JPWO2015136595A1 - Heat pump equipment - Google Patents

Abstract

An object of this invention is to provide the heat pump apparatus which can suppress that a liquid refrigerant accumulates inside the airtight container at the time of a low heating operation. The heat pump device of the present invention includes a first suction passage that guides the low-pressure refrigerant to the compression element in the sealed container, a first discharge passage that discharges the high-pressure refrigerant compressed by the compression element, and after being discharged from the first discharge passage. A compressor having a second suction passage that discharges the high-pressure refrigerant that has undergone heat exchange to the internal space of the sealed container without compression, and a second discharge passage that discharges the high-pressure refrigerant in the internal space of the sealed container; A first heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the passage, a second heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the second discharge passage, and a high heating operation And a control means for performing a low heating operation having a smaller heating capacity than that of the high heating operation, and the control means is configured so that the state of the refrigerant in the second suction passage becomes an overheated gas state during the low heating operation. Control.

Description

  The present invention relates to a heat pump device that heats a target fluid.

  Patent Document 1 below discloses a hot water supply cycle apparatus as follows. The hot water supply cycle apparatus includes a compressor, a gas cooler, an expansion valve, an evaporator, and the like. The compressor has a compression element and an electric element in a sealed container. The compressor includes a suction pipe that directly guides the low-pressure refrigerant to the compression element, a discharge pipe that discharges the high-pressure refrigerant compressed by the compression element to the outside of the sealed container without releasing it into the sealed container, and heat discharged from the discharge pipe. Refrigerant reintroduction pipe for reintroducing the refrigerant after replacement into the sealed container, and refrigerant redischarge pipe for discharging the refrigerant after being introduced into the sealed container from the refrigerant reintroduction pipe and passing through the electric element to the outside of the sealed container With. In the gas cooler, the temperature of water in the water pipe is increased by the refrigerant in the refrigerant pipe by exchanging heat between the water pipe through which hot water is circulated and the refrigerant pipe through which the compressed refrigerant is circulated. Of the refrigerant pipes, the high-temperature side refrigerant pipe connected to the discharge pipe exchanges heat with the outlet side of the water pipe of the gas cooler, and the low-temperature side refrigerant pipe connected to the refrigerant re-discharge pipe exchanges heat with the inlet side of the water pipe of the gas cooler. .

  The following Patent Document 2 discloses the following heat pump water heater. The heat pump water heater has a heat pump cycle in which a compressor, a water refrigerant heat exchanger, an expansion valve, and an evaporator are connected in an annular shape. The heat pump water heater has a hot water storage operation mode and a hot water operation mode. In the hot water filling operation mode, the heat pump cycle is operated, the water supplied from the hot water storage tank and the water flowing out of the water refrigerant heat exchanger of the heat pump cycle are mixed by the hot water mixing valve, and the mixed water is mixed. Supply to the bathtub and make the hot water temperature lower than the hot water storage operation mode.

Japanese Unexamined Patent Publication No. 2006-132427 Japanese Unexamined Patent Publication No. 2011-21828

  When the hot water operation mode as in Patent Document 2 is performed with the apparatus of Patent Document 1, the discharge pressure of the compressor becomes below the critical point, and the refrigerant condenses to a gas-liquid two-phase state in the high-temperature side refrigerant pipe, and is sealed. Gas-liquid two-phase refrigerant may flow into the container. The discharge pipe is attached to the top of the sealed container. When the gas-liquid two-phase refrigerant flows into the sealed container, gas-liquid separation is performed in the sealed container, and only the gas refrigerant flows out from the discharge pipe. As a result, liquid refrigerant is accumulated in the sealed container, and the entire refrigeration cycle becomes insufficient. Further, when the liquid refrigerant accumulated in the sealed container is heated by the heat of the compression element or the electric element, the refrigerant mixed with the refrigerating machine oil evaporates, so that the refrigerating machine oil is foamed and the refrigerating machine oil is combined with the refrigerant. It flows out from the discharge pipe. As a result, the refrigerating machine oil in the sealed container may be insufficient.

  The present invention has been made to solve the above-described problems, and an object of the present invention is to provide a heat pump device capable of suppressing liquid refrigerant from accumulating inside a sealed container during low heating operation.

  The heat pump device of the present invention includes a sealed container, a compression element provided inside the sealed container, and a first low-pressure refrigerant sucked from the outside of the sealed container that is led to the compression element without being discharged into the internal space of the sealed container. A suction passage, a first discharge passage that discharges high-pressure refrigerant compressed by the compression element to the outside of the sealed container without releasing it into the inner space of the sealed container, and a high pressure that exchanges heat after being discharged from the first discharge passage A compressor having a second suction passage that discharges the refrigerant to the inner space of the sealed container without compression, and a second discharge passage that discharges the high-pressure refrigerant in the inner space of the sealed container to the outside of the sealed container without being compressed. A first heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the first discharge passage; and a second heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the second discharge passage. , Expanded high-pressure refrigerant that passed through the second heat exchanger The total heating amount of the expansion part to be a low-pressure refrigerant, the evaporator that evaporates the low-pressure refrigerant that has passed through the expansion part, the high heating operation, and the first heat exchanger and the second heat exchanger is the high heating operation. And a control means for performing a low heating operation that is smaller than the control means, and the control means controls the refrigerant state in the second suction passage to be in an overheated gas state during the low heating operation.

  According to the heat pump device of the present invention, it is possible to suppress liquid refrigerant from accumulating inside the sealed container during low heating operation.

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 heat pump unit with which the heat pump apparatus of Embodiment 1 of this invention is provided. It is a pressure-enthalpy diagram of the high heating operation of the heat pump device of Embodiment 1 of the present invention. It is a figure showing an example of the temperature change of the refrigerant | coolant and water in a 1st heat exchanger and a 2nd heat exchanger in the high heating operation of the heat pump apparatus of Embodiment 1 of this invention. It is a flowchart which shows the control action of the control apparatus in the high heating operation of the heat pump apparatus of Embodiment 1 of this invention. It is a pressure-enthalpy diagram of the low heating operation of the heat pump device of Embodiment 1 of the present invention. It is a figure showing an example of the temperature change of the refrigerant | coolant and water in the 1st heat exchanger and the 2nd heat exchanger in the low heating operation of Embodiment 1 of this invention. It is a flowchart which shows the control action of the control apparatus in the low heating operation of the heat pump apparatus of Embodiment 1 of this invention. It is a flowchart which shows the control action of the control apparatus in the low heating operation of the heat pump apparatus of Embodiment 2 of this invention. It is a flowchart which shows the control action of the control apparatus in the low heating operation of the heat pump apparatus of Embodiment 3 of this invention. It is a block diagram which shows the heat pump unit with which the heat pump apparatus of Embodiment 4 of this invention is provided. It is a flowchart which shows the control action of the control apparatus in the low heating operation of the heat pump apparatus of Embodiment 4 of this invention.

  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 addition, the present invention includes all combinations of the embodiments described below.

Embodiment 1 FIG.
FIG. 1 is a configuration diagram illustrating a heat pump apparatus according to Embodiment 1 of the present invention. As shown in FIG. 1, the heat pump device 1 according to the first embodiment includes a heat pump unit 2 that heats water, a hot water storage tank 10, and a control device 50. The hot water storage tank 10 stores water by forming a temperature stratification in which the upper side is hot and the lower side is low. The lower part of the hot water storage tank 10 and the inlet 12 of the heat pump unit 2 are connected via an inlet pipe 11. A pump 13 is installed in the middle of the inlet pipe 11. One end of the upper pipe 14 is connected to the upper part of the hot water storage tank 10. The other end of the upper pipe 14 branches into two and is connected to the first inlet of the hot water mixing valve 15 and the first inlet of the bath mixing valve 16, respectively. The outlet 17 of the heat pump unit 2 is connected to a position in the middle of the upper pipe 14 via the outlet pipe 18. Details of the heat pump unit 2 will be described later. In the first embodiment, the case where the target fluid to be heated is water will be described. However, the target fluid in the present invention may be a fluid other than water, such as brine or antifreeze.

  A water supply pipe 19 for supplying water from a water source such as a water supply is connected to the lower part of the hot water storage tank 10. A pressure reducing valve 20 for reducing the water source pressure to a predetermined pressure is installed in the middle of the water supply pipe 19. As water flows in from the water supply pipe 19, the interior of the hot water storage tank 10 is always kept full. A water supply pipe 21 branches from a water supply pipe 19 between the hot water storage tank 10 and the pressure reducing valve 20. The downstream side of the water supply pipe 21 branches in two and is connected to the second inlet of the hot water mixing valve 15 and the second inlet of the bath mixing valve 16, respectively. The outlet of the hot water supply mixing valve 15 is connected to a hot water tap 23 via a hot water supply pipe 22. A hot water supply flow rate detection means 24 and a hot water supply temperature sensor 25 are installed in the hot water supply pipe 22. The outlet of the bath mixing valve 16 is connected to a bathtub 27 via a bath pipe 26. An open / close valve 28 and a bath temperature sensor 29 are installed in the bath pipe 26. A heat pump outlet temperature sensor 30 that detects a heat pump outlet temperature, which is a temperature of water exiting the heat pump unit 2, is installed in the outlet pipe 18 in the vicinity of the outlet 17 of the heat pump unit 2. The heat pump outlet temperature sensor 30 may be provided in a pipe (water channel 48 described later) inside the heat pump unit 2. In the following description, the temperature of water entering the heat pump unit 2 is referred to as “heat pump inlet temperature”.

  The control device 50 is a control means configured by, for example, a microcomputer. The control device 50 includes a ROM (Read Only Memory), a RAM (Random Access Memory), a storage unit including a nonvolatile memory, an arithmetic processing unit (CPU) that executes arithmetic processing based on a program stored in the storage unit, An input / output port for inputting / outputting external signals to / from the arithmetic processing unit, a timer for measuring time, and the like are provided. The control device 50 is electrically connected to various actuators and sensors included in the heat pump device 1. The control device 50 is connected to the operation unit 60 so as to communicate with each other. The user can set the hot water supply temperature, the amount of bath water, the bath temperature, etc., or can make a timer reservation for the bath hot water time by operating the operation unit 60. The control device 50 controls the operation of the heat pump device 1 by controlling the operation of each actuator according to the program stored in the storage unit based on the information detected by each sensor, the instruction information from the operation unit 60, and the like. .

  Next, hot water storage operation will be described. The hot water storage operation is an operation for increasing the amount of hot water stored and the amount of heat stored in the hot water storage tank 10. During the hot water storage operation, the control device 50 operates the heat pump unit 2 and the pump 13. In the hot water storage operation, the low temperature water led out from the lower part of the hot water storage tank 10 by the pump 13 is sent to the heat pump unit 2 through the inlet pipe 11 and heated by the heat pump unit 2 to become high temperature water. This high-temperature water flows into the upper part of the hot water storage tank 10 through the outlet pipe 18 and the upper pipe 14. By such hot water storage operation, hot water is accumulated in the hot water storage tank 10 from above.

  In the hot water storage operation, the control device 50 controls the heat pump outlet temperature detected by the heat pump outlet temperature sensor 30 to be, for example, about 65 ° C. to 90 ° C. The heat pump outlet temperature is lowered by controlling the pump 13 so that the flow rate of the water flowing through the heat pump unit 2 is increased. The heat pump outlet temperature rises by controlling the pump 13 so that the flow rate of the water flowing through the heat pump unit 2 becomes low. In the hot water storage operation, the control device 50 controls the heat pump unit 2 so as to perform the high heating operation. The high heating operation of the heat pump unit 2 is an operation for setting the heating capacity of the heat pump unit 2 to a predetermined rated capacity.

  Next, the hot water supply operation will be described. The hot water supply operation is an operation for supplying hot water to the hot water tap 23. When the user opens the hot-water tap 23, the water from the water supply pipe 19 flows into the lower part of the hot water storage tank 10 due to the water source pressure, so that the high-temperature water in the upper part of the hot water storage tank 10 flows out to the upper pipe 14. In the hot water supply mixing valve 15, the low temperature water supplied from the water supply pipe 21 and the high temperature water supplied from the hot water storage tank 10 through the upper pipe 14 are mixed. This mixed water is discharged to the outside from the hot water tap 23 through the hot water supply pipe 22. At this time, when the passage of the mixed water is detected by the hot water supply flow rate detection means 24, the control device 50 sets the hot water supply temperature detected by the hot water supply temperature sensor 25 in the operation section 60 in advance by the user. The mixing ratio of the hot water supply mixing valve 15 is controlled so as to be a value.

  Next, hot water filling operation will be described. The hot water filling operation is an operation in which hot water is accumulated in the bathtub 27. When the user performs the start operation of the hot water filling operation with the operation unit 60, or when the timer reserved time comes, the hot water filling operation is started. At the time of hot water filling operation, the control device 50 operates the heat pump unit 2 and the pump 13 to open the on-off valve 28. When water from the water supply pipe 19 flows into the lower part of the hot water storage tank 10 due to the water source pressure, the high-temperature water at the upper part of the hot water storage tank 10 flows out to the upper pipe 14. Further, the low temperature water led out from the lower part of the hot water storage tank 10 by the pump 13 is sent to the heat pump unit 2 through the inlet pipe 11 and heated by the heat pump unit 2. The water heated by the heat pump unit 2 passes through the outlet pipe 18 and flows into the upper pipe 14. The high temperature water supplied from the hot water storage tank 10 and the water heated by the heat pump unit 2 merge through the upper pipe 14 and are supplied to the bath mixing valve 16. In the bath mixing valve 16, low temperature water supplied from the water supply pipe 21 and hot water supplied through the upper pipe 14 are mixed. This mixed water passes through the bath pipe 26, passes through the on-off valve 28, and is discharged into the bathtub 27. At this time, the control device 50 controls the mixing ratio of the bath mixing valve 16 so that the hot water supply temperature detected by the bath temperature sensor 29 becomes the bathtub temperature set value set in advance by the user using the operation unit 60. .

  Thus, in the hot water filling operation of the first embodiment, hot water is supplied to the bathtub 27 by using not only high-temperature water stored in the hot water storage tank 10 but also water heated by the heat pump unit 2. In the hot water operation, the control device 50 controls the heat pump outlet temperature detected by the heat pump outlet temperature sensor 30 to be lower than the bathtub temperature set value. In the hot water filling operation, the control device 50 controls the heat pump unit 2 to perform the low heating operation. The low heating operation of the heat pump unit 2 is an operation that lowers the heating capacity of the heat pump unit 2 compared to the high heating operation.

  Here, the hot water filling ability will be described. The hot water filling capacity is heat energy per unit time required when a bathtub is filled at a target bath temperature under the conditions of a predetermined bath capacity, feed water temperature, and hot water flow rate. For example, assuming that the bath capacity is 180 L, the feed water temperature is 9 ° C., the target bath temperature is 45 ° C., and the hot water flow rate is 10 L / min to 20 L / min, the standard hot water filling capacity is 25 kW to 50 kW. On the other hand, the rated heating capacity of the heat pump unit 2 is, for example, about 4.5 kW to 9 kW. Only the heating capacity of the heat pump unit 2 cannot satisfy the standard hot water filling capacity. In order to satisfy the standard hot water filling capacity, it is necessary to use hot water stored in the hot water storage tank 10.

The heat pump unit 2 has a characteristic that COP (Coefficient of Performance) becomes higher as the heating capacity is lowered. Moreover, the heat pump unit 2 has a characteristic that COP increases as the heat pump outlet temperature decreases. For this reason, the COP for the low heating operation is higher than the COP for the high heating operation. The COP of the low heating operation is C1, the COP of the high heating operation is C2, the hot water COP that is a substantial COP of the hot water operation is C3, and the heating capacity of the low heating operation of the heat pump unit 2 relative to the hot water capacity When the ratio is Rhp, the hot water filling COP is expressed by the following equation.
C3 = C1 * Rhp + C2 * (1-Rhp) (1)

  As described above, a high heating operation is performed in the hot water storage operation. That is, the high temperature water in the hot water storage tank 10 is generated by a high heating operation. Accordingly, when hot water is supplied to the bathtub 27 using only high-temperature water stored in the hot water storage tank 10 without operating the heat pump unit 2 in the hot water operation, the hot water COP is equal to the COP of the high heating operation. . That is, if Rhp = 0 in the above equation (1), C3 = C2. On the other hand, according to the hot water filling operation of the first embodiment, the low water heating operation is performed and the water heated by the heat pump unit 2 is used as an auxiliary to make the hot water filling COP higher than the COP of the high heating operation. be able to. That is, since Rhp> 0 in the above equation (1), C3> C2. As described above, in the first embodiment, by performing the hot water filling operation accompanied by the low heating operation, it is possible to improve the hot water filling COP while ensuring the aforementioned hot water filling ability. Moreover, by operating the heat pump unit 2 during the hot water filling operation, a part of the thermal energy necessary for the hot water filling can be covered. For this reason, the amount of heat stored in the hot water storage tank 10 can be reduced. Therefore, the heat dissipation loss from the hot water storage tank 10 can be reduced, and the energy efficiency as a whole can be further improved. Moreover, since the heat storage amount to be secured in the hot water storage tank 10 can be reduced, the capacity of the hot water storage tank 10 can be reduced, and the hot water storage tank 10 can be downsized.

  FIG. 2 is a configuration diagram showing the heat pump unit 2 included in the heat pump device 1 according to the first embodiment of the present invention. As shown in FIG. 2, the heat pump unit 2 includes a refrigerant circuit in which a compressor 3, a first heat exchanger 4, a second heat exchanger 5, an expansion valve 6, and an evaporator 7 are connected by a refrigerant pipe. Is provided. The first heat exchanger 4 and the second heat exchanger 5 are heat exchangers that heat water with the heat of the refrigerant. The evaporator 7 is composed of an air refrigerant heat exchanger that performs heat exchange between air and refrigerant. The heat pump unit 2 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. In the first embodiment, carbon dioxide is used as the refrigerant. In addition, the evaporator 7 in this invention is not restricted to what heat-exchanges air and a refrigerant | coolant, For example, you may exchange a heat | fever with a refrigerant | coolant and groundwater, solar hot water, etc., for example.

  The compressor 3 includes a sealed container 31, a compression element 32 and an electric element 33 provided inside the sealed container 31, a first suction passage 34, a first discharge passage 35, a second suction passage 36, And two discharge passages 37. A compression element 32 is disposed below the electric element 33. Inside the sealed container 31, an internal space 38 between the compression element 32 and the electric element 33 and an internal space 39 above the electric element 33 are provided. The first suction passage 34 does not discharge the low-pressure refrigerant sucked into the compressor 3 into the internal spaces 38 and 39 of the sealed container 31, and guides the low-pressure refrigerant directly to the compression element 32. The compression element 32 compresses the low-pressure refrigerant into a high-pressure refrigerant. The compression element 32 is driven by the electric element 33. The electric element 33 is an electric motor having a stator 33a and a rotor 33b. The compression element 32 discharges the compressed high-pressure refrigerant to the first discharge passage 35. The first discharge passage 35 does not discharge the high-pressure refrigerant to the internal spaces 38 and 39 of the sealed container 31, but directly discharges the high-pressure refrigerant to the outside of the sealed container 31. The high-pressure refrigerant discharged from the first discharge passage 35 passes through the refrigerant flow path 40 and flows into the first heat exchanger 4. The high-pressure refrigerant cooled with water in the first heat exchanger 4 passes through the refrigerant passage 41 and the second suction passage 36 and is re-inhaled into the compressor 3.

  The outlet of the second suction passage 36 is located in the internal space 38 between the electric element 33 and the compression element 32. The second suction passage 36 does not compress the high-pressure refrigerant re-inhaled by the compressor 3 and discharges the high-pressure refrigerant to the internal space 38 between the electric element 33 and the compression element 32. The inlet of the second discharge passage 37 is located in the internal space 39 above the electric element 33. The high-pressure refrigerant discharged from the outlet of the second suction passage 36 to the internal space 38 between the electric element 33 and the compression element 32 passes through a gap between the rotor 33b and the stator 33a of the electric element 33 and the like. It reaches the internal space 39 above the element 33. At this time, the electric element 33 having a high temperature is cooled by the high-pressure refrigerant, and the high-pressure refrigerant is heated by the heat of the electric element 33. The second discharge passage 37 does not compress the high-pressure refrigerant in the internal space 39 above the electric element 33 and discharges the high-pressure refrigerant to the outside of the sealed container 31.

  The high-pressure refrigerant discharged from the second discharge passage 37 passes through the refrigerant flow path 42 and flows into the second heat exchanger 5. The high-pressure refrigerant cooled with water in the second heat exchanger 5 passes through the refrigerant flow path 43 and reaches the expansion valve 6. The expansion valve 6 is an expansion part that expands the high-pressure refrigerant to make it a low-pressure refrigerant. The low-pressure refrigerant expanded by the expansion valve 6 flows into the evaporator 7 through the refrigerant flow path 44. In the evaporator 7, the low-pressure refrigerant is heated and evaporated by exchanging heat with the outside air guided by the blower 8. The low-pressure refrigerant that has passed through the evaporator 7 reaches the first suction passage 34 of the compressor 3 through the refrigerant flow path 45 and is sucked into the compressor 3. The high-low pressure heat exchanger 9 exchanges heat between the high-pressure refrigerant in the refrigerant channel 43 and the low-pressure refrigerant in the refrigerant channel 45.

  In the following description, the pressure of the refrigerant discharged from the compression element 32 is referred to as “compression element discharge pressure”, and the pressure of the refrigerant sucked into the compression element 32 is referred to as “compression element suction pressure”. The refrigerant temperature is referred to as “compression element discharge temperature”, and the refrigerant temperature drawn into the compression element 32 is referred to as “compression element suction temperature”. The pressure of the high-pressure refrigerant discharged from the first discharge passage 35 is equal to the compression element discharge pressure. The pressure of the high-pressure refrigerant discharged from the first discharge passage 35 decreases due to the pressure loss that reaches the second suction passage 36 via the first heat exchanger 4. For this reason, the pressure of the high-pressure refrigerant in the internal space 38 of the sealed container 31 is slightly lower than the pressure of the high-pressure refrigerant discharged from the first discharge passage 35, that is, the compression element discharge pressure.

  The heat pump unit 2 includes a water flow path 46 that guides water flowing from the inlet 12 to the second heat exchanger 5, a water flow path 47 that guides water that has passed through the second heat exchanger 5 to the first heat exchanger 4, A water flow path 48 that guides the water that has passed through the heat exchanger 4 to the outlet 17 is further provided. During the heating operation, water flowing from the inlet 12 flows into the second heat exchanger 5 through the water flow path 46 and is heated by the heat of the refrigerant in the second heat exchanger 5. The water heated in the second heat exchanger 5 flows into the first heat exchanger 4 and is further heated by the heat of the refrigerant in the first heat exchanger 4. The water further heated in the first heat exchanger 4 reaches the outlet 17 through the water channel 48 and flows to the outlet pipe 18.

  The first discharge passage 35 or the refrigerant flow path 40 is provided with a discharge temperature sensor 51 that detects the compression element discharge temperature. A refrigerant temperature sensor 52 that detects the refrigerant temperature of the second suction passage 36 is provided in the second suction passage 36 or the refrigerant flow path 41.

  Next, the high heating operation during the hot water storage operation will be further described. As described above, in the high heating operation, the control device 50 controls the temperature of the water exiting the first heat exchanger 4, that is, the heat pump outlet temperature, to about 65 ° C. to 90 ° C., for example. FIG. 3 is a pressure-enthalpy diagram of the high heating operation. A to H in FIG. 3 correspond to A to H in FIG. As shown in FIG. 3, the refrigerant is compressed by the compression element 32 to a pressure exceeding the critical pressure (A → B). The high-pressure refrigerant in the supercritical state is cooled by the first heat exchanger 4 (B → C). The state of the high-pressure refrigerant sucked into the internal space 38 of the sealed container 31 from the second suction passage 36 is C in FIG. This high-pressure refrigerant is heated by the heat of the electric element 33 while reaching the internal space 39 (C → D). The state of the high-pressure refrigerant discharged from the second discharge passage 37 is D in FIG. This high-pressure refrigerant is cooled by the second heat exchanger 5 (D → E). Thereafter, the high-pressure refrigerant is further cooled by the high-low pressure heat exchanger 9 (E → F). The high-pressure refrigerant that has passed through the high-low pressure heat exchanger 9 is decompressed by the expansion valve 6 and becomes low-pressure refrigerant (F → G). This low-pressure refrigerant evaporates in the evaporator 7 (G → H). The low-pressure refrigerant evaporated in the evaporator 7 is heated in the high-low pressure heat exchanger 9 (H → A). The heat pump outlet temperature in such a high heating operation, which is 65 ° C. to 90 ° C., is sufficiently higher than the critical temperature of carbon dioxide, which is a refrigerant. In the high heating operation, the compression element discharge pressure and the pressure of the refrigerant inside the first heat exchanger 4, the sealed container 31, and the second heat exchanger 5 are pressures exceeding the critical pressure.

  FIG. 4 is a diagram illustrating an example of temperature changes of the refrigerant and water in the first heat exchanger 4 and the second heat exchanger 5 in the high heating operation. B to E in FIG. 4 correspond to B to E in FIGS. The horizontal axis of FIG. 4 represents the position in the 1st heat exchanger 4 and the 2nd heat exchanger 5 by ratio of flow path length. That is, the horizontal axis of FIG. 4 represents the ratio of the water flow path length from the water inlet of the second heat exchanger 5 when the total length of the water flow paths of the first heat exchanger 4 and the second heat exchanger 5 is 1. Or the ratio of the refrigerant | coolant flow path length from the refrigerant | coolant exit of the 2nd heat exchanger 5 when the full length of the refrigerant | coolant flow path of the 1st heat exchanger 4 and the 2nd heat exchanger 5 is set to 1 is represented. In the example shown in FIG. 4, the temperature of water entering the second heat exchanger 5, that is, the heat pump inlet temperature, is about 9 ° C., and the temperature of water exiting the first heat exchanger 4, ie, the heat pump outlet temperature, is about 65 ° C. . The temperature of the refrigerant entering the first heat exchanger 4, that is, the compression element discharge temperature is about 85 ° C.

  FIG. 5 is a flowchart showing the control operation of the control device 50 in the high heating operation. In the high heating operation, the control device 50 controls each actuator as follows. The control device 50 controls the compressor 3 so that the heating capacity of the heat pump unit 2 becomes the rated capacity (step S1). Here, the heating capacity is the total amount of water heating per hour for the first heat exchanger 4 and the second heat exchanger 5. The control device 50 can control the heating capacity by controlling the capacity of the compressor 3. The control device 50 can control the capacity of the compressor 3 by controlling the driving speed, driving frequency, and the like of the compressor 3. Moreover, the control apparatus 50 controls the water flow rate by the pump 13 so that the heat pump outlet temperature detected by the heat pump outlet temperature sensor 30 becomes a predetermined heating temperature set value in the range of 65 ° C. to 90 ° C. . Moreover, the control apparatus 50 controls the ventilation volume of the air blower 8 according to required evaporation capability. The evaporation capacity is the amount of heat absorbed by the refrigerant from the air in the evaporator 7.

  During the high heating operation, the control device 50 controls the refrigerant flow rate with the expansion valve 6 so that the compression element discharge temperature matches the target value. The smaller the opening of the expansion valve 6 and the lower the refrigerant flow rate, the higher the compression element discharge temperature. The compression element discharge temperature can be detected by a discharge temperature sensor 51 provided at B in FIGS. 2 and 6. The control device 50 stores a table that defines the relationship between parameters such as the heat pump outlet temperature, the outside air temperature, the heating capacity, and the target value of the compression element discharge temperature. The target value of the compression element discharge temperature is determined so as to obtain the maximum COP according to parameters such as the heat pump outlet temperature, the outside air temperature, and the heating capacity. The control device 50 determines a target value of the compression element discharge temperature based on parameters such as the heat pump outlet temperature, the outside air temperature, and the heating capacity, and the table. Then, the control device 50 determines whether or not the compression element discharge temperature matches the target value (step S2). If the compression element discharge temperature matches the target value in step S2, the control device 50 returns to step S1. If the compression element discharge temperature does not match the target value in step S2, the control device 50 controls the refrigerant flow rate with the expansion valve 6 so that the compression element discharge temperature matches the target value (step S3). . By controlling as described above, the COP during the high heating operation can be made sufficiently high.

  In the following description, the degree of superheat of the refrigerant coming out of the evaporator 7 is referred to as “evaporator outlet superheat degree”. There is a correlation between the discharge temperature of the compression element and the degree of superheat of the evaporator outlet. For this reason, at the time of the high heating operation, the control device 50 may control the refrigerant flow rate with the expansion valve 6 so that the evaporator outlet superheat degree matches the target value, instead of the steps S2 and S3. The degree of superheat at the outlet of the evaporator 7 increases as the opening degree of the expansion valve 6 is reduced to lower the refrigerant flow rate. The evaporator outlet superheat degree can be detected as, for example, a temperature difference between the two temperature sensors provided with temperature sensors G and H in FIGS. In this case, the control device 50 stores a table that defines the relationship between parameters such as the heat pump outlet temperature, the outside air temperature, and the heating capacity and the target value of the evaporator outlet superheat degree. The target value of the evaporator outlet superheat degree is determined so as to obtain the maximum COP according to parameters such as the heat pump outlet temperature, the outside air temperature, and the heating capacity. The control device 50 determines a target value of the evaporator outlet superheat degree based on parameters such as the heat pump outlet temperature, the outside air temperature, the heating capacity, and the above table. And it replaces with said step S2 and the control apparatus 50 judges whether the evaporator exit superheat degree corresponds with target value. If the evaporator outlet superheat degree matches the target value, the control device 50 returns to step S1. When the evaporator outlet superheat degree does not coincide with the target value, the control device 50 replaces the refrigerant flow rate with the expansion valve 6 so that the evaporator outlet superheat degree coincides with the target value instead of step S3. Control. By controlling as described above, the COP during the high heating operation can be made sufficiently high.

  Next, the low heating operation during the hot water operation will be further described. In the low heating operation, the control device 50 controls the heat pump outlet temperature to about 20 ° C. to 30 ° C., for example. FIG. 6 is a pressure-enthalpy diagram of the low heating operation. A to H in FIG. 6 correspond to A to H in FIG. As shown in FIG. 6, the refrigerant is compressed by the compression element 32 to a pressure equal to or lower than the critical pressure (A → B). The high-pressure refrigerant (B in FIG. 6) after being compressed by the compression element 32 is in a superheated gas state. The high-pressure refrigerant in the superheated gas state is cooled by the first heat exchanger 4 (B → C). The high-pressure refrigerant (C in FIG. 6) after being cooled by the first heat exchanger 4 is also in the superheated gas state. This high-pressure refrigerant is heated by the heat of the electric element 33 while reaching the internal space 39 (C → D). The state of the high-pressure refrigerant discharged from the second discharge passage 37 is D in FIG. This high-pressure refrigerant is condensed and liquefied by being cooled by the second heat exchanger 5 (D → E). Thereafter, the high-pressure refrigerant is further cooled by the high-low pressure heat exchanger 9 (E → F). The high-pressure refrigerant that has passed through the high-low pressure heat exchanger 9 is decompressed by the expansion valve 6 and becomes low-pressure refrigerant (F → G). This low-pressure refrigerant evaporates in the evaporator 7 (G → H). The low-pressure refrigerant evaporated in the evaporator 7 is heated in the high-low pressure heat exchanger 9 (H → A). 20 to 30 degreeC which is the heat pump exit temperature of such a low heating operation is lower than the critical temperature of the carbon dioxide which is a refrigerant | coolant. In the low heating operation, the compression element discharge pressure and the pressure of the refrigerant inside the first heat exchanger 4, the sealed container 31, and the second heat exchanger 5 become pressures below the critical pressure.

  FIG. 7 is a diagram illustrating an example of temperature changes of the refrigerant and water in the first heat exchanger 4 and the second heat exchanger 5 in the low heating operation. 7 correspond to B to E in FIGS. 2 and 6. The meaning of the horizontal axis in FIG. 7 is the same as the horizontal axis in FIG. In the example shown in FIG. 7, the temperature of water entering the second heat exchanger 5, that is, the heat pump inlet temperature is about 9 ° C., and the heat pump outlet temperature is about 25 ° C. The temperature of the refrigerant entering the first heat exchanger 4, that is, the compression element discharge temperature is about 45 ° C. The condensation saturation temperature of the refrigerant in the second heat exchanger 5 is about 22 ° C. In the following description, the condensation saturation temperature of the refrigerant in the second heat exchanger 5 is simply referred to as “condensation saturation temperature”. Further, the evaporation saturation temperature of the refrigerant in the evaporator 7 is simply referred to as “evaporation saturation temperature”.

  FIG. 8 is a flowchart showing the control operation of the control device 50 in the low heating operation. In the low heating operation, the control device 50 controls each actuator as follows. The control device 50 controls the compressor 3 so that the heating capacity of the heat pump unit 2 is lower than the heating capacity in the high heating operation, that is, lower than the rated capacity (step S11). In step S <b> 11, the control device 50 controls the capacity of the compressor 3 to be lower than that in the high heating operation. For example, the driving speed, driving frequency, etc. of the compressor 3 are lowered as compared with the high heating operation. The refrigerant flow rate in the low heating operation is lower than the refrigerant flow rate in the high heating operation. Moreover, the control apparatus 50 controls the water flow rate by the pump 13 so that the heat pump outlet temperature detected by the heat pump outlet temperature sensor 30 becomes a predetermined heating temperature set value in the range of 20 ° C. to 30 ° C. . The water flow rate in the low heating operation is higher than the water flow rate in the high heating operation. Moreover, the control apparatus 50 controls the ventilation volume of the air blower 8 according to required evaporation capability.

  In the low heating operation, the control device 50 sets the refrigerant flow rate to the expansion valve 6 so that the state of the refrigerant flowing into the internal space 38 of the sealed container 31, that is, the state of the refrigerant in the second suction passage 36 becomes a superheated gas state. To control. At this time, if the opening degree of the expansion valve 6 is decreased, the refrigerant flow rate is decreased, and the degree of superheat of the refrigerant in the second suction passage 36 is increased. The degree of superheat is the difference between the temperature of superheated gas (that is, superheated steam) and the temperature of saturated steam. If the degree of superheat is greater than zero, the refrigerant is in a superheated gas state.

  In the first embodiment, during the low heating operation, the control device 50 estimates the superheat degree SHsi of the refrigerant in the second suction passage 36 by a method described later (step S12). The control device 50 compares the estimated superheat degree SHsi with the reference value α (step S13). The reference value α is a predetermined value of zero or more. If the superheat degree SHsi is larger than the reference value α in step S13, it can be determined that the superheat degree SHsi is sufficiently large and the state of the refrigerant in the second suction passage 36 can be reliably maintained in the superheated gas state. In this case, the process returns to step S11. On the other hand, if the degree of superheat SHsi is equal to or less than the reference value α in step S13, it can be determined that the degree of superheat SHsi is not sufficient. In this case, the control device 50 increases the superheat degree SHsi by reducing the opening degree of the expansion valve 6 (step S14). Thereby, the state of the refrigerant in the second suction passage 36 can be reliably maintained in the superheated gas state.

Next, a method for estimating the superheat degree SHsi of the refrigerant in the second suction passage 36 during the low heating operation will be described. The degree of superheat SHsi of the refrigerant in the second suction passage 36 can be calculated from the condensation saturation temperature Tc and the refrigerant temperature Ts2 in the second suction passage 36 by the following equation.
SHsi = Ts2-Tc (2)

A method for estimating the condensation saturation temperature from the heat pump outlet temperature will be described with reference to FIG. As shown in FIG. 7, the refrigerant has a higher heat transfer coefficient in the gas-liquid two-phase region than the region where the refrigerant is superheated gas and the region where the refrigerant is supercooled liquid. Heat transfer is promoted. For this reason, most of the region where the water temperature rises is a gas-liquid two-phase region. In the low heating operation, the water temperature change from the heat pump inlet temperature to the heat pump outlet temperature is small compared to the high heating operation. In addition, the low heating operation has a higher water flow rate and a higher heat transfer rate than the high heating operation. As a result, the temperature difference between the refrigerant and water at the pinch point is smaller in the low heating operation than in the high heating operation. The pinch point is a point at which the refrigerant temperature and the water temperature are closest. In the superheated gas region, the gradient of the temperature change of the refrigerant with respect to the flow path direction is larger as it is closer to B (refrigerant inlet of the first heat exchanger 4) in FIG. Therefore, the inclination of the temperature change of the water with respect to the flow path direction is also small near the pinch point. Therefore, the water temperature Twp at the pinch point and the water temperature Twgc1i at the water inlet of the first heat exchanger 4 can be regarded as substantially equal. That is, the following equation holds.
Twp≈Twgc1i (3)

The heat exchange amount Qgc1 in the first heat exchanger 4 can be calculated by the temperature difference between the refrigerant inlet and outlet. That is, the following equation holds.
Qgc1 = Gr × Cpr × (Td1−Ts2) (4)
Where Gr is the refrigerant flow rate, Cpr is the constant pressure specific heat of the refrigerant, and Td1 is the compression element discharge temperature. The refrigerant flow rate Gr can be estimated from the capacity of the compressor 3 and the outside air temperature. The compression element discharge temperature Td1 can be detected by the discharge temperature sensor 51.

The heat exchange amount Qgc1 in the first heat exchanger 4 can also be calculated from the temperature difference between the refrigerant and water. That is, the following equation holds.
Qgc1 = Agc1 × Kgc1 × ΔTgc1 (5)
However, Agc1 is the heat transfer area of the first heat exchanger 4, Kgc1 is the heat passage rate of the first heat exchanger 4, and ΔTgc1 is the temperature difference between the refrigerant and water in the first heat exchanger 4. Agc1 is a fixed value. Kgc1 can be regarded as a substantially constant value. ΔTgc1 can be calculated by the following equation, assuming an arithmetic average temperature difference.
ΔTgc1 = (Td1 + Ts2) / 2− (Two + Twgc1i) / 2 (6)
However, Two is the heat pump outlet temperature. The arithmetic average temperature difference can be easily calculated by the above equation (6). In addition, a logarithmic average temperature difference may be used as ΔTgc1. In that case, ΔTgc1 can be calculated more accurately.

The first heat exchange is performed based on the compression element discharge temperature Td1, the refrigerant temperature Ts2 of the second suction passage 36, and the heat pump outlet temperature Two by combining the expressions (4), (5), and (6). The water temperature Twgc1i at the water inlet of the vessel 4 can be estimated. By assuming that Twgc1i thus estimated is equal to the water temperature Twp at the pinch point, the water temperature Twp at the pinch point can be obtained. The condensation saturation temperature Tc can be calculated by the following equation.
Tc = Twp + ΔTp (7)
However, ΔTp is the temperature difference between the refrigerant and water at the pinch point. The temperature difference ΔTp between the refrigerant and the water at the pinch point is about 1 ° C. to 3 ° C. By substituting the condensation saturation temperature Tc calculated by the above equation (7) into the above equation (2), the superheat degree SHsi of the refrigerant in the second suction passage 36 can be obtained. In step S12 of FIG. 8, as described above, the superheat degree SHsi of the refrigerant in the second suction passage 36 is set based on the compression element discharge temperature Td1, the refrigerant temperature Ts2 in the second suction passage 36, and the heat pump outlet temperature Two. Can be estimated.

The method for estimating the water temperature Twgc1i at the water inlet of the first heat exchanger 4 may be as follows instead of the above method. The heat exchange amount Qgc1 in the first heat exchanger 4 can be calculated by the temperature difference between the water inlet and outlet. That is, the following equation holds.
Qgc1 = Gw × Cpw × (Two−Twgc1i) (8)
However, Gw is a water flow rate and Cpw is the specific heat of water. The water flow rate Gw can be estimated from the driving speed of the pump 13. Alternatively, the water flow rate Gw may be detected by a flow rate sensor. Based on the compression element discharge temperature Td1, the refrigerant temperature Ts2 of the second suction passage 36, and the heat pump outlet temperature Two by combining any one of the formulas (4) and (5) with the formula (8). Thus, the water temperature Twgc1i at the water inlet of the first heat exchanger 4 can be estimated. By assuming that Twgc1i estimated in this way is equal to the water temperature Twp at the pinch point, the degree of superheat SHsi of the refrigerant in the second suction passage 36 can be estimated by using the above equations (7) and (2).

  In the first embodiment, the refrigerant temperature Ts2 in the second suction passage 36 can be detected by the refrigerant temperature sensor 52. However, instead of the refrigerant temperature sensor 52, the refrigerant temperature Ts2 in the second suction passage 36 may be detected by a temperature sensor provided on the outer surface of the sealed container 31 or the like.

Further, without providing the refrigerant temperature sensor 52, the refrigerant temperature Ts2 of the second suction passage 36 may be estimated as follows based on the compression element discharge temperature Td1 and the heat pump outlet temperature Two. The heat exchange amount Qgc1 in the first heat exchanger 4 can be regarded as being proportional to the temperature difference between the compression element discharge temperature Td1 and the heat pump outlet temperature Two. Therefore, if the proportionality coefficient is F, the following equation is established.
Qgc1 = F × (Td1−Two) (9)
By combining the above equations (4) and (9), the refrigerant temperature Ts2 of the second suction passage 36 can be estimated based on the compression element discharge temperature Td1 and the heat pump outlet temperature Two. Moreover, the water temperature Twgc1i at the water inlet of the first heat exchanger 4 can be estimated based on the compression element discharge temperature Td1 and the heat pump outlet temperature Two by combining the above equations (8) and (9). By assuming that Twgc1i estimated in this way is equal to the water temperature Twp at the pinch point, the degree of superheat SHsi of the refrigerant in the second suction passage 36 can be estimated by using the above equations (7) and (2).

  The theory described above assumes that the refrigerant in the second suction passage 36 is in a superheated gas state when calculating the heat exchange amount Qgc1 in the first heat exchanger 4. For this reason, assuming that the refrigerant in the second suction passage 36 is in a gas-liquid two-phase state, the heat exchange amount Qgc1 in the first heat exchanger 4 is estimated to be smaller than actual. Estimating the amount of heat exchange Qgc1 in the first heat exchanger 4 to be smaller than actual leads to estimating the temperature difference ΔTgc1 of refrigerant and water in the first heat exchanger 4 to be smaller than actual. Estimating the temperature difference ΔTgc1 between the refrigerant and water in the first heat exchanger 4 to be smaller than actual leads to estimating the water temperature Twgc1i at the water inlet of the first heat exchanger 4 higher than actual. Estimating the water temperature Twgc1i at the water inlet of the first heat exchanger 4 higher than actual leads to estimating the condensation saturation temperature Tc higher than actual. The superheat degree SHsi of the refrigerant in the second suction passage 36 is calculated by the above equation (2). For this reason, estimating the condensation saturation temperature Tc higher than actual leads to estimating the superheat degree SHsi of the refrigerant in the second suction passage 36 smaller than actual. Therefore, if the superheat degree SHsi of the refrigerant in the second suction passage 36 estimated by the above-described theory is controlled to be greater than zero, the state of the refrigerant in the second suction passage 36 can be reliably controlled to the superheated gas state.

In step S12 of FIG. 8, instead of the method described above, the degree of superheat of the refrigerant in the second suction passage 36 based on the evaporation saturation temperature Te, the compression element suction temperature Ts1, and the compression element discharge temperature Td1 as follows. SHsi may be estimated. When the compression process in the compression element 32 is a polytropic change and the polytropic index is n, the following equation holds. The polytropic index n can be regarded as a constant value determined from the physical property value of the refrigerant and the compressor efficiency.
Td1 = Ts1 × (Pd1 / Ps1) ^{(n−1) / n} (10)
However, Pd1 is a compression element discharge pressure. The compression element suction temperature Ts1 can be detected by providing a temperature sensor at A in FIGS. The compression element discharge temperature Td1 can be detected by a discharge temperature sensor 51 provided at B in FIGS.

  The evaporation saturation temperature Te can be detected by providing a temperature sensor at G in FIGS. Based on the evaporation saturation temperature Te, the evaporation saturation pressure Pe can be calculated from the relationship between the saturation temperature and the saturation pressure. If the pressure loss in the evaporator 7 is ignored, the compression element suction pressure Ps1 can be regarded as being equal to the evaporation saturation pressure Pe. Alternatively, the compression element suction pressure Ps1 may be calculated by subtracting a constant value of pressure loss from the evaporation saturation pressure Pe. The compression element discharge pressure Pd1 is calculated based on the evaporation saturation temperature Te, the compression element suction temperature Ts1, and the compression element discharge temperature Td1 by substituting the compression element suction pressure Ps1 thus determined in the above equation (10). it can. Based on the compression element discharge pressure Pd1, the condensation saturation temperature Tc can be calculated from the relationship between the saturation temperature and the saturation pressure. By substituting the condensation saturation temperature Tc calculated in this way into the above equation (2), the superheat degree SHsi of the refrigerant in the second suction passage 36 can be estimated.

According to the heat pump device 1 of the first embodiment, the following effects can be obtained.
(1) During hot water storage operation, that is, during high heating operation, the compression element discharge pressure Pd1 becomes supercritical pressure, and the refrigerant in the first heat exchanger 4 and the second heat exchanger 5 is not condensed. For this reason, the refrigerant | coolant flow volume can be controlled with the expansion valve 6 so that COP may become the maximum.
(2) During the hot water filling operation, that is, the low heating operation, the gas-liquid two-phase refrigerant flows into the internal space 38 of the hermetic container 31 by controlling the refrigerant state of the second suction passage 36 to be the superheated gas state. Can be surely prevented. Thereby, it is possible to appropriately control the refrigerant flow rate while improving the reliability. As a result, efficient operation is possible.
(3) The superheat degree SHsi of the refrigerant in the second suction passage 36 can be estimated based on the detected temperatures of temperature sensors such as the heat pump outlet temperature sensor 30, the discharge temperature sensor 51, and the refrigerant temperature sensor 52. For this reason, it can control so that the state of the refrigerant | coolant of the 2nd suction passage 36 at the time of low heating operation may be in a superheated gas state, without using an expensive pressure sensor.
(4) Since the gas-liquid two-phase refrigerant can be reliably prevented from flowing into the sealed container 31 even during the low heating operation, the liquid refrigerant is not accumulated in the sealed container 31. When the liquid refrigerant accumulated in the hermetic container 31 is heated by the compression element 32 or the electric element 33, the liquid refrigerant mixed with the refrigerating machine oil is vaporized, so that the refrigerating machine oil is foamed. When the refrigerating machine oil is foamed, the refrigerating machine oil is mixed with the gas refrigerant, and the refrigerating machine oil flows out of the second discharge passage 37 along with the gas refrigerant. As a result, the refrigerating machine oil in the hermetic container 31 is insufficient, resulting in poor lubrication of the sliding portion of the compressor 3. Moreover, refrigeration oil stagnates in the 2nd heat exchanger 5, the heat transfer of a refrigerant | coolant is inhibited, and performance falls. On the other hand, according to the first embodiment, the liquid refrigerant is not accumulated in the sealed container 31, and these problems can be reliably prevented.

  The present invention is not limited to the configuration described above, and may be as follows, for example. A pressure sensor that detects the compression element discharge pressure Pd1 may be provided, and the degree of superheat of the refrigerant in the second suction passage 36 may be controlled based on the value of the pressure sensor. A temperature sensor may be provided in the water flow path 47 connecting the first heat exchanger 4 and the second heat exchanger 5, and the water temperature Twgc1i at the water inlet of the first heat exchanger 4 may be detected by the temperature sensor. A temperature sensor may be provided at an intermediate point of the refrigerant flow path of the second heat exchanger 5, and the condensation saturation temperature Tc may be directly detected by the temperature sensor.

  The effect of the present invention is particularly prominent when a refrigerant (for example, carbon dioxide) having a pressure at which the compression element discharge pressure during the high heating operation exceeds the critical pressure is used as in the first embodiment. However, the present invention can also be applied to the case of using a refrigerant whose compression element discharge pressure during high heating operation is a pressure equal to or lower than the critical pressure. Examples of such a refrigerant include R410A, R32, R22, R407C, propane, propylene, HFO-1234yf, HFO-1234ze, or a mixed refrigerant thereof.

  A general design method in the case of using a refrigerant whose compression element discharge pressure at the time of high heating operation is a pressure equal to or lower than the critical pressure is that the second suction passage 36 is adjusted according to the conditions at the time of high heating operation that operates at the rated capacity. In this method, the ratio of the size (heat exchange amount) of the first heat exchanger 4 and the second heat exchanger 5 is designed so that the refrigerant becomes superheated gas. However, even in such a design, at the time of low heating operation, the heat pump outlet temperature is low, the compression element discharge pressure is lowered, and the enthalpy difference between the first heat exchanger 4 and the second heat exchanger 5 is reduced. The refrigerant in the second suction passage 36 is likely to be in a gas-liquid two-phase state. On the other hand, the same effects as described above can be obtained by applying the present invention. Therefore, it is meaningful to apply the present invention even when a refrigerant is used in which the discharge pressure of the compression element during high heating operation is a critical pressure or less.

Embodiment 2. FIG.
Next, a second embodiment of the present invention will be described with reference to FIG. 9. The description will focus on the differences from the first embodiment described above, and the same or corresponding parts will be denoted by the same reference numerals. Is omitted. FIG. 9 is a flowchart showing the control operation of the control device 50 in the low heating operation of the heat pump device 1 according to the second embodiment of the present invention. Since step S11 to step S13 in FIG. In the second embodiment, when the superheat degree SHsi of the refrigerant in the second suction passage 36 is equal to or less than the reference value α in step S13, the control device 50 causes the compressor 3 to increase the capacity of the compressor 3. Is controlled (for example, by increasing the driving speed of the compressor 3), the superheat degree SHsi of the refrigerant in the second suction passage 36 is increased (step S15). Thereby, in the second embodiment, the state of the refrigerant in the second suction passage 36 can be more reliably maintained in the superheated gas state. Since the second embodiment is the same as the first embodiment except for the matters described above, further description is omitted.

Embodiment 3 FIG.
Next, a third embodiment of the present invention will be described with reference to FIG. 10. The description will focus on the differences from the first embodiment described above, and the same or corresponding parts will be denoted by the same reference numerals. Is omitted. FIG. 10 is a flowchart showing the control operation of the control device 50 in the low heating operation of the heat pump device 1 according to the third embodiment of the present invention. Since step S11 to step S13 of FIG. In the third embodiment, when the superheat degree SHsi of the refrigerant in the second suction passage 36 is equal to or less than the reference value α in step S13, the control device 50 controls the pump 13 so that the water flow rate becomes low. Thus, the superheat degree SHsi of the refrigerant in the second suction passage 36 is increased (step S16). When the water flow rate is lowered, the compression element discharge pressure increases. However, since the saturated vapor line in the pressure-enthalpy diagram is inclined to the upper left, the enthalpy of the saturated gas becomes smaller as the pressure increases. As a result, the degree of superheat SHsi increases. According to the third embodiment, the state of the refrigerant in the second suction passage 36 can be more reliably maintained in the superheated gas state. Since the third embodiment is the same as the first embodiment except for the matters described above, further description is omitted.

Embodiment 4 FIG.
Next, the fourth embodiment of the present invention will be described with reference to FIG. 11 and FIG. 12. The description will focus on the differences from the first embodiment described above, and the same or corresponding parts will be denoted by the same reference numerals. The description is omitted. FIG. 11 is a configuration diagram showing the heat pump unit 2 included in the heat pump device 1 according to the fourth embodiment of the present invention. As shown in FIG. 11, the heat pump unit 2 according to the fourth embodiment has a configuration similar to that of the first embodiment, and a second suction passage from the first discharge passage 35 without passing through the first heat exchanger 4. Further, a bypass passage 55 for flowing the refrigerant to 36 and a bypass valve 56 provided in the bypass passage 55 are further provided. The bypass channel 55 is a channel that bypasses the refrigerant channel of the first heat exchanger 4. The bypass valve 56 is a flow path control element that makes the amount of refrigerant passing through the bypass flow path 55 variable. When the bypass valve 56 is closed, all of the superheated gas refrigerant discharged from the first discharge passage 35 passes through the first heat exchanger 4. That is, the state is the same as in the first embodiment. On the other hand, when the bypass valve 56 is opened, the superheated refrigerant discharged from the first discharge passage 35 flows separately into the refrigerant flow path of the first heat exchanger 4 and the bypass flow path 55. . Then, the refrigerant that has passed through the refrigerant flow path of the first heat exchanger 4 and the refrigerant that has passed through the bypass flow path 55 merge and flow to the second suction passage 36.

  FIG. 12 is a flowchart showing the control operation of control device 50 in the low heating operation of heat pump device 1 according to the fourth embodiment of the present invention. Since step S11 to step S13 in FIG. In the present fourth embodiment, when the superheat degree SHsi of the refrigerant in the second suction passage 36 is equal to or less than the reference value α in step S13, the control device 50 causes the bypass valve 56 to increase the opening degree of the bypass valve 56. By controlling 56, the superheat degree SHsi of the refrigerant in the second suction passage 36 is increased (step S17). That is, in step S17, when the bypass valve 56 is closed, the bypass valve 56 is opened to a predetermined opening, and when the bypass valve 56 is already open, the opening of the bypass valve 56 is increased. When the opening degree of the bypass valve 56 is increased, the ratio of the refrigerant passing through the bypass passage 55 increases. Since the refrigerant in the superheated gas state that has passed through the bypass flow path 55 has not exchanged heat, it is at a higher temperature than the refrigerant that has passed through the refrigerant flow path of the first heat exchanger 4. Therefore, the degree of superheat SHsi of the refrigerant in the second suction passage 36 is increased by increasing the opening degree of the bypass valve 56 and increasing the ratio of the refrigerant passing through the bypass flow path 55. With such a configuration, in the fourth embodiment, the state of the refrigerant in the second suction passage 36 can be more reliably maintained in the superheated gas state. Since the fourth embodiment is the same as the first embodiment except for the matters described above, further description is omitted.

DESCRIPTION OF SYMBOLS 1 Heat pump apparatus, 3 Compressor, 4 1st heat exchanger, 5 2nd heat exchanger, 6 Expansion valve, 7 Evaporator, 8 Blower, 9 Feed water temperature, 9 High / low pressure heat exchanger, 10 Hot water tank, 11 Inlet Piping, 12 inlet, 13 pump, 14 upper piping, 15 hot water mixing valve, 16 bath mixing valve, 17 outlet, 18 outlet piping, 19 water supply piping, 20 pressure reducing valve, 21 water supply piping, 22 hot water piping, 23 hot water tap, 24 Hot water flow rate detection means, 25 Hot water temperature sensor, 26 Bath piping, 27 Bathtub, 28 On-off valve, 29 Bath temperature sensor, 30 Heat pump outlet temperature sensor, 31 Airtight container, 32 Compression element, 33 Electric element, 33a Stator, 33b Rotation 34, first suction passage, 35 first discharge passage, 36 second suction passage, 37 second discharge passage, 38, 39 internal space, 40, 41 42, 43, 44 and 45 refrigerant passages, 46, 47, 48 water passage, 50 control unit, 51 a discharge temperature sensor, 52 a refrigerant temperature sensor, 55 bypass passage, 56 bypass valve, 60 an operation unit

The heat pump device of the present invention includes a sealed container, a compression element provided inside the sealed container, and a first low-pressure refrigerant sucked from the outside of the sealed container that is led to the compression element without being discharged into the internal space of the sealed container. A suction passage, a first discharge passage that discharges high-pressure refrigerant compressed by the compression element to the outside of the sealed container without releasing it into the inner space of the sealed container, and a high pressure that exchanges heat after being discharged from the first discharge passage A compressor having a second suction passage that discharges the refrigerant to the inner space of the sealed container without compression, and a second discharge passage that discharges the high-pressure refrigerant in the inner space of the sealed container to the outside of the sealed container without being compressed. A first heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the first discharge passage; and a second heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the second discharge passage. , Expanded high-pressure refrigerant that passed through the second heat exchanger The total heating amount of the expansion part to be a low-pressure refrigerant, the evaporator that evaporates the low-pressure refrigerant that has passed through the expansion part, the high heating operation, and the first heat exchanger and the second heat exchanger is the high heating operation. Control means for performing a low heating operation that is smaller than the control means, and the control means controls the refrigerant state in the second suction passage to be in a superheated gas state during the low heating operation, and during the high heating operation. The pressure of the refrigerant discharged from the compression element exceeds the critical pressure, and the temperature of the target fluid exiting the first heat exchanger exceeds the critical temperature of the refrigerant. During low heating operation, the refrigerant is discharged from the compression element. that the temperature of the target fluid exiting the first heat exchanger with the refrigerant pressure becomes critical pressure below the pressure is shall such to a temperature lower than the critical temperature.

Claims (7)

A hermetic container, a compression element provided inside the hermetic container, and a first suction passage for guiding low-pressure refrigerant sucked from the outside of the hermetic container to the compression element without releasing the refrigerant into the inner space of the hermetic container The high pressure refrigerant compressed by the compression element is discharged to the outside of the sealed container without being discharged into the inner space of the sealed container, and heat exchange is performed after being discharged from the first discharge path. A second suction passage for discharging the high-pressure refrigerant into the inner space of the sealed container without compressing, and a second discharge passage for discharging the high-pressure refrigerant in the inner space of the sealed container to the outside of the sealed container without compressing. A compressor having,
A first heat exchanger that heats the target fluid with the heat of the high-pressure refrigerant discharged from the first discharge passage;
A second heat exchanger for heating the target fluid with the heat of the high-pressure refrigerant discharged from the second discharge passage;
An expansion section that expands the high-pressure refrigerant that has passed through the second heat exchanger into a low-pressure refrigerant;
An evaporator for evaporating the low-pressure refrigerant that has passed through the expansion section;
Control means for performing a high heating operation and a low heating operation in which the total heating amount of the first heat exchanger and the second heat exchanger is smaller than the high heating operation;
With
The said control means is a heat pump apparatus which controls so that the state of the refrigerant | coolant of said 2nd suction passage may be in a superheated gas state at the time of the said low heating driving | operation.   The pressure of the refrigerant discharged from the compression element becomes higher than the critical pressure during the high heating operation, and the pressure of the refrigerant discharged from the compression element becomes lower than the critical pressure during the low heating operation. Item 2. The heat pump device according to Item 1.   The said control means controls the refrigerant | coolant flow volume by the said expansion part so that the temperature of the refrigerant | coolant discharged from the said compression element or the superheat degree of the refrigerant | coolant which exits from the said evaporator becomes a target value at the time of the said high heating operation. The heat pump device according to claim 1 or claim 2.   The control means, during the low heating operation, to the temperature of the refrigerant discharged from the compression element, the temperature of the refrigerant in the second suction passage, and the temperature of the target fluid exiting from the first heat exchanger On the basis of the degree of superheat of the refrigerant in the second suction passage, and based on the estimation result, at least one of the opening of the expansion section, the capacity of the compressor, and the flow rate of the target fluid is calculated. The heat pump device according to any one of claims 1 to 3, which is controlled.   The control means, during the low heating operation, based on the temperature of the refrigerant sucked into the compression element, the temperature of the refrigerant discharged from the compression element, and the evaporation saturation temperature of the evaporator. 2. The degree of superheat of the refrigerant in the two suction passages is estimated, and based on the estimation result, at least one of the opening of the expansion section, the capacity of the compressor, and the flow rate of the target fluid is controlled. The heat pump device according to claim 3. A bypass passage for flowing a refrigerant from the first discharge passage to the second suction passage without going through the first heat exchanger;
A flow path control element for varying the amount of refrigerant passing through the bypass flow path;
With
The said control means controls the quantity of the refrigerant | coolant which passes along the said bypass flow path by the said flow-path control element so that the state of the refrigerant | coolant of a said 2nd suction passage may be in a superheated gas state at the time of the said low heating operation. The heat pump device according to any one of claims 1 to 3. Equipped with a hot water storage tank,
The control means performs the high heating operation during hot water storage operation in which water heated by the first heat exchanger and the second heat exchanger flows into the hot water storage tank, and the first heat exchanger and the The heat pump device according to any one of claims 1 to 6, wherein the low heating operation is performed during a hot water supply operation in which water heated by the second heat exchanger is supplied to a bathtub.

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