JP2017150705A - Heat pump device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a heat pump device capable of preventing dew condensation of a housing of a heat exchange terminal caused by making circulation liquid into a cryogenic temperature, in a case of performing return temperature control.SOLUTION: Temperatures of evaporator inlet refrigerant flowing into first and second heat exchangers 41, 51 are detected, rotational frequencies of first and second compressor 43, 53 and openings of first and second expansion valves 44, 54 are detected, and in accordance with theses results, upper limit vales of the rotational frequencies of the first and second compressor 43, 53 are set. Consequently, since a heat absorption amount from a terminal circulation circuit 30 with first and second heat pump circuits 40, 50 can be suppressed, outgoing cooling water temperature is prevented from being extremely low, and occurrence of dew condensation in a housing of a heat exchange terminal 36 can be prevented.SELECTED DRAWING: Figure 3

Description

  The present invention relates to a heat pump device capable of supplying a circulating liquid obtained by performing heat exchange to a refrigerant in a heat pump circuit after heat exchange with a heat source side to a load terminal and performing cooling.

  Conventionally, in this type of heat pump apparatus, as described in Patent Document 1, the temperature of the circulating fluid that returns from the load terminal to the load-side heat exchanger is detected in the load-side circuit, and the return circulating fluid temperature is set in advance. In some cases, the rotation speed of the compressor of the heat pump circuit is controlled (return temperature control) so that the target return circulating fluid temperature is obtained.

JP2015-129616A

  In this conventional apparatus, by performing such return temperature control, when the cooling load of the load circuit is small, it is possible to supply a relatively high-temperature circulating fluid to the load terminal, thereby improving efficiency. On the other hand, when the cooling load of the load circuit is large, a relatively low temperature circulating fluid is supplied to the load terminal. However, because the return temperature control is performed, the temperature of the circulating fluid toward the load terminal does not increase. Depending on the operating conditions, the circulating fluid may become extremely cold. At this time, when the low-temperature circulating fluid is supplied to the load terminal, the case of the load terminal may be cooled to cause condensation on the surface of the load terminal.

  In order to solve the above-mentioned problems, in claim 1 of the present invention, a compressor, a load-side heat exchanger as an evaporator, a decompression means, and a heat-source-side heat exchanger as a condenser capable of exchanging heat with the heat-source side Are connected by refrigerant piping to form a heat pump circuit, and the load-side heat exchanger and at least one load terminal are connected by circulating fluid piping to form a load-side circuit, and in the load circuit, from the load terminal Returning to the load-side heat exchanger, circulating fluid temperature detecting means for detecting the circulating fluid temperature, and the compressor so that the circulating fluid temperature detected by the circulating fluid temperature detecting means becomes the target circulating fluid temperature. A compressor control means for controlling the number of revolutions, wherein in the heat pump circuit, a refrigerant temperature detection means for detecting a refrigerant temperature flowing into the load-side heat exchanger, and the heat pump circuit Corresponding to the state quantity detection means for detecting the refrigerant state quantity related to the refrigerant circulation amount in the refrigerant, the refrigerant temperature detected by the refrigerant temperature detection means and the refrigerant state quantity detected by the state quantity detection means, Upper limit value setting means for setting an upper limit value of the rotation speed of the compressor in the control of the compressor control means is provided.

  According to a second aspect of the present invention, supply from the load-side heat exchanger to the load terminal is performed based on the refrigerant temperature detected by the refrigerant temperature detection means and the refrigerant state quantity detected by the state quantity detection means. Estimated temperature calculating means for calculating the estimated temperature of the circulating fluid, and the upper limit value setting means controls the compressor control means according to the estimated temperature calculated by the estimated temperature calculating means. The upper limit value of the rotation speed of the compressor is set.

  According to a third aspect of the present invention, the upper limit value setting means sets the upper limit value higher as the estimated temperature is higher, and sets the upper limit value lower as the estimated temperature is lower.

  According to a fourth aspect of the present invention, the state quantity detection means detects the rotation speed of the compressor and the valve opening of the pressure reduction means as the refrigerant state quantity, and the estimated temperature calculation means is the refrigerant temperature. The estimated temperature is calculated based on the refrigerant temperature detected by the detection means and the rotation speed and the valve opening detected by the state quantity detection means.

  According to a fifth aspect of the present invention, the estimated temperature calculating means uses the estimated temperature by using the refrigerant temperature, a rotational speed coefficient corresponding to the rotational speed, and a valve opening coefficient corresponding to the valve opening. Is calculated.

  According to a sixth aspect of the present invention, the estimated temperature calculating means calculates the estimated temperature based on a sum of the refrigerant temperature detected by the refrigerant temperature detecting means, the rotation speed coefficient, and the valve opening coefficient.

  According to a seventh aspect of the present invention, in the heat pump circuit, a discharge temperature detecting means for detecting a refrigerant temperature discharged from the compressor, and the pressure reducing means according to the refrigerant temperature detected by the discharge temperature detecting means. Pressure reduction control means for controlling the valve opening degree.

  According to the first aspect of the present invention, for example, at the time of cooling operation, the refrigerant discharged from the compressor in the heat pump circuit and having a high temperature and high pressure exchanges heat in the heat source side heat exchanger, and the refrigerant condenses while dissipating heat to the heat source side. It becomes a high-pressure liquid. After the condensation, the refrigerant, which has been decompressed by the decompression means in the heat pump circuit and becomes low temperature and low pressure, exchanges heat with the circulating fluid circulated in the load side circuit and the load side heat exchanger, absorbs heat from the circulating fluid and evaporates. . As described above, the circulating fluid in the load-side circuit absorbs heat in the load-side heat exchanger and is cooled, and then is supplied to the load terminal to cool the air-conditioned space.

  At this time, the load side circuit is provided with circulating fluid temperature detecting means for detecting the circulating fluid temperature returning from the load terminal to the load side heat exchanger. The rotation speed of the compressor of the heat pump circuit corresponds to the instruction value of the operator, for example, according to the circulating fluid temperature detected by the circulating fluid temperature detecting means by the compressor control means. The target circulating fluid temperature is controlled.

  By performing temperature control on the circulating fluid returning from the load terminal (so-called return temperature control), relatively high-temperature circulating fluid can be supplied to the load terminal when the cooling load of the load circuit is small, improving efficiency. Can be achieved. On the other hand, when the cooling load of the load circuit is large, a relatively low temperature circulating fluid is supplied to the load terminal. At that time, depending on the operating conditions, the circulating fluid becomes extremely low temperature, and there is a possibility that the housing may condense at the load terminal to which the low-temperature circulating fluid is supplied.

