passage of the second gas cooler 5 through which the refrigerant discharged from the refrigerant 37 and the refrigerating machine oil pass.

In this specification, the cross-sectional area of the flow path refers to the area of the fluid flow range in a cross section perpendicular to the fluid flow direction. Further, when there are a plurality of first refrigerant heat transfer channels of the first gas cooler 4, that is, when the refrigerant and the refrigerating machine oil flowing into the first gas cooler 4 are divided into a plurality of first refrigerant heat transfer channels and flow in parallel. In addition, the total cross-sectional area of the first refrigerant heat transfer channel refers to the sum of the cross-sectional areas of the first refrigerant heat transfer channels. Similarly, when there are a plurality of second refrigerant heat transfer channels of the second gas cooler 5, that is, the refrigerant and the refrigerating machine oil flowing into the second gas cooler 5 are divided into a plurality of second refrigerant heat transfer channels and flow in parallel. In this case, the total cross-sectional area of the second refrigerant heat transfer channel is the sum of the cross-sectional areas of the first refrigerant heat transfer channels.

As will be described below, in the first embodiment, the total cross-sectional area of the first refrigerant heat transfer channel of the first gas cooler 4 is compared with the total cross-sectional area of the second refrigerant heat transfer channel of the second gas cooler 5, By increasing the size, an increase in the refrigerant pressure loss of the first gas cooler 4 can be reliably suppressed. As a result, the discharge pressure of the compressor 3 is lowered, the input of the compressor 3 is reduced, and the COP is improved.

FIG. 3 is a perspective view showing a main part of the first gas cooler 4 of the first embodiment. FIG. 4 is a cross-sectional view showing a main part of the first gas cooler 4 of the first embodiment. As shown in FIGS. 3 and 4, the first gas cooler 4 has one first torsion tube 41 and three first refrigerant heat transfer tubes 42. FIG. 4 shows a cross section along the longitudinal direction of the first torsion tube 41. In FIG. 3, for convenience, the three first refrigerant heat transfer tubes 42 are denoted by

reference numerals

42 a, 42 b, and 42 c, respectively. In FIG. 3, the first refrigerant

heat transfer tubes

42 a and 42 c are hatched for the sake of convenience in order to easily distinguish the first refrigerant

heat transfer tubes

42 a, 42 b and 42 c. That is, the hatching in FIG. 3 does not mean a cross section.

In the first gas cooler 4 of the first embodiment, the refrigerant and the refrigerating machine oil flow inside the first refrigerant heat transfer tube 42. That is, the first refrigerant heat transfer pipe 42 forms a first refrigerant heat transfer channel. The first gas cooler 4 of the first embodiment has three first refrigerant

heat transfer tubes

42a, 42b, and 42c, that is, three first refrigerant heat transfer channels. The refrigerant and refrigerating machine oil that have flowed into the first gas cooler 4 are divided into these three first refrigerant

heat transfer tubes

42a, 42b, 42c, that is, three first refrigerant heat transfer passages, and flow in parallel. However, in the present invention, the number of the first refrigerant heat transfer channels of the first gas cooler 4, that is, the first heat exchanger is not limited to three, but may be one, two, or four or more.

The first torsion tube 41 has a spiral groove 411 on the outer periphery thereof. The number of grooves 411 is the same as the number of first refrigerant heat transfer tubes 42. That is, in the first embodiment, the first torsion tube 41 has three grooves 411 that are parallel to each other. In FIG. 3,

reference numerals

411a, 411b, and 411c are assigned to the three grooves 411, respectively. Each

groove

411a, 411b, 411c is continuously spiraling. The first refrigerant

heat transfer tubes

42a, 42b, 42c are fitted in the

grooves

411a, 411b, 411c, respectively, and are wound spirally along the shapes of the

grooves

411a, 411b, 411c. With such a configuration, the contact heat transfer area between the first torsion tube 41 and the first refrigerant heat transfer tube 42 can be increased.

In the first gas cooler 4 of the first embodiment, the first torsion tube 41 forms a first liquid heat transfer channel through which water passes. The number of first torsion pipes 41 of the first gas cooler 4 of the first embodiment, that is, the number of first liquid heat transfer channels is one. However, in the present invention, a plurality of first liquid heat transfer channels are provided in the first gas cooler 4, that is, the first heat exchanger, and a liquid such as water is divided into these first liquid heat transfer channels and flows in parallel. It may be configured.

Water flows from the right to the left in FIGS. 3 and 4 through the inside of the first torsion pipe 41. The refrigerant and the refrigerating machine oil spirally flow inside the first refrigerant heat transfer tube 42 from the left to the right in FIGS. 3 and 4. That is, the flow direction of water and the traveling direction of the refrigerant flowing in a spiral form are opposite to each other, resulting in a counter flow.

In this specification, the inner diameter SRi of the first torsion tube 41 is defined as the length of the portion shown in FIG. That is, the inner diameter SRi of the first torsion tube 41 refers to the inner diameter of the portion where the inner diameter is the smallest in the first torsion tube 41.

FIG. 5 is an enlarged cross-sectional view showing the main parts of the first gas cooler 4 and the second gas cooler 5 of the first embodiment. (1) in FIG. 5 shows the first gas cooler 4, and (2) in FIG. 5 shows the second gas cooler 5. As shown in FIG. 5, the first torsion tube 41 and the first refrigerant heat transfer tube 42 are joined via a heat transfer material 60 such as solder. The second gas cooler 5 includes a second torsion tube 51 and a second refrigerant heat transfer tube 52. The second torsion tube 51 has a spiral groove 511 on the outer periphery thereof. In the second gas cooler 5 of Embodiment 1, a second refrigerant heat transfer channel is formed by the second refrigerant heat transfer tube 52, and a second liquid heat transfer channel is formed by the second torsion tube 51. Since the second gas cooler 5 has substantially the same structure as the first gas cooler 4, the drawings corresponding to FIGS. 3 and 4 are omitted. The above description of the first gas cooler 4 is similarly applied to the second gas cooler 5. FIG. 5 shows a cross section along the longitudinal direction of the first torsion tube 41 or the second torsion tube 51.

As shown in FIG. 5, when the first refrigerant heat transfer tube 42 or the second refrigerant heat transfer tube 52 whose original shape is a circular tube shape is spirally wound around the first torsion tube 41 or the second torsion tube 51. In the state after winding, the cross-sectional shape of the first refrigerant heat transfer tube 42 or the second refrigerant heat transfer tube 52 is not a circle, but is a flat shape or an ellipse that is long in the axial direction of the first torsion tube 41 or the second torsion tube 51. It becomes a shape. In the present specification, the inner diameter di1 of the first refrigerant heat transfer tube 42 or the inner diameter di2 of the second refrigerant heat transfer tube 52 is the inner diameter in a circular state before being wound around the first torsion tube 41 or the second torsion tube 51. Means.

Normally, in the first gas cooler 4 or the second gas cooler 5, a portion that is not wound around the first torsion tube 41 or the second torsion tube 51 around the end of the first refrigerant heat transfer tube 42 or the second refrigerant heat transfer tube 52. Exists. For this reason, in such a portion, the inner diameter di1 of the first refrigerant heat transfer tube 42 or the inner diameter di2 of the second refrigerant heat transfer tube 52 is measured before being wound around the first torsion tube 41 or the second torsion tube 51. Can do.

Instead of the above definition, the shape of the first refrigerant heat transfer tube 42 or the second refrigerant heat transfer tube 52 in the state wound around the first torsion tube 41 or the second torsion tube 51 is regarded as an ellipse, and the major axis of the ellipse And the average value of the short diameter may be treated as the inner diameter di1 of the first refrigerant heat transfer tube 42 or the inner diameter di2 of the second refrigerant heat transfer tube 52.

As shown in FIG. 5, in the first embodiment, the inner diameter di1 of the first refrigerant heat transfer tube 42 of the first gas cooler 4 is made larger than the inner diameter di2 of the second refrigerant heat transfer tube 52 of the second gas cooler 5. Is desirable. Further, it is desirable that the torsion pitch p of the first torsion tube 41 of the first gas cooler 4 is larger than the torsion pitch p2 of the second torsion tube 51 of the second gas cooler 5. In this specification, the torsion pitch p of the first torsion tube 41 of the first gas cooler 4 and the torsion pitch p2 of the second torsion tube 51 of the second gas cooler 5 are respectively the lengths of the portions shown in FIG. Define. That is, the torsion pitch p of the first torsion tube 41 is a distance between the centers of two peaks sandwiching the groove 411 in the cross section along the longitudinal direction of the first torsion tube 41. Similarly, the twist pitch p <b> 2 of the second twisted tube 51 is a distance between the centers of two peaks sandwiching the groove 511 in the cross section along the longitudinal direction of the second twisted tube 51.

In the example described below, a case where carbon dioxide is used as a refrigerant will be described. In the example described below, it is assumed that the number of first refrigerant heat transfer channels of the first gas cooler 4 is equal to the number of second refrigerant heat transfer channels of the second gas cooler 5. FIG. 6 is a diagram illustrating temperature changes of the refrigerant and water in the entire first gas cooler 4 and the second gas cooler 5 and the division positions of the first gas cooler 4 and the second gas cooler 5. The horizontal axis in FIG. 6 is a ratio to the total length of the first torsion tube 41 and the second torsion tube 51 (that is, the sum of the length of the first liquid heat transfer channel and the length of the second liquid heat transfer channel). The origin (0) on the horizontal axis in FIG. 6 represents the water outlet and refrigerant inlet of the first gas cooler 4, and the right end (1) on the horizontal axis represents the water inlet and refrigerant outlet of the second gas cooler 5.

As described above, not only refrigerant gas but also a large amount of refrigerating machine oil circulates in the first refrigerant heat transfer tube 42 of the first gas cooler 4. In the first gas cooler 4, the high-temperature refrigeration oil also exchanges heat with water. However, when the specific heat of the refrigeration oil becomes smaller than the specific heat of the refrigerant gas, there is a concern about a decrease in heating capacity and a reduction in hot water supply efficiency associated therewith. Is done. Regarding the relationship between the temperature of the refrigerant gas and the refrigeration oil and the specific heat, the specific heat of the refrigerant gas greatly increases when the temperature is between 20 ° C and 60 ° C, whereas the specific heat of the refrigeration oil is almost constant regardless of the temperature. It is. In order to prevent a decrease in heating capacity due to a large amount of refrigeration oil contained in the refrigerant gas, it is necessary to make the refrigerant gas contain almost no refrigeration oil in a temperature range where the specific heat of the refrigerant gas greatly increases. is there. As shown in FIG. 6, the temperature of the pinch point at which the temperatures of the refrigerant gas and water are closest is about 50 ° C. Therefore, the upper limit temperature in the range where the specific heat of the refrigerant gas rapidly increases is a temperature obtained by adding 10 ° C. to the pinch point temperature. Therefore, if the outlet temperature of the first refrigerant heat transfer tube 42 of the first gas cooler 4 (≈the temperature of the second suction passage 36) is higher by 10 ° C. or more than the pinch point temperature, a reduction in heating capacity is prevented. can do. If at least the outlet temperature of the first refrigerant heat transfer tube 42 of the first gas cooler 4 is higher than the pinch point temperature, it is possible to prevent a significant decrease in heating capacity. From the above, it is desirable that the division position of the first gas cooler 4 and the second gas cooler 5 is higher than the pinch point where the temperature difference between the refrigerant gas and water is closest. In particular, in Embodiment 1, as shown in FIG. 6, the length of the first torsion tube 41 of the first gas cooler 4 with respect to the total length of the first torsion tube 41 and the second torsion tube 51 is It is desirable to configure so as to correspond to about 10% on the high temperature side.

FIG. 7 is a diagram showing changes in refrigerant density in the entire first gas cooler 4 and second gas cooler 5. The meaning of the horizontal axis in FIG. 7 is the same as the horizontal axis in FIG. As shown in FIG. 7, the refrigerant has a lower density as the temperature is higher.

Here, the pressure loss ΔP of the refrigerant in the refrigerant heat transfer tube is obtained by the following equation 1. Here, in order to simplify the description, the sectional shape of the refrigerant heat transfer tube is circular.
ΔP = λ / di · ρ / 2 · u ² · L (Formula 1)
Where λ: pipe friction coefficient, di [m]: inner diameter of refrigerant heat transfer tube, ρ [kg / m ³ ]: refrigerant density, u [m / s]: refrigerant flow velocity, L [m]: flow path length

Further, assuming that the mass flow rate of the refrigerant is Gr [kg / s] and the flow path cross-sectional area of the refrigerant heat transfer tube is A [m ² ], the refrigerant flow velocity u is obtained by the following

formulas

2 and 3.
u = Gr / (ρ · A) (Formula 2)
A = π / 4 · di ² (Formula 3)

Here, in order to simplify the explanation, it is assumed that the shape and the refrigerant flow rate of the first refrigerant heat transfer tube 42 and the second refrigerant heat transfer tube 52 are constant, and the pipe friction coefficient λ does not change. From the above formula, the refrigerant pressure loss ΔP per unit flow path length is proportional to 1 / ρ.

In the first embodiment, refrigerant gas containing a large amount of refrigerating machine oil circulates in the first gas cooler 4, and refrigerant gas containing little refrigerating machine oil circulates in the second gas cooler 5. When the viscosity of the CO ₂ gas refrigerant in the first gas cooler 4 is 1, the average viscosity ratio of the refrigerating machine oil is 311. Thus, the viscosity of the refrigerating machine oil is very large compared to the viscosity of the CO ₂ gas refrigerant. For this reason, the pressure loss of the refrigerant gas containing a large amount of refrigerating machine oil increases.

The mass flow rate of the refrigerating machine oil is Goil [kg / s]. The oil circulation rate OC [%] of the first gas cooler 4 or the second gas cooler 5 is expressed by the following equation 4.
OC = Goil / (Gr + Goil) × 100 (Formula 4)

The oil circulation rate OC is a ratio of the mass flow rate of the refrigerating machine oil to the sum of the mass flow rate of the refrigerant and the mass flow rate of the refrigerating machine oil. In the rated operation state of the heat pump device 1, the oil circulation rate OC of the first gas cooler 4 is preferably 2% or more, and more preferably 5% or more. Further, in the rated operation state of the heat pump device 1, the oil circulation rate OC of the first gas cooler 4 is preferably 20% or less, and more preferably 10% or less. By setting the oil circulation rate OC of the first gas cooler 4 to be equal to or higher than the lower limit described above, the heat of the high-temperature refrigeration oil in the compressor 3 can be effectively used for heating the water in the first gas cooler 4, Heating capacity can be improved. Moreover, the refrigerant | coolant pressure loss of the 1st gas cooler 4 can be suppressed reliably by making the oil circulation rate OC of the 1st gas cooler 4 or less into the upper limit mentioned above, and the quantity of the refrigerating machine oil in the compressor 3 Can be reliably prevented from being excessively lowered.

In the rated operation state of the heat pump device 1, the oil circulation rate OC of the second gas cooler 5 is preferably 0.01% or more, and more preferably 0.1% or more. Further, in the rated operation state of the heat pump device 1, the oil circulation rate OC of the second gas cooler 5 is preferably 1% or less, and more preferably 0.5% or less. By making the oil circulation rate OC of the second gas cooler 5 equal to or less than the above-described upper limit value, the refrigerant pressure loss of the second gas cooler 5 can be reliably suppressed. Further, if the oil circulation rate OC of the second gas cooler 5 is close to the above lower limit value, there is almost no influence of the refrigerating machine oil, so the oil circulation rate OC of the second gas cooler 5 is further lower than the above lower limit value. There is no need to do. Depending on the operating conditions of the heat pump device 1, the oil circulation rate OC of the second gas cooler 5 may be lower than the lower limit value described above.

When the oil circulation rate OC is about 5% to 10%, when the other conditions are the same as compared with the case where the oil circulation rate OC is 0.5% or less, the refrigerant pressure loss is 1.6. Increases to about 2.0 times.

FIG. 8 is a diagram showing a ratio of refrigerant pressure loss of the first gas cooler 4 and the second gas cooler 5 when the shapes of the first gas cooler 4 and the second gas cooler 5 are the same except for the channel length. FIG. 9 is a configuration diagram of a conventional heat pump apparatus. First, the conventional heat pump apparatus 70 shown in FIG. 9 will be described. Elements common to the heat pump apparatus 1 of the first embodiment are denoted by the same reference numerals, and redundant description is omitted. A heat pump device 70 shown in FIG. 9 includes a compressor 71 having one intake passage and one discharge passage instead of the compressor 3 in the heat pump device 1 of the first embodiment. The heat pump device 70 includes a single gas cooler 72 instead of the first gas cooler 4 and the second gas cooler 5. In the heat pump device 70, the low-pressure refrigerant sucked into the compressor 71 from the pipe 21 is compressed by the compressor 71 to become a high-pressure refrigerant. The high-pressure refrigerant is discharged from the compressor 71, passes through the pipe 10 and the gas cooler 72, and reaches the pipe 19.

In FIG. 8, “the oil circulation rate of the entire gas cooler is 0.5% or less” means that the gas cooler 72 is replaced with the first gas cooler 4 and the second gas cooler 5 as in the conventional heat pump device 70 of FIG. It means a case where the refrigerant after separating the refrigerating machine oil in the airtight container of the compressor 71 is allowed to flow into the gas cooler 72 without being divided. That is, it means the case of the conventional refrigeration cycle in which the refrigerant is not returned between the first gas cooler 4 and the second gas cooler 5 into the sealed container 31 of the compressor 3. In this case, when the refrigerant pressure loss of the entire gas cooler 72 is 1, the ratio of the refrigerant pressure loss of the portion corresponding to 10% of the refrigerant high temperature side of the entire flow length of the gas cooler 72 is: 0.17. The ratio of the remaining refrigerant pressure loss in the portion corresponding to the flow path length of 90% on the low temperature side of the refrigerant is 0.83. As shown in FIG. 7, since the refrigerant density is small on the high temperature side of the refrigerant gas, the ratio of the refrigerant pressure loss occupied by the portion corresponding to 10% of the total flow path length is 17% of the total refrigerant pressure loss. Thus, it becomes larger than the ratio of the channel length.

In FIG. 8, the case where “the first gas cooler has a large oil circulation rate and the second gas cooler has an oil circulation rate of 0.5% or less” means that the first gas cooler 4 has an oil circulation rate of 5% to 10%. Therefore, the refrigerant pressure loss is doubled as compared with the case where the oil circulation rate is 0.5% or less. However, here, 10% of the flow path length on the refrigerant high temperature side corresponds to the first gas cooler 4 with respect to the entire flow path length of the first gas cooler 4 and the second gas cooler 5. In this case, when the refrigerant pressure loss of the entire gas cooler 72 is 1, the ratio of the refrigerant pressure loss of the first gas cooler 4 is 0.17 × 2 = 0.34. Therefore, the ratio of the refrigerant pressure loss in the entire first gas cooler 4 and second gas cooler 5 is 0.34 + 0.83 = 1.17. Thus, when the refrigerant pressure loss per unit channel length is large and the refrigerant pressure loss is doubled on the refrigerant high temperature side, the influence on the refrigerant pressure loss of the entire gas cooler is large. For this reason, the refrigerant | coolant pressure loss of the whole gas cooler becomes 1.17 time compared with the case where an oil circulation rate is small as a whole. Moreover, the ratio which the refrigerant | coolant pressure loss of the 1st gas cooler 4 accounts to the whole is as large as 29%.

Although the first gas cooler 4 has a higher oil circulation rate than the second gas cooler 5, the mainly flowing medium is a refrigerant. For this reason, the form of the heat exchanger which comprises the 1st gas cooler 4 is not an oil cooler type form, but the form of the heat exchanger for normal refrigerant | coolants is preferable. For example, the first gas cooler 4 preferably has a configuration using a torsion tube, similarly to the second gas cooler 5.

From the above, if the oil circulation rate of the first gas cooler 4 is large, the refrigerant pressure loss of the first gas cooler 4 tends to increase, and the discharge pressure of the compressor 3 tends to increase. As a result, the input of the compressor 3 increases and the COP tends to decrease. Therefore, in the first embodiment, the refrigerant pressure loss of the first gas cooler 4 is reduced as follows.

A relationship between the inner diameter di1 of the first refrigerant heat transfer tube 42 of the first gas cooler 4, the flow path length L of the first refrigerant heat transfer tube 42, and the refrigerant pressure loss will be described. The refrigerant pressure loss ΔP in the first refrigerant heat transfer tube 42 has the following proportional relationship from the above equations 1 to 3, provided that the pipe friction coefficient, the refrigerant density, and the refrigerant flow rate are constant.
ΔP∝L / (di1) ⁵

Therefore, in order to reduce the refrigerant pressure loss of the first gas cooler 4, it is advantageous to shorten the flow path length L of the first refrigerant heat transfer tube 42, and to increase the inner diameter di1 of the first refrigerant heat transfer tube 42. It is advantageous to do so.

Next, the effect of increasing the twist pitch p of the first torsion pipe 41 will be described. FIG. 10 is a diagram showing the relationship between the ratio between the twist pitch p and the inner diameter SRi of the first torsion pipe 41 and the heat transfer coefficient on the water side. FIG. 10 shows a change in the heat transfer coefficient on the water side when the inner diameter SRi of the first torsion tube 41 is constant and the torsion pitch p is increased. In FIG. 10, the water-side heat transfer coefficient is expressed as a ratio to the water-side heat transfer coefficient value when the value of p / SRi is 1. As shown in FIG. 10, the heat transfer coefficient on the water side tends to increase as p / SRi increases, that is, as the twist pitch p of the first torsion pipe 41 increases.

FIG. 11 is a diagram showing the relationship between the ratio between the twist pitch p and the inner diameter SRi of the first torsion tube 41 and the required length of the first torsion tube 41. In FIG. 11, when the torsion pitch p is increased while keeping the inner diameter SRi of the first torsion tube 41 constant, the length of the first torsion tube 41 required to obtain the same heat exchange amount is set as a reference length. It is expressed as a ratio to thickness. The first gas cooler 4, which is a torsion tube heat exchanger, has a structure in which the first refrigerant heat transfer tube 42 is wound along the spiral groove 411 of the first torsion tube 41. Therefore, when the torsion pitch p of the first torsion tube 41 is increased, the length of the first refrigerant heat transfer tube 42 wound around the unit length of the first torsion tube 41 decreases, and the first refrigerant heat transfer tube 42 and the first torsion tube 42 The contact area with the one torsion pipe 41 is reduced. For this reason, as the twist pitch p of the first torsion pipe 41 is increased, the length of the first torsion pipe 41 necessary for equalizing the heat exchange amount between the refrigerant and water increases. On the other hand, as shown in FIG. 10, as the twist pitch p is increased, the heat transfer coefficient on the water side is increased, so that the heat exchange efficiency per unit length of the first torsion pipe 41 is increased. From these relationships, the relationship shown in FIG. 11 is obtained.

FIG. 12 is a diagram showing the relationship between the ratio between the twist pitch p and the inner diameter SRi of the first torsion tube 41 and the required length of the first refrigerant heat transfer tube. FIG. 12 shows the length of the first refrigerant heat transfer tube 42 required to obtain the same heat exchange amount when the inner diameter SRi of the first torsion tube 41 is constant and the torsion pitch p is increased. Is expressed as a ratio to the length of the first refrigerant heat transfer tube 42 required when the value of 1 is 1. As described with reference to FIG. 11, the required length of the first torsion tube 41 becomes longer as the torsion pitch p of the first torsion tube 41 is increased. On the other hand, as the twist pitch p of the first torsion tube 41 is increased, the length of the first refrigerant heat transfer tube 42 wound around the unit length of the first torsion tube 41 decreases. As a result, as shown in FIG. 12, the required length of the first refrigerant heat transfer tube 42 is shortened as the twist pitch p of the first twist tube 41 is increased. However, in the region where p / SRi is greater than approximately 1.8, the tendency for the required length of the first refrigerant heat transfer tube 42 to decrease decreases.

Summarizing the above characteristics, if the torsion pitch p of the first torsion tube 41 is increased, the length of the first torsion tube 41 required to obtain the same heat exchange amount is increased, but the heat transfer coefficient on the water side is increased. As the ratio increases, the required length of the first torsion tube 41 increases relatively slowly. And as shown in FIG. 12, since the length of the 1st refrigerant | coolant heat exchanger tube 42 can be shortened effectively as the twist pitch p of the 1st twisted tube 41 is enlarged, the refrigerant | coolant pressure of the 1st gas cooler 4 is shown. This is advantageous in reducing loss.

FIG. 13 is a diagram showing the relationship between the refrigerant pressure loss of the first gas cooler 4, the ratio of the twist pitch p and the inner diameter SRi of the first torsion pipe 41, and the inner diameter di1 of the first refrigerant heat transfer pipe. In FIG. 13 and below, the ratio di1 / di2 of the inner diameter di1 of the first refrigerant heat transfer tube 42 of the first gas cooler 4 to the inner diameter di2 of the second refrigerant heat transfer tube 52 of the second gas cooler is referred to as “inner diameter ratio”. FIG. 13 shows the torsion pitch p of the first torsion pipe 41 for each of the cases where the inner diameter ratio di1 / di2 is set to a plurality of values shown in the figure under the condition that the heat exchange amount of the first gas cooler 4 is constant. The change of the refrigerant | coolant pressure loss of the 1st gas cooler 4 when changing is shown is represented. In FIG. 13, the refrigerant pressure loss of the first gas cooler 4 is expressed as a ratio with respect to the refrigerant pressure loss of the first gas cooler 4 when the values of the inner diameter ratio di1 / di2 and p / SRi are both 1.

FIG. 14 is a diagram showing the relationship between the ratio of the torsion pitch p and the inner diameter SRi of the first torsion tube 41 of the first gas cooler 4 and the length of the first torsion tube 41 in each case shown in FIG. . In FIG. 14, the length of the first torsion tube 41 is expressed as a ratio with respect to the length of the first torsion tube 41 when both the inner diameter ratio di1 / di2 and the p / SRi value are 1. 13 and 14, the ratio between the twist pitch p2 of the second torsion pipe 51 of the second gas cooler 5 and the inner diameter SRi is approximately 1. Further, it is assumed that the inner diameter SRi of the first torsion tube 41 of the first gas cooler 4 is equal to the inner diameter SRi of the second torsion tube 51 of the second gas cooler 5.

As shown in FIG. 13, when the inner diameter ratio di1 / di2 is the same, the refrigerant pressure loss of the first gas cooler 4 decreases as p / SRi increases, that is, the torsion pitch p of the first torsion pipe 41 increases. . When the twist pitch p of the first torsion pipe 41 is the same, the refrigerant pressure loss of the first gas cooler 4 increases as the inner diameter ratio di1 / di2 increases, that is, the inner diameter di1 of the first refrigerant heat transfer pipe 42 increases. Decrease.

As described above, the effect of reducing the refrigerant pressure loss of the first gas cooler 4 increases as the inner diameter ratio di1 / di2 increases, that is, as the inner diameter di1 of the first refrigerant heat transfer tube 42 increases. However, the larger the inner diameter di1 of the first refrigerant heat transfer tube 42, the lower the flow rate of the refrigerant in the first refrigerant heat transfer tube 42 and the lower the heat transfer coefficient in the first refrigerant heat transfer tube 42. Therefore, as shown in FIG. 14, as the inner diameter ratio di1 / di2 is increased, that is, as the inner diameter di1 of the first refrigerant heat transfer tube 42 is increased, the first twisted tube necessary for obtaining the same heat exchange amount. 41 becomes longer. As shown in the figure, the required length of the first torsion tube 41 increases as p / SRi increases, that is, the torsion pitch p of the first torsion tube 41 increases.

When the length of the first torsion pipe 41 of the first gas cooler 4 is increased, the entire gas cooler including the first gas cooler 4 and the second gas cooler 5 is increased, and the housing of the heat pump device 1 may be enlarged. Further, when the length of the first torsion tube 41 of the first gas cooler 4 is increased, the material required for the first torsion tube 41 is increased, so that the weight and cost are increased. In addition, if the first torsion pipe 41 serving as a water flow path becomes excessively long, there may be a concern about an increase in the amount of heat released from the first gas cooler 4 to the outside of the heat pump device 1 or an increase in water-side pressure loss. .

As described above, as the twist pitch p of the first torsion pipe 41 of the first gas cooler 4 is increased, the refrigerant pressure loss of the first gas cooler 4 is reduced, but the length of the first torsion pipe 41 is increased. For this reason, if the torsion pitch p of the first torsion tube 41 is excessively increased, the length of the first torsion tube 41 becomes too long, and the above-described adverse effects may occur. In view of this point, the value of p / SRi, which is the ratio between the twist pitch p and the inner diameter SRi of the first twisted tube 41, is preferably 1.8 or less. As described above, in the region where the value of p / SRi exceeds 1.8, the effect of shortening the required length of the first refrigerant heat transfer tube 42 by increasing the torsion pitch p of the first torsion tube 41 is slowed down. To do. For this reason, in the region where the value of p / SRi exceeds 1.8, when the torsion pitch p of the first torsion tube 41 is further increased, not only the effect of further reducing the refrigerant pressure loss is weak, but also the first torsion It is easy to cause an adverse effect of the length of the tube 41 becoming long. On the other hand, if the value of p / SRi is 1.8 or less, the adverse effect of the length of the first twisted tube 41 can be reliably suppressed.

Further, the value of p / SRi of the first torsion pipe 41 of the first gas cooler 4 is preferably 1.1 or more, more preferably 1.2 or more, and further preferably 1.4 or more. By making the value of p / SRi preferably 1.1 or more, more preferably 1.2 or more, and even more preferably 1.4 or more, the length of the first refrigerant heat transfer tube 42 can be effectively shortened. Yes (see FIG. 12). As a result, the refrigerant pressure loss of the first gas cooler 4 can be more reliably reduced. In summary, the value of p / SRi of the first torsion pipe 41 of the first gas cooler 4 is preferably 1.1 or more and 1.8 or less, more preferably 1.2 or more and 1.8 or less, and 1.4 or more. 1.8 or less is more preferable. By setting the value of p / SRi in such a range, the first torsion pipe 41 is sufficiently increased in the torsion pitch p, and the effect of reducing the refrigerant pressure loss of the first gas cooler 4 is sufficiently increased. A markedly superior effect is obtained in that the adverse effects associated with the length of the tube 41 can be reliably suppressed.

Next, a preferable maximum value of the inner diameter ratio di1 / di2 of the first refrigerant heat transfer tube 42 and the second refrigerant heat transfer tube 52 will be described. FIG. 15 shows the first case where the inner diameter ratio di1 / di2 of the first refrigerant heat transfer tube 42 and the second refrigerant heat transfer tube 52 is changed when the value of p / SRi of the first torsion tube 41 is 1.8. It is a figure which shows the change of the refrigerant | coolant pressure loss of 1 gas cooler. In FIG. 15, the refrigerant pressure loss of the first gas cooler 4 is expressed as a ratio to the sum of the refrigerant pressure loss of the first gas cooler 4 and the refrigerant pressure loss of the second gas cooler 5 (that is, the refrigerant pressure loss of the entire gas cooler). As shown in FIG. 15, as the inner diameter ratio di1 / di2 is larger, that is, as the inner diameter di1 of the first refrigerant heat transfer tube 42 is larger, the refrigerant pressure loss of the first gas cooler 4 decreases, and the refrigerant pressure loss with respect to the refrigerant pressure loss of the entire gas cooler decreases. The ratio of the refrigerant pressure loss of the one gas cooler 4 becomes small. However, as shown in FIG. 14, as the inner diameter ratio di1 / di2 is larger, that is, as the inner diameter di1 of the first refrigerant heat transfer tube 42 is larger, the length of the first torsion tube 41 becomes longer. Further, in the first gas cooler 4 in which a large amount of refrigerating machine oil circulates, if the inner diameter di1 of the first refrigerant heat transfer tube 42 is too large, the flow rate of the refrigerating machine oil may deteriorate as the refrigerant flow rate decreases. As a result, the amount of refrigerating machine oil in the first gas cooler 4 may increase significantly. For these reasons, it is desirable to set the inner diameter ratio di1 of the first refrigerant heat transfer tube 42 of the first gas cooler to a value that is not too large.

As shown in FIG. 6, the channel length of the first gas cooler 4 occupies about 10% of the channel length of the entire gas cooler. Therefore, if the ratio of the refrigerant pressure loss of the first gas cooler 4 to the refrigerant pressure loss of the entire gas cooler can be reduced to about 10%, it can be said that the refrigerant pressure loss of the first gas cooler 4 is sufficiently reduced. The refrigerant pressure loss of the first gas cooler 4 is further reduced, that is, the refrigerant pressure loss per unit flow path length of the first gas cooler 4 is made larger than the refrigerant pressure loss per unit flow path length of the second gas cooler 5. Making it smaller can be said to go too far. As shown in FIG. 15, when the inner diameter ratio di1 / di2 is about 1.4, the ratio of the refrigerant pressure loss of the first gas cooler 4 to the refrigerant pressure loss of the entire gas cooler is about 10%. Therefore, if the value of the inner diameter ratio di1 / di2 is set to 1.4, it can be said that the refrigerant pressure loss of the first gas cooler 4 is sufficiently reduced in relation to the ratio of the flow path length. On the other hand, if the value of the inner diameter ratio di1 / di2 is excessively increased, that is, if the inner diameter di1 of the first refrigerant heat transfer tube 42 is excessively increased, the length of the first torsion tube 41 becomes excessive or the first gas cooler 4 There is a possibility that the above-mentioned detrimental effect of increasing the amount of refrigerating machine oil remaining will occur. On the other hand, if the value of the inner diameter ratio di1 / di2 is 1.4 or less, the inner diameter di1 of the first refrigerant heat transfer tube 42 will not be too large, and such an adverse effect can be reliably suppressed.

Further, the value of the inner diameter ratio di1 / di2 of the first refrigerant heat transfer tube 42 and the second refrigerant heat transfer tube 52 is preferably 1.1 or more, and more preferably 1.2 or more. By setting the value of the inner diameter ratio di1 / di2 to preferably 1.1 or more, more preferably 1.2 or more, the refrigerant pressure loss of the first gas cooler 4 can be more reliably reduced (see FIG. 13). . In summary, the value of the inner diameter ratio di1 / di2 is preferably 1.1 or more and 1.4 or less, and more preferably 1.2 or more and 1.4 or less. By setting the value of the inner diameter ratio di1 / di2 in such a range, the refrigerant pressure of the first gas cooler 4 can be surely suppressed while preventing the above-described adverse effects caused by making the inner diameter di1 of the first refrigerant heat transfer tube 42 too large. A remarkable effect is obtained that the loss can be sufficiently reduced.

As described above, according to the first embodiment, the refrigerant pressure loss of the first gas cooler 4 can be reliably suppressed, so that the input of the compressor 3 can be reduced and the COP can be improved.

As shown in FIG. 7, the refrigerant density in the second gas cooler 5 is larger than the refrigerant density in the first gas cooler 4. As described above, the larger the refrigerant density, the smaller the refrigerant pressure loss per unit channel length. Therefore, when other conditions are the same, the refrigerant pressure loss per length of the second refrigerant heat transfer tube 52 of the second gas cooler 5 becomes the refrigerant pressure loss per length of the first refrigerant heat transfer tube 42 of the first gas cooler 4. Smaller than that. Therefore, the inner diameter di2 of the second refrigerant heat transfer tube 52 of the second gas cooler 5 or the cross-sectional area of each second refrigerant heat transfer channel is equal to the inner diameter di1 of the first refrigerant heat transfer tube 42 of the first gas cooler 4 or the first of each. Even if it is smaller than the cross-sectional area of the refrigerant heat transfer channel, the refrigerant pressure loss of the second gas cooler 5 can be sufficiently suppressed. Further, by making the inner diameter di2 of the second refrigerant heat transfer tube 52 of the second gas cooler 5 or the cross-sectional area of each second refrigerant heat transfer channel relatively small, the second refrigerant heat transfer tube 52, that is, each of the second refrigerant transfer tubes. Since the refrigerant flow rate in the heat flow path is increased, the heat transfer coefficient of the refrigerant can be increased. As a result, the length of the second torsion pipe 51 of the second gas cooler, that is, the second liquid heat transfer channel can be shortened. From the above, the inner diameter di1 of the first refrigerant heat transfer tube 42 of the first gas cooler 4 or the cross-sectional area of each first refrigerant heat transfer channel is equal to the inner diameter di2 of the second refrigerant heat transfer tube 52 of the second gas cooler 5 or each. The cross-sectional area of the second refrigerant heat transfer channel is preferably large.

FIG. 16 shows the case where the twist pitch p of the first torsion tube 41 is equal to the torsion pitch p2 of the second torsion tube 51 and the inner diameter SRi of the first torsion tube 41 and the second torsion tube 51 is equal. It is a figure which shows the change of the heat transfer coefficient. The meaning of the horizontal axis in FIG. 16 is the same as the horizontal axis in FIG. In FIG. 16, the water-side heat transfer coefficient is expressed as a ratio to the water-side heat transfer coefficient at the water outlet of the first gas cooler 4. As shown in FIG. 16, the heat transfer coefficient on the water side decreases as the distance from the refrigerant inlet and the water outlet of the first gas cooler 4 increases, that is, as the water temperature decreases. Therefore, when the torsion pitch p of the first torsion tube 41 and the torsion pitch p2 of the second torsion tube 51 are equal and the inner diameter SRi of the first torsion tube 41 and the second torsion tube 51 is equal, the second gas cooler 5 The water-side heat transfer coefficient is smaller than the water-side heat transfer coefficient of the first gas cooler 4. In view of this point, in the second gas cooler 5, the contact area between the second refrigerant heat transfer tube 52 and the second torsion tube 51 can be increased by relatively reducing the torsion pitch p2 of the second torsion tube 51. desirable. Thereby, the length of the 2nd torsion pipe | tube 51 of the 2nd gas cooler 5 can be shortened. On the other hand, as described above, the torsion pitch p of the first torsion pipe 41 of the first gas cooler 4 is desirably relatively large. From the above, it is preferable that the torsion pitch p of the first torsion tube 41 of the first gas cooler 4 is larger than the torsion pitch p2 of the second torsion tube 51 of the second gas cooler 5.

In the first embodiment, it is preferable to make the inner diameter SRi of the first torsion tube 41 of the first gas cooler 4 equal to the inner diameter SRi of the second torsion tube 51 of the second gas cooler 5. When the second gas cooler 5 is installed near the first gas cooler 4, the upstream end of the first torsion pipe 41 and the downstream end of the second torsion pipe 51 are connected. In this case, by making the inner diameter SRi of the first torsion tube 41 equal to the inner diameter SRi of the second torsion tube 51, both can be easily connected. Further, by making the inner diameter SRi of the first torsion tube 41 equal to the inner diameter SRi of the second torsion tube 51, the material and the manufacturing method used for both can be made common, and the cost is reduced.

In the first embodiment, the number of first refrigerant heat transfer tubes 42 of the first gas cooler 4, that is, the number of first refrigerant heat transfer channels, and the number of second refrigerant heat transfer tubes 52 of the second gas cooler 5, that is, second refrigerant heat transfer flows. Preferably the number of paths is equal. By making the number of the first refrigerant heat transfer tubes 42 equal to the number of the second refrigerant heat transfer tubes 52, the forms of the first torsion tube 41 and the second torsion tube 51 can be designed similarly, and the cost is reduced. To do.

In the first embodiment, the case where the first heat exchanger (first gas cooler 4) and the second heat exchanger (second gas cooler 5) are constituted by a twisted tube heat exchanger has been described as an example. However, in the present invention, the first heat exchanger and the second heat exchanger are not limited to the twisted tube heat exchanger, and various types of heat exchangers can be used.

As described above, the inner diameter ratio di1 / di2 of the first refrigerant heat transfer tube 42 and the second refrigerant heat transfer tube 52 is preferably 1.1 or more and 1.4 or less, and preferably 1.2 or more and 1.4 or less. More preferred. When the inner diameter ratio di1 / di2 is 1.1, the ratio of the total cross-sectional area of the first refrigerant heat transfer channel of the first heat exchanger to the total cross-sectional area of the second refrigerant heat transfer channel of the second heat exchanger is: (1.1) ² ≈ 1.2. When the inner diameter ratio di1 / di2 is 1.2, the ratio of the total cross-sectional area of the first refrigerant heat transfer channel of the first heat exchanger to the total cross-sectional area of the second refrigerant heat transfer channel of the second heat exchanger is (1.2) ² ≈1.4. When the inner diameter ratio di1 / di2 is 1.4, the ratio of the total cross-sectional area of the first refrigerant heat transfer channel of the first heat exchanger to the total cross-sectional area of the second refrigerant heat transfer channel of the second heat exchanger is (1.4) ² ≈2. Therefore, when the numerical range of the inner diameter ratio di1 / di2 is replaced with the numerical range of the ratio of the channel cross-sectional area, the ratio of the total cross-sectional area of the first refrigerant heat transfer channel to the total cross-sectional area of the second refrigerant heat transfer channel is 1.2 or more and 2 or less, and preferably 1.4 or more and 2 or less. By setting the ratio of the channel cross-sectional areas in such a range, an effect similar to the above-described effect can be obtained.

In the first embodiment described above, the number of first refrigerant heat transfer channels of the first heat exchanger (first gas cooler 4) and the number of second refrigerant heat transfer channels of the second heat exchanger (second gas cooler 5). However, in the present invention, the number of first refrigerant heat transfer channels may be larger than the number of second refrigerant heat transfer channels. When the number of the first refrigerant heat transfer channels is increased as compared with the number of the second refrigerant heat transfer channels, the entire cross-sectional area of the first refrigerant heat transfer channel is set to the total of the second refrigerant heat transfer channels with a simple configuration. It can be larger than the cross-sectional area. When the number of first refrigerant heat transfer channels is larger than the number of second refrigerant heat transfer channels, for example, the cross-sectional area of the first refrigerant heat transfer channel is equal to the cross-sectional area of the second refrigerant heat transfer channel. May be. For this reason, the 1st refrigerant | coolant heat exchanger tube 42 of the 1st gas cooler 4 and the 2nd refrigerant | coolant heat exchanger tube 52 of the 2nd gas cooler 5 can be manufactured with a common material, and cost reduces.

In the first embodiment, the heat pump device that heats water using the first heat exchanger and the second heat exchanger has been described as an example. However, in the present invention, the first heat exchanger and the second heat exchanger are used. The liquid to be heated is not limited to water, and may be, for example, brine or antifreeze.

1 heat pump device, 1a water inlet, 1b water outlet, 2 tank unit, 2a hot water storage tank, 2b water pump, 2c hot water mixing valve, 3 compressor, 4 1st gas cooler, 5 2nd gas cooler, 6 expansion valve, 7 evaporation , 8 blower, 9 high / low

pressure heat exchanger

10, 11, 12, 17, 18, 19, 20, 21 pipe, 13 water supply pipe, 14 hot water outlet pipe, 15 water supply branch pipe, 16 hot water supply pipe, 23, 26 water flow Path, 31 sealed container, 32 compression element, 33 electric element, 34 first suction passage, 35 first discharge passage, 36 second suction passage, 37 second discharge passage, 38 internal space, 41 first torsion pipe, 42, 42a, 42b, 42c 1st refrigerant heat transfer tube, 50 control part, 51 2nd torsion tube, 52 2nd refrigerant heat transfer tube, 60 heat transfer material, 70 heat pump equipment , 71 compressor, 72 a gas cooler, 331 rotor, 332 stator, 411,411a, 411b, 411c, 511 grooves

Claims

The first discharge passage for discharging the refrigerant and the refrigeration oil and the second discharge passage for discharging the refrigerant and the refrigeration oil, and the mass flow rate of the refrigeration oil discharged from the first discharge passage is from the second discharge passage. More compressors than the mass flow rate of the refrigerating machine oil discharged,
One or a plurality of first refrigerant heat transfer channels through which the refrigerant discharged from the first discharge passage and the refrigerating machine oil pass, and one or a plurality of first liquid heat transfer channels through which the liquid passes, the first refrigerant transfer A first heat exchanger for exchanging heat between the heat flow path and the first liquid heat transfer flow path;
One or a plurality of second refrigerant heat transfer channels through which the refrigerant discharged from the second discharge passage and the refrigerating machine oil pass, and one or a plurality of second liquid heat transfer channels through which the liquid passes, the second refrigerant transfer A second heat exchanger for exchanging heat between the heat flow path and the second liquid heat transfer flow path;
With
A heat pump device in which a total cross-sectional area of the first refrigerant heat transfer channel is larger than a total cross-sectional area of the second refrigerant heat transfer channel.
The number of the first refrigerant heat transfer channels is equal to the number of the second refrigerant heat transfer channels,
The heat pump device according to claim 1, wherein a cross-sectional area of each of the first refrigerant heat transfer channels is larger than a cross-sectional area of each of the second refrigerant heat transfer channels.
The first refrigerant heat transfer channel is formed by a first refrigerant heat transfer tube,
The second refrigerant heat transfer channel is formed by a second refrigerant heat transfer tube,
The first liquid heat transfer channel is formed by a first torsion tube having a spiral groove on the outer periphery,
The second liquid heat transfer channel is formed by a second torsion tube having a spiral groove on the outer periphery,
The first refrigerant heat transfer tube is disposed along the groove of the first torsion tube,
The heat pump apparatus according to claim 1 or 2, wherein the second refrigerant heat transfer tube is disposed along the groove of the second torsion tube.
The heat pump device according to claim 3, wherein a ratio of an inner diameter of the first refrigerant heat transfer tube to an inner diameter of the second refrigerant heat transfer tube is 1.1 or more and 1.4 or less.
The heat pump device according to claim 3 or 4, wherein a torsion pitch of the first torsion tube is larger than a torsion pitch of the second torsion tube.
The value of p / SRi is 1.1 or more and 1.8 or less, where p is the torsion pitch of the first torsion tube and SRi is the inner diameter of the first torsion tube. The heat pump device according to claim 1.
The heat pump device according to any one of claims 3 to 6, wherein an inner diameter of the first torsion tube is equal to an inner diameter of the second torsion tube.
The heat pump device according to any one of claims 3 to 7, wherein the number of the first refrigerant heat transfer channels is larger than the number of the second refrigerant heat transfer channels.
The heat pump according to any one of claims 1 to 8, wherein a ratio of a total cross-sectional area of the first refrigerant heat transfer channel to a total cross-sectional area of the second refrigerant heat transfer channel is 1.2 or more and 2 or less. apparatus.
The ratio of the mass flow rate of the refrigerating machine oil to the sum of the mass flow rate of the refrigerant and the mass flow rate of the refrigerating machine oil in the first heat exchanger is 2% or more and 20% or less. Heat pump device.
The ratio of the mass flow rate of the refrigeration oil to the sum of the mass flow rate of the refrigerant and the mass flow rate of the refrigeration oil in the second heat exchanger is 0.01% or more and 1% or less. The heat pump device according to item.
The heat pump device according to any one of claims 1 to 11, wherein the liquid is water and has a function of supplying hot water obtained by heating the water.