  Therefore, according to the first aspect, the refrigerant temperature flowing into the load-side heat exchanger is detected by the refrigerant temperature detecting means in the heat pump circuit, and the refrigerant state quantity related to the refrigerant circulation amount in the heat pump circuit is detected by the state quantity detecting means. Is detected. As the refrigerant state quantity, for example, the number of revolutions of the compressor provided in the heat pump circuit, the valve opening degree of the decompression means, and the like can be considered.

  Then, the upper limit value setting means sets an upper limit value of the rotation speed of the compressor in the control of the compressor control means in correspondence with the detected refrigerant temperature and refrigerant state quantity. Thereby, for example, when the circulating fluid temperature may be remarkably lowered as described above, for example, the refrigerant temperature is relatively low or the rotational speed of the compressor is relatively high, the heat pump circuit is provided with the above Set a lower upper limit for the compressor speed. As a result, the amount of heat absorbed from the load circuit of the heat pump circuit can be suppressed, so that the circulating fluid temperature is prevented from becoming excessively low as described above, thereby preventing the occurrence of condensation in the casing of the load terminal. Can do.

  According to claim 2, the temperature of the circulating fluid supplied to the load terminal is estimated, and the circulating fluid temperature is excessively lowered by setting the upper limit value of the rotational speed of the compressor according to the estimated temperature. Can be surely prevented.

  According to the third aspect, the lower the estimated temperature of the circulating fluid, the lower the upper limit value of the number of revolutions of the compressor, thereby further reliably preventing the circulating fluid temperature from becoming excessively low.

  According to the fourth aspect, the temperature of the circulating fluid can be accurately estimated by calculating the estimated temperature based on the rotational speed of the compressor and the valve opening of the pressure reducing means.

  In addition, according to claim 5, by calculating the estimated temperature using coefficients corresponding to the rotation speed and the valve opening degree, it is possible to estimate the circulating fluid temperature quickly and accurately by calculation. it can.

  According to the sixth aspect, the circulating fluid temperature can be estimated with high accuracy reflecting the characteristic that the heat exchange efficiency between the refrigerant and the circulating fluid in the load side heat exchanger is proportional to the refrigerant circulation amount. .

  According to the seventh aspect, the valve opening degree of the pressure reducing means is controlled so that the refrigerant temperature detected by the discharge temperature detecting means becomes, for example, the target temperature. Thereby, the valve opening degree of the pressure reducing means can be optimally controlled in accordance with the heat exchange amount to be exchanged between the heat pump circuit and the load side circuit in the load side heat exchanger.

1 is an external configuration diagram of main units of a heat pump apparatus according to an embodiment of the present invention. Circuit diagram of the entire heat pump device The figure explaining the action at the time of cooling operation Functional configuration diagram of geothermal control device Functional configuration diagram of air heat control device Explanatory drawing explaining the supercooling in the comparative example which does not set the upper limit of compressor speed The figure showing the behavior of the 1st heat pump circuit of a comparative example The figure showing the relationship between the refrigerant | coolant circulation amount in an 1st heat exchanger, the evaporator inlet_port | entrance refrigerant | coolant temperature, and the going cold water temperature. The figure which shows an example of a compressor rotational speed, a rotational speed coefficient, an expansion valve opening degree, and a valve opening coefficient, and each corresponding relationship The figure showing the setting aspect of the compressor rotation speed upper limit by an upper limit setting part The figure showing the behavior of the 1st heat pump circuit of an embodiment 11. Corresponding to the behavior of FIG. 11, the estimated cold water temperature, the evaporator inlet refrigerant temperature, the first compressor rotation speed and the corresponding rotation speed coefficient, the valve opening of the first expansion valve and the corresponding valve A diagram showing the detailed time course of the opening coefficient The flowchart figure showing the control procedure which the compressor control part and temperature estimation part of a geothermal control device perform

  Hereinafter, an embodiment of the present invention will be described with reference to FIGS.

  The present embodiment is an embodiment when the present invention is applied to a composite heat source type heat pump apparatus. FIG. 1 shows an external configuration of a main unit of the heat pump apparatus 1 of the present embodiment. In FIG. 1, the heat pump device 1 of the present embodiment is a load-side circuit that circulates a circulating liquid L (for example, water or antifreeze liquid) through a ground heat heat pump unit 4, an air heat heat pump unit 5, and a heat exchange terminal 36. Terminal circulation circuit 30 and a ground heat circulation circuit 20 as a heat source connection path.

  FIG. 2 shows a circuit configuration of the entire heat pump apparatus 1 of the present embodiment. As shown in FIG. 2, the heat pump device 1 is provided in the geothermal heat pump unit 4 and can heat or cool the circulating liquid L on the heat exchange terminal 36 side using a geothermal heat source. A circuit 40, a second heat pump circuit 50 provided in the air heat heat pump unit 5 and capable of heating or cooling the circulating liquid L on the heat exchange terminal 36 side using an air heat source, the terminal circulation circuit 30, And the underground heat circulation circuit 20.

  In FIG. 2, the first heat pump circuit 40 includes a variable capacity first compressor 43, a first heat exchanger 41 as a load side heat exchanger, a first expansion valve 44 as a decompression means, and heat source side heat. A ground heat source heat exchanger 45 as an exchanger is connected in an annular shape by a first refrigerant pipe 42. The first refrigerant pipe 42 is provided with a four-way valve 46 as a switching valve for switching the flow direction of the first refrigerant C1 (see FIG. 3 described later) in the first heat pump circuit 40.

  The first heat exchanger 41 and the underground heat source heat exchanger 45 are constituted by, for example, plate heat exchangers. In this plate heat exchanger, a plurality of heat transfer plates are stacked, and the circulating fluid L (or heat medium H, see FIG. 3 described later) that is the heat medium and the refrigerant flow path through which the first refrigerant C1 is circulated. The fluid flow paths to be formed are alternately formed with each heat transfer plate as a boundary.

  Moreover, the temperature of the 1st refrigerant | coolant C1 discharged from the 1st compressor 43 is detected by the 1st refrigerant | coolant discharge temperature sensor 42a as a discharge temperature detection means. Similarly, of the refrigerant temperature sensors 42c and 42b provided in the first refrigerant pipe 42 from the first heat exchanger 41 to the underground heat source heat exchanger 45 through the first expansion valve 44, the first expansion is performed. The first refrigerant temperature sensor 42c as refrigerant temperature detecting means provided in the first refrigerant pipe 42 from the valve 44 to the first heat exchanger 41 is used for the first low pressure side (cooling) or high pressure side (heating). The temperature of the refrigerant C1 is detected. The detection results of the first refrigerant discharge temperature sensor 42 a and the first refrigerant temperature sensor 42 c are input to the underground heat control device 61.

  The second heat pump circuit 50 includes a variable capacity second compressor 53, a second heat exchanger 51 as a load side heat exchanger, a second expansion valve 54 as a pressure reducing means, and a heat source side heat exchanger. The air heat source heat exchanger 55 is annularly connected by a second refrigerant pipe 52. The second refrigerant pipe 52 is provided with a four-way valve 58 as a switching valve for switching the flow direction of the second refrigerant C2 (see FIG. 3 described later) in the second heat pump circuit 50.

  As described above, the second heat exchanger 51 is configured by, for example, a plate heat exchanger, and each of the refrigerant flow path through which the second refrigerant C2 flows and the fluid flow path through which the circulating liquid L flows. It is formed alternately with the heat transfer plate as a boundary.

  The temperature of the second refrigerant C2 discharged from the second compressor 53 is detected by a second refrigerant discharge temperature sensor 52a serving as a discharge temperature detecting means. Similarly, of the refrigerant temperature sensors 52c and 52b provided in the second refrigerant pipe 52 from the second heat exchanger 51 to the air heat source heat exchanger 55 via the second expansion valve 54, the second expansion valve The second refrigerant on the low-pressure side (during cooling) or the high-pressure side (during heating) is obtained by a second refrigerant temperature sensor 52c as refrigerant temperature detection means provided in the second refrigerant pipe 52 from 54 to the second heat exchanger 51. The temperature of C2 is detected. Furthermore, the outside air temperature is detected by the outside air temperature sensor 57. The detection results of the second refrigerant discharge temperature sensor 52a, the second refrigerant temperature sensor 52c, and the outside air temperature sensor 57 are input to the air heat control device 62.

  As the first refrigerant C1 of the first heat pump circuit 40 and the second refrigerant C2 of the second heat pump circuit 50, for example, an arbitrary refrigerant such as an HFC refrigerant such as R410A or R32 or a carbon dioxide refrigerant is used. Can be used.

  The underground heat circulation circuit 20 is configured to circulate through the underground heat circulation pump 22 having a variable rotation speed (number of rotations per unit time), the underground heat source heat exchanger 45, and the underground heat source heat exchanger 45. A ground heat exchanger 23 installed as a heat source for exchanging heat with one refrigerant C <b> 1 (in this example, in the ground) is annularly connected by a ground heat pipe 21. An antifreeze liquid to which ethylene glycol, propylene glycol, or the like is added is circulated as a heat medium H1 (see FIG. 3 to be described later) by the geothermal circulation pump 22, and the heat medium H1 is supplied to the underground heat pipe 21. An underground system turn 24 that stores and adjusts the pressure of the underground heat circulation circuit 20 is provided. The underground heat exchanger 23 is not limited to being provided in the ground, but is provided in a relatively large capacity water source such as a lake, a reservoir, a river, the sea, a hot spring, a well, and the like. You may make it heat.

  In the terminal circulation circuit 30, the first heat exchanger 41, the second heat exchanger 51, and a heat exchange terminal 36 as a load terminal such as a fan coil, a floor heating panel, a panel convector, etc. Are connected in an annular form in order from the upstream side. The load pipe 31 is provided with a circulating fluid circulating pump 32 that circulates the circulating fluid L in the terminal circulating circuit 30 and a load systern 35 that stores the circulating fluid L and adjusts the pressure of the terminal circulating circuit 30. Yes. In this example, the circulating fluid circulation pump 32 is configured to rotate at a constant speed (a constant rotational speed). The heat exchange terminal 36 can be operated by a terminal remote controller 60b. Although one heat exchange terminal 36 is provided in FIG. 2, two or more heat exchange terminals 36 may be provided in parallel, and the number and specifications are not particularly limited.

  At this time, in the terminal circulation circuit 30, the first heat exchanger 41 and the second heat exchanger 51 are connected in series, and as described above, the circulating fluid that circulates in the terminal circulation circuit 30. The first heat exchanger 41 is disposed upstream of the second heat exchanger 51 with respect to the flow of L. That is, the heat pump device 1 uses the underground heat source to heat or cool the circulating liquid L on the heat exchange terminal 36 side, and the first heat exchanger 41 of the first heat pump circuit 40 and heat using the air heat source. The second heat exchanger 51 of the second heat pump circuit 50 that heats or cools the circulating liquid L on the exchange terminal 36 side is a composite heat source heat pump device connected in series to the terminal circulation circuit 30. It is.

  The load pipe 31 is returned as circulating fluid temperature detecting means for detecting the temperature of the circulating fluid L that flows into the first heat exchanger 41 from the heat exchange terminal 36 (in other words, returns from the heat exchange terminal 36). The liquid temperature sensor 34 is provided, and the detection result is input to the underground heat control device 61 and the air heat control device 62. In addition, the structure which acquires the detection result of the return liquid temperature sensor 34 via the said underground heat control apparatus 61 may be sufficient as the air heat control apparatus 62 not directly connected to the return liquid temperature sensor 34.

  Here, the heat pump device 1 can selectively function as a heating device that performs a heating operation by switching the four-way valves 46 and 58 or a cooling device that performs a cooling operation. As an example, the case of performing the cooling operation will be described with reference to FIG.

  FIG. 3 shows a state during the cooling operation. Note that various signal lines shown in FIG. 2 are omitted in order to prevent the illustration from being complicated. In the cooling operation shown in FIG. 3, the first heat pump circuit 40 switches the four-way valve 46 as shown in the figure, so that the first refrigerant C1 discharged from the first compressor 43 is discharged into the ground. After the heat source heat exchanger 45, the first expansion valve 44, and the first heat exchanger 41 are circulated in this order, a flow path that returns to the first compressor 43 is formed. Thus, in the underground heat source exchanger 45 functioning as a condenser after the first refrigerant C1 in the gas state sucked at low temperature and low pressure is compressed by the first compressor 43 to become high temperature and high pressure gas. Then, heat exchange is performed with the heat medium H1 flowing through the underground heat circulation circuit 20, and the heat medium H1 is changed into a high-pressure liquid while releasing heat. Thus, the first refrigerant C1 that has become liquid is decompressed by the first expansion valve 44 and becomes a low-pressure liquid that easily evaporates. In the first heat exchanger 41 that functions as an evaporator, the terminal circulation circuit Heat is exchanged with the circulating fluid L flowing through 30 to evaporate and change into a gas to absorb heat and cool the circulating fluid L, and then return to the first compressor 43 again as a low-temperature and low-pressure gas.

  On the other hand, in the second heat pump circuit 50, when the four-way valve 58 is switched as shown in the figure, the second refrigerant C2 discharged from the second compressor 53 is supplied to the air heat source heat exchanger 55 and the second expansion valve. 54 and the 2nd heat exchanger 51 are distribute | circulated in order, Then, the flow path which returns to the 2nd compressor 53 is formed. Thereby, the second refrigerant C2 in the gas state sucked at a low temperature and a low pressure is compressed by the second compressor 53 to become a high temperature and a high pressure gas, and then the air heat source heat exchanger 55 functioning as a condenser. , The heat exchange with the air sent by the operation of the blower fan 56 is performed to change the liquid into a high-pressure liquid while releasing heat to the outside air. The second refrigerant C2 that has become liquid in this manner is reduced in pressure by the second expansion valve 54, becomes a low-pressure liquid, and easily evaporates. In the second heat exchanger 51 that functions as an evaporator, the terminal circulation circuit Heat is exchanged with the circulating liquid L flowing through 30 to evaporate and change into gas, thereby absorbing the heat and cooling the circulating liquid L, and then returning to the second compressor 53 again as a low-temperature and low-pressure gas.

  In the underground heat circulation circuit 20, the heat medium H <b> 1 is supplied to the underground heat source heat exchanger 45 by the underground heat circulation pump 22. In the underground heat source heat exchanger 45 that functions as a condenser, the first refrigerant C1 that flows through the refrigerant flow path of the underground heat source heat exchanger 45 and the fluid flow path of the underground heat source heat exchanger 45 are provided. After the circulating heat medium H1 faces and exchanges heat, the heat of the first refrigerant C1 at a high temperature is dissipated to the heat medium H1 side and the first refrigerant C1 is cooled, The heat of the medium H1 is radiated to the ground by the underground heat exchanger 23.

  In the terminal circulation circuit 30, the circulating liquid L that has flowed into the first heat exchanger 41 by the circulating liquid circulation pump 32 is heated in the first heat exchanger 41 that functions as an evaporator. In the second heat exchanger 51 functioning as an evaporator, the air heat source heat exchanger in the second heat exchanger 51 after being cooled by performing heat exchange with the first refrigerant C1 that has been heat-exchanged with the medium H1 and cooled as described above. At 55, heat is exchanged with the outside air, and heat is exchanged with the second refrigerant C2 cooled as described above to further cool. The circulating fluid L thus cooled is then supplied to the heat exchange terminal 36 to cool the air-conditioned space.

  In addition, in the above, although the state at the time of the air_conditionaing | cooling operation which operated both the underground heat pump unit 4 and the air heat pump unit 5 was demonstrated and demonstrated in FIG. 4, it is not restricted to this. That is, a cooling operation in which only the geothermal heat pump unit 4 is operated or a cooling operation in which only the air heat heat pump unit 5 is operated is possible.

  Next, the underground heat control device 61 and the air heat control device 62 will be described. Although not shown in detail, the underground heat control device 61 and the air heat control device 62 include a storage unit that stores various data and programs, and a control unit that performs calculation / control processing. Functional configurations of the underground heat control device 61 and the air heat control device 62 will be described with reference to FIGS. 4 and 5.

  As shown in FIG. 4, the underground heat control device 61 functionally includes a compressor control unit 61A as a compressor control unit, an expansion valve control unit 61B as a pressure reduction control unit, and a pump control unit 61C. I have. The underground heat control device 61 is communicably connected to the terminal control device 36a and the main remote controller 60a provided in the heat exchange terminal 36 (see FIG. 2). The temperature estimation unit 61D will be described later.

  The compressor control unit 61A controls the number of revolutions of the first compressor 43 according to the temperature of the circulating fluid L detected by the return fluid temperature sensor 34. Particularly in this example, the compressor controller 61A determines that the return temperature of the circulating fluid L detected by the return fluid temperature sensor 34 is a desired target temperature (circulating fluid target temperature) corresponding to, for example, the operation of the main remote controller 60a. Thus, the rotational speed of the first compressor 43 is controlled (return temperature control). The upper limit setting unit 61p will be described later.

  The expansion valve control unit 61B controls the valve opening degree of the first expansion valve 44 according to the temperature of the first refrigerant C1 detected by the first refrigerant discharge temperature sensor 42a. Particularly in this example, the expansion valve control unit 61B is configured so that the temperature of the first refrigerant C1 detected by the first refrigerant discharge temperature sensor 42a becomes a control target temperature corresponding to the operation of the main remote controller 60a, for example. The valve opening degree of the first expansion valve 44 is controlled.

  The pump control unit 61C performs the above-described operation according to the temperature of the first refrigerant C1 detected by the refrigerant temperature sensor 42b provided in the first refrigerant pipe 42 from the first expansion valve 44 to the underground heat source heat exchanger 45. The rotational speed of the intermediate heat circulation pump 22 is controlled (see also FIG. 2). Particularly in this example, the pump control unit 61C controls the rotation speed of the underground heat circulation pump 22 so that the temperature of the first refrigerant C1 detected by the refrigerant temperature sensor 42b becomes a substantially constant value. To do.

  As shown in FIG. 5, the air heat control device 62 functionally includes a compressor control unit 62A as a compressor control unit, an expansion valve control unit 62B as a pressure reduction control unit, and a fan control unit 62C. ing. The temperature estimation unit 62D will be described later. The air heat control device 62 is communicably connected to the underground heat control device 61 (see FIG. 2).

  The compressor control unit 62 </ b> A controls the rotation speed of the second compressor 53 according to the temperature of the circulating fluid L detected by the return fluid temperature sensor 34. Particularly in this example, the compressor control unit 62A determines that the return temperature of the circulating fluid L detected by the return fluid temperature sensor 34 is a desired target temperature corresponding to the operation of the main remote controller 60a, for example. The rotational speed of the second compressor 53 is controlled (return temperature control). The upper limit setting unit 62p will be described later. The compressor control unit 62A of the air heat control device 62 and the compressor control unit 61A of the underground air heat control device 61 cooperate with each other as necessary, and are the target first compressor 43. Alternatively, the second compressor 53 is controlled.

  The expansion valve control unit 62B controls the valve opening degree of the second expansion valve 54 in accordance with the temperature of the second refrigerant C2 detected by the second refrigerant discharge temperature sensor 52a. In particular, in this example, the expansion valve control unit 62B causes the temperature of the second refrigerant C2 detected by the second refrigerant discharge temperature sensor 52a to be a control target temperature corresponding to, for example, the operation of the main remote controller 60a. The valve opening degree of the second expansion valve 54 is controlled. The expansion valve control unit 62B of the air heat control device 62 and the expansion valve control unit 61B of the underground air heat control device 61 cooperate with each other as necessary, and are the target first expansion valve 44. Alternatively, the second expansion valve 54 is controlled.

  The fan control unit 62C controls the rotational speed of the blower fan 56 according to the temperature of the outside air detected by the outside air temperature sensor 57 (see also FIG. 2).

  In the heat pump device 1 having the above basic configuration and operation, the main part of this embodiment is the temperature estimation units 61D and 62D newly provided in the underground heat control device 61 and the air heat control device 62 (details will be described later). ) Based on the calculation results in), the control content (upper limit of the compressor speed) for the compressors 43 and 53 by the upper limit value setting units 61p and 62p (details will be described later) newly provided in the compressor control units 61A and 62A. Value setting, details are given later. Hereinafter, the details will be described in order.

  First, as a comparative example of the present embodiment, when the cooling operation is performed by operating only the geothermal heat pump unit 4 alone as described above, the upper limit value setting of the compressor rotation speed by the upper limit value setting unit 61p is performed. The case where no is performed will be described.

  That is, as shown in FIG. 6, when two heat exchange terminals 36A and 36B are connected in parallel to the terminal circulation circuit 30 of the heat pump device 1, and each is installed in a separate A room and B room. Is assumed. For example, as shown in FIG. 6A, when each of the heat exchange terminals 36A and 36B is operated with a cooling load of 2 [kw] and cools the rooms A and B, the first heat pump circuit 40 is Therefore, the vehicle is operated with a total output of 4 [kw]. Here, the heat exchange terminals 36A and 36B typically include a heat exchanger, a blower fan, and the like therein, and the cold water cooled by the heat pump unit 1 is passed through the heat exchanger inside. And a fan coil that drives the blower fan to exchange heat with room air to cool the room.

  Thereafter, if only the heat exchange terminal 36A in the A room is stopped while the operation of the heat exchange terminal 36B in the B room is continued, the entire cooling load is rapidly reduced from 4 [kW] to 2 [kW]. . However, at this time, as described above, only the return temperature control (= rotational speed control according to the temperature of the circulating fluid L detected by the return fluid temperature sensor 34) by the compressor control unit 61A is performed. Since the response of the output control to the increase / decrease in the cooling load is slow, as shown in FIG. 6 (b), the output of the first heat pump circuit 40 is decreased to 2 [kw] corresponding to the decrease. A large time lag occurs. As a result, the heat exchange terminal 36B in the operating B room is in a supercooled state, and the housing of the heat exchange terminal 36B may be condensed.

  A specific example of the detailed behavior of the first heat pump circuit 40 in this case is shown in the graph of FIG.

  In the drawing, FIG. 7A shows a change over time of the circulation amount [L / min] of the circulating liquid L in the terminal circulation circuit 30. Moreover, FIG.7 (b) has shown the time-dependent transition of the temperature of the circulating fluid L at that time. Note that the solid line indicates the temperature of the circulating fluid L flowing into the heat exchange terminal 36 from the first heat exchanger 41 (= forward chilled water temperature) [° C.], and the broken line indicates the first heat exchange from the heat exchange terminal 36. The temperature of the circulating liquid L returning to the vessel 41 (hereinafter referred to as “return cold water temperature” as appropriate. In this embodiment, detected by the return liquid temperature sensor 34 as described above. The same applies hereinafter) [° C.]

  FIG. 7C shows the change over time in the number of revolutions to the first compressor 43 in the first heat pump circuit 40. Moreover, FIG.7 (d) has shown the time-dependent transition of the valve opening degree of the 1st expansion valve 44 at that time. 7 (e) shows the temperature of the first refrigerant C1 flowing into the first heat exchanger 41 (= evaporator inlet refrigerant temperature. In the present embodiment, the temperature is detected by the first refrigerant temperature sensor 42c. The same applies hereinafter) [ [C] over time.

  As described above, in the state in which the cooling operation is performed by the two heat exchange terminals 36A and 36B in the A room and the B room (see the times Ta to Tb in FIGS. 7A to 7E), When the heat exchange terminal 36A is stopped and the cooling load is reduced, the circulation amount of the circulating liquid L in the terminal circulation circuit 30 is reduced by the known control of the underground heat control device 61 (time Tb in FIG. 7A). See Tc). In conjunction with this, the circulating cold water temperature of the circulating fluid L rapidly decreases (see times Tb to Tc in FIG. 7B), and the evaporator inlet refrigerant temperature on the first heat pump circuit 40 side (FIG. 7 ( e)) also decreases rapidly (see times Tb to Tc in FIG. 7E).

  Thereafter, after a time difference due to circulation of the circulating fluid L in the terminal circulation circuit 30, the return chilled water temperature of the circulating fluid L also decreases from the target temperature in the return temperature control (in this example, set to 17 ° C.). (See time Tc in FIG. 7B). However, the temperature is decreased while maintaining the temperature difference from the outgoing chilled water temperature by the amount absorbed by the heat exchange terminal 36b that is still in operation at this time (see the times Tc to Td in FIG. 7B). ).

  Here, as described above, the rotation speed of the first compressor 43 of the first heat pump circuit 40 is controlled according to the return cold water temperature under the control of the compressor control unit 61A. Therefore, as described above, the return chilled water temperature divides the target temperature (17 ° C. in the above example), and further reaches, for example, an allowable temperature that is lower than the target temperature by a predetermined deviation, thereby rotating the first compressor 43. The number is controlled so as to decrease sequentially (see times Td to Tg in FIG. 7C). At this time, the rotational speed of the first compressor 43 is normally increased or decreased in steps at predetermined time intervals (for example, 1 minute intervals) in consideration of operational stability. When the number is reduced, the temperature of the outgoing chilled water becomes extremely low (for example, 3 ° C. or lower; see times Td and Te in FIG. 7B). In this state, the heat exchange terminal 36B in operation is supercooled, and the housing of the heat exchange terminal 36B may be condensed.

  On the other hand, as described above, the opening degree of the first expansion valve 44 of the first heat pump circuit 40 is controlled according to the refrigerant discharge temperature by the control of the expansion valve control unit 61B. Therefore, as a result of the control for reducing the rotational speed of the first compressor 43, when the refrigerant discharge temperature decreases, the valve opening of the first expansion valve 44 decreases, that is, the valve of the first expansion valve 44. The opening degree is controlled in the closing direction (see times Te to Ti in FIG. 7D). Further, due to the stepwise reduction in the number of revolutions of the first compressor 43, the output of the first heat pump circuit 40 is reduced stepwise, and the forward chilled water temperature and the return chilled water temperature rise again (time of FIG. 7B). In conjunction with Te to Ti), the evaporator inlet refrigerant temperature also slightly increases.

  FIG. 7 shows the case where only the first heat pump circuit 40 is operated for cooling, but the same behavior is obtained when only the second heat pump circuit 50 is operated for cooling (detailed description is omitted).

  Therefore, in the present embodiment, in order to prevent excessive output (overcooling) by the heat pump circuit 40 immediately after the reduction of the cooling load as described above, the underground heat control device 61 is provided with a temperature estimation unit 61D and the compression The machine control unit 61A is provided with an upper limit setting unit 61p as an upper limit setting unit. Similarly, in order to prevent the excessive output (overcooling) of the heat pump circuit 50 immediately after the cooling load is reduced, the air heat control device 62 is provided with a temperature estimation unit 62D and the compressor control unit 62A as an upper limit setting means. The upper limit setting unit 62p is provided.

  The temperature estimation unit 61D of the underground heat control device 61 controls the temperature of the first refrigerant C1 (the evaporator inlet refrigerant temperature) detected by the first refrigerant temperature sensor 42c and the rotation speed control by the compressor control unit 61A. Refrigerant obtained at the time (for example, using a control command value) of the rotation speed of the first compressor 43 as the refrigerant state quantity related to the refrigerant circulation amount obtained at the time and the valve opening degree control by the expansion valve control unit 61B Based on the opening degree of the first expansion valve 44 as the refrigerant state quantity related to the circulation amount, the temperature of the circulating liquid L flowing into the heat exchange terminal 36 from the first heat exchanger 41 (the outgoing cold water temperature) is determined. presume. Then, the upper limit value setting unit 61p of the compressor control unit 61A is based on the forward chilled water temperature estimated by the temperature estimation unit 61D (hereinafter referred to as “forward chilled water estimated temperature” as appropriate). An upper limit value of the rotational speed of the first compressor 43 in the control of 61A is set.

  The temperature estimation unit 62D of the air heat control device 62 is configured to control the temperature of the second refrigerant C2 (the evaporator inlet refrigerant temperature) detected by the second refrigerant temperature sensor 52c and the rotational speed control by the compressor control unit 62A. The obtained refrigerant speed (the rotation speed of the second compressor 53 as a refrigerant state quantity related to the refrigerant circulation amount) and the refrigerant obtained at the time of valve opening control by the expansion valve control unit 62B (for example, using a control command value) Based on the valve opening degree of the second expansion valve 54 as the refrigerant state quantity related to the circulation amount, the temperature of the circulating liquid L flowing into the heat exchange terminal 36 from the second heat exchanger 51 (the forward cold water temperature) The said going cold water estimated temperature is calculated. Then, the upper limit value setting unit 62p of the compressor control unit 62A rotates the second compressor 53 in the control of the compressor control unit 62A based on the estimated forward chilled water temperature calculated by the temperature estimation unit 62D. Set the upper limit for numbers.

  Here, a method for calculating the estimated temperature of the incoming cold water in the temperature estimation units 61D and 62D will be described. Generally, in the heat exchange in the first heat exchanger 41 and the second heat exchanger 51, as shown in FIG. 8, as the evaporator inlet refrigerant temperature is lower, the outgoing chilled water temperature is also lowered and becomes the output side. The outgoing chilled water temperature is proportional to the evaporator inlet refrigerant temperature on the input side.

  At this time, as the mechanical characteristics of the heat exchangers 41 and 51, the heat conversion efficiency is higher as the circulation amount of the refrigerant on the input side is larger, and the heat conversion efficiency is lower as the circulation amount of the refrigerant on the input side is smaller. is there. That is, as described above, the graph line indicating the proportional relationship between the evaporator inlet refrigerant temperature and the outgoing cold water temperature is located on the left side as the refrigerant circulation amount increases, and located on the right side as the refrigerant circulation amount decreases. For this reason, in order to calculate the estimated temperature of the going cold water from the evaporator inlet refrigerant temperature as described above, it is necessary to calculate the refrigerant state quantity related to the refrigerant circulation amount.

  In this example, since the refrigerant circulation amount is approximately proportional to the compressor rotational speed and the valve opening, the refrigerant state quantity is the rotation of the first compressor 43 and the second compressor 53 as described above. 8 and the opening of the first expansion valve 44 and the second expansion valve 54 are used to correct the characteristic line of the solid line in FIG. 8 (see the double-headed white arrow in FIG. 8). The estimated cold water temperature is calculated from the refrigerant temperature.

That is, the correction value related to the compressor speed is the correction value A, the correction value based on the valve opening is the correction value B (valve opening coefficient), and the sum “correction value A + correction value B” is the refrigerant state quantity. As a correction value based on
Estimated cold water temperature = Evaporator inlet refrigerant temperature + Correction value A + Correction value B (Equation 1)
The estimated cold water temperature is calculated by the equation (1).

  An example of the correction value A at this time is shown in FIG. Here, the correction value A is referred to as “rotational speed coefficient” (the same applies hereinafter). In FIG. 9A, in this example, the rotational speed coefficient corresponds discretely to the three sections of the rotational speed [rps] of the compressor, and in the range where the rotational speed is 40 [rps] or less. The rotational speed coefficient is 8, and the rotational speed exceeds 40 [rps] (however, for convenience, it is expressed as “41”, the same applies hereinafter), and in the range of 60 [rps] or lower, the rotational speed coefficient is 6. In the range where the rotational speed exceeds 60 [rps] and is equal to or lower than 90 [rps], the rotational speed coefficient is 4.

  An example of the correction value B at this time is shown in FIG. Here, the correction value B is referred to as a “valve opening coefficient” (the same applies hereinafter). In FIG. 9B, in this example, the valve opening coefficient corresponds discretely to the three sections of the expansion valve opening, and in the range where the expansion valve opening (relative value) is 150 or less. The valve opening coefficient is 1, the expansion valve opening exceeds 150 (however, for convenience, it is expressed as “151” in the figure, the same applies hereinafter), and the valve opening coefficient is 0 in the range of 350 or less. In the range where the expansion valve opening exceeds 350 and is 450 or less, the valve opening coefficient is -1.

  In addition, the correspondence (correlation) shown by Fig.9 (a) and FIG.9 (b) is beforehand memorize | stored and hold | maintained in the appropriate location (for example, above-mentioned memory | storage part) of the underground heat control apparatus 61.

  Next, a method for setting the upper limit value of the compressor rotation speed in the upper limit value setting units 61p and 62p based on the estimated going-out cold water temperature calculated as described above will be described.

  In the upper limit setting units 61p and 62p, the upper limit value of the rotation speed of the compressors 43 and 53 is set higher as the estimated outgoing chilled water temperature is higher, and the rotation of the compressors 43 and 53 is decreased as the estimated outgoing chilled water temperature is lower. Set the upper limit value of the number by switching sequentially. Specifically, as shown in FIG. 10, there is no upper limit value of the rotation speed of the compressors 43 and 53 (that is, unlimited) until the estimated temperature of the incoming cold water decreases to 6 ° C. or lower (when higher than 6 ° C.). ). The upper limit value of the rotation speed of the compressors 43 and 53 is set to 50 [rps] until the estimated temperature of the incoming cold water drops below 6 ° C. and further lower than 3 ° C. (when higher than 3 ° C.) When the estimated cold water temperature falls below 3 ° C., the upper limit value of the rotational speed of the compressors 43 and 53 is set to 35 [rps].

  In this way, in the decreasing direction of the estimated going-up chilled water temperature, the dividing temperature that becomes the setting for switching the upper limit value is 6 ° C. and 3 ° C., but in the increasing direction of the estimated chilled water temperature, the dividing temperature is Change to 8 ° C and 4 ° C (= Hysteresis is given to the upper limit setting). That is, the upper limit value of the rotation speed of the compressors 43 and 53 is set to 35 [rps] until the estimated temperature of the going cold water rises to 4 ° C. or higher (when it is less than 4 ° C.), Until the temperature rises to 4 ° C. or higher and becomes 8 ° C. or higher (less than 8 ° C.), the upper limit value of the rotation speed of the compressors 43 and 53 is set to 50 [rps], Is set to no upper limit (that is, unlimited) for the rotation speed of the compressors 43 and 53.

  In the present embodiment, the upper limit value setting units 61p and 62p perform the setting of the compressor rotation speed upper limit value as described above. Thereby, for example, the first heat pump circuit 40 is cooled by the behavior shown in FIG. 11 corresponding to FIG. That is, in the state where the cooling operation is performed by the two heat exchange terminals 36A and 36B (see times T1 to T7 in FIGS. 11A to 11E), the cooling load is reduced by stopping the heat exchange terminal 36A. Then (see time T7 in FIG. 11 (a)), the circulation amount of the circulating liquid L, the outgoing chilled water temperature, and the evaporator inlet refrigerant temperature are respectively reduced (FIG. 11 (a), FIG. 11 (b), FIG. 11). (See time T7 to T8 in (e)).

  At this time, based on the decrease in the evaporator inlet refrigerant temperature, the estimated temperature of the chilled water is from 9 ° C. (refer to times T1 to T7 in FIG. 11 (f)) to 5 ° C. by the temperature estimation units 61D and 62D. It is calculated that it has decreased (see time T8 in FIG. 11 (f)). Thus, as a result of the upper limit value setting unit 61p changing the upper limit value of the rotation speed of the compressor to 50 [rps] based on the setting shown in FIG. 10, the actual rotation speed of the first compressor 43 is also 50 [ rps] (forcibly) (see time T8 in FIG. 11C). As a result, the output of the first heat pump circuit 40 decreases, and the outgoing chilled water temperature that has started to decrease as described above is indicated by the return temperature control to decrease the compressor rotational speed based on the aforementioned decrease in the return chilled water temperature. For example, it stops at around 5 ° C. (see time T8 in FIG. 11B). As a result, as described above with reference to FIG. 7B, it is prevented that the temperature of the forward chilled water becomes extremely low and the heat exchange terminal 36B is overcooled (see times T8 to T12 in FIG. 11B). ).

  In the figure, the rotational speed of the first compressor 43 is reduced even after the going-out chilled water temperature rises (see times T12 to T13 in FIG. 11C), which is compressed by the return temperature control. Since the rotational speed instructed to the machine falls below the rotational speed upper limit value by the upper limit value setting unit 61p at this time, the rotational speed is reduced in such a manner that the return temperature control is prioritized (time T12). To T13).

  In the behavior of FIG. 11 as described above, the estimated cold water temperature, the evaporator inlet refrigerant temperature, the rotational speed of the first compressor 43 and the corresponding rotational speed coefficient, and the first expansion valve 44. FIG. 12 shows a detailed time course of the valve opening degree and the valve opening coefficient corresponding thereto.

  11 shows the case where only the first heat pump circuit 40 is cooled, as in FIG. 7, but the same applies to the second compressor 53 even when only the second heat pump circuit 50 is cooled. And the same behavior is obtained (detailed explanation is omitted).

  Next, in order to realize the above method, a control procedure executed by the compressor control unit 61A and the temperature estimation unit 61D of the underground heat control device 61 when the first heat pump circuit 40 performs a cooling operation is illustrated in FIG. This will be described with reference to the flowchart of FIG. In FIG. 13, first, in step S10, the compressor control unit 61A determines whether or not the heat pump device 1 has started operation. Specifically, the operation start state is, for example, a case where the operation is started from a stop state by an appropriate operation start operation of the heat pump device 1 by an operator, or a return from a standby state described later to the heat pump device. 1 is started again (details will be described later). Until the operation start state is reached, the determination in step S10 is not satisfied (S10: No), and the loop waits. When the operation start state is reached, the determination in step S10 is satisfied (S10: Yes), and the process proceeds to step S15.

  In step S15, the compressor control unit 61A waits for a predetermined time (in this example, 1 minute) as an idle period for stabilizing the operation of the first compressor 43. Thereafter, the process proceeds to step S20.

  In step S20, the compressor control unit 61A acquires the detection result of the return chilled water temperature by the return liquid temperature sensor 34 (in other words, detects the return chilled water temperature; the same applies hereinafter). Thereafter, the process proceeds to step S25.

  In step S25, the compressor control unit 61A calculates an instruction rotational speed N1 for the first compressor 43 for setting the return chilled water temperature to the target temperature based on the return chilled water temperature detected in step S20 ( = Return temperature control). Thereafter, the process proceeds to step S30.

  In step S30, the temperature estimation unit 61D acquires the detection result of the evaporator inlet refrigerant temperature by the first refrigerant temperature sensor 42c (in other words, detects the evaporator inlet refrigerant temperature; the same applies hereinafter). Thereafter, the process proceeds to step S35.

  In step S35, the temperature estimation unit 61D detects the rotation speed detection result of the first compressor 43 at this time (detected by an appropriate known method and input to the compressor control unit 61A). Obtained from 61A. Thereafter, the process proceeds to step S40.

  In step S40, the temperature estimation unit 61D refers to the correspondence relationship in FIG. 9A and determines the rotation speed coefficient corresponding to the compressor rotation speed acquired in step S35. Thereafter, the process proceeds to step S45.

  In step S45, the temperature estimation unit 61D performs expansion valve control on the detection result of the valve opening degree of the first expansion valve 44 at this time point (detected by an appropriate known method and input to the expansion valve control unit 61B). Obtained from the unit 61B. Thereafter, the process proceeds to step S50.

  In step S50, the temperature estimation unit 61D determines the valve opening coefficient corresponding to the valve opening acquired in step S45 with reference to the correspondence relationship in FIG. Thereafter, the process proceeds to step S55.

  In step S55, the temperature estimation unit 61D adds the evaporator inlet refrigerant temperature detected in step S30, the rotation speed coefficient determined in step S40, and the valve opening coefficient acquired in step S50. The estimated cold water temperature is calculated. Thereafter, the process proceeds to step S60.

  In step S60, the compressor control unit 61A obtains the estimated amount of forward chilled water calculated by the temperature estimation unit 61D in step S55 by the upper limit value setting unit 61p, and refers to the setting of FIG. An upper limit value N2 of the compressor rotational speed corresponding to the acquired forward cold water estimated temperature is set. Thereafter, the process proceeds to step S65.

  In step S65, the compressor controller 61A determines whether the upper limit value N2 of the compressor speed set in step S60 is unlimited (in other words, corresponds to the “no upper limit of compressor speed” shown in FIG. 10). Determine whether or not. If the upper limit value N2 is set without limitation, the determination at Step S65 is satisfied (S65: YES), and the routine goes to Step S70.

  In step S70, the compressor control unit 61A instructs (outputs a control signal) the rotational speed of the first compressor 43 to be the rotational speed N1 calculated in step S25. Thereafter, the process proceeds to step S75 described later.

  On the other hand, in the determination in step S65, if the upper limit value N2 is not unlimited (in other words, does not correspond to the region of “no upper limit of compressor speed” shown in FIG. 10), the determination in step S65 is satisfied. If not (S65: NO), the process proceeds to step S80.

  In step S80, the compressor control unit 61A determines whether or not the rotation speed N1 calculated in step S25 is greater than the upper limit value N2 set in step S60. If N1 ≦ N2, the determination in step S80 is not satisfied (S80: NO), and the process proceeds to step S70. If N1> N2, the determination in step S80 is satisfied (S80: YES), and the process proceeds to step S85.

  In step S85, the compressor control unit 61A instructs (outputs a control signal) the rotational speed of the first compressor 43 to be the rotational speed N2, which is equal to the upper limit value set in the step S60. Thereafter, the process proceeds to step S75.

  In step S75, the compressor control unit 61A determines whether or not the heat pump device 1 is in an operation end state. That is, when the cooling operation is performed under the control of the compressor rotation speed as described above and the cooling load is reduced, the return liquid temperature sensor 34 of the terminal circulation circuit 30 detects the cooling load without operating the heat pump device 1. In some cases, the return cold water temperature of the circulating fluid L reaches the target temperature or lower. In this case, the heat pump device 1 is stopped by a known control by the underground heat control device 61 and enters a standby state (that is, the operation of the heat pump device 1 is once terminated). In step S75, the compressor control unit 61A determines whether or not the heat pump device 1 is in this standby state. While the operation is not finished (that is, in the standby state), the determination in step S75 is not satisfied (S75: No), the process returns to step S15, and the same procedure is repeated. When the heat pump device 1 is in the operation end state (that is, the standby state), the determination in step S75 is satisfied (step S75: YES), and the process returns to step S10. At this time, the determination of whether or not the operation start state has been reached in step S10 is a determination of whether or not the standby state has been released. That is, after the return chilled water temperature of the circulating fluid L reaches the target temperature or less and enters a standby state as described above, when the return chilled water temperature of the circulating fluid L again exceeds the target temperature, The operation of the heat pump device 1 is started again by known control by the intermediate heat control device 61. Therefore, in step S10 at this time, the compressor control unit 61A determines whether or not the heat pump device 1 has thus returned from the standby state and has been restarted. When the operation is resumed and the determination in step S10 is satisfied, the same procedure is repeated.

  In addition, although illustration is abbreviate | omitted, when the operation completion | finish operation of the appropriate heat pump apparatus 1 is made by the operator at the arbitrary timings in each of the above procedures, this flow is ended and the heat pump apparatus 1 is stopped. .

  In addition, in the above, the temperature estimation part 61D which performs step S35 and step S45 functions as a state quantity detection means as described in each claim. Moreover, the temperature estimation part 61D which performs step S40, step S50, and step S55 functions as the estimated temperature calculation means described in each claim.

  When the second heat pump circuit 50 performs a cooling operation, the compressor control unit 62A and the temperature estimation unit 62D of the air heat control device 62 execute the same control procedure as in FIG. The rotation speed is controlled (details are omitted).

  As described above, according to the heat pump device 1 of the present embodiment, in the first and second heat pump circuits 40, 50, the evaporator inlet refrigerant temperature is set by the first and second refrigerant temperature sensors 42c, 52c. A refrigerant state quantity that is detected and related to the refrigerant circulation quantity in the first and second heat pump circuits 40, 50 (in this example, the rotational speeds of the first and second compressors 43, 53, the first and second The valve opening degree of the expansion valves 44 and 54) is acquired.

  Then, the upper limit values of the rotation speeds of the first and second compressors 43 and 53 are set by the upper limit value setting units 61p and 62p corresponding to the evaporator inlet refrigerant temperature, the rotation speed and the valve opening. The As a result, when there is a possibility that the actual outgoing chilled water temperature may be significantly lowered as described above, the rotational speeds of the first and second compressors 43 and 53 are kept low. As a result, the amount of heat absorbed from the terminal circulation circuit 30 by the first and second heat pump circuits 40, 50 can be suppressed, so that it is possible to prevent the outgoing chilled water temperature from becoming excessively low as described above. It is possible to prevent the occurrence of condensation on the casing surface of the exchange terminal 36 (the heat exchange terminal 36B in the above example).

  In this embodiment, in particular, the temperature of the forward chilled water supplied to the heat exchange terminal 36 is estimated by the temperature estimation units 61D and 62D, and the upper limit setting units 61p and 62p are configured to perform the first operation according to the estimated temperature of the forward chilled water. And the upper limit of the rotational speed of the second compressors 43 and 53 is set. Thereby, it can prevent reliably that the going cold water temperature becomes low too much.

  In the present embodiment, in particular, the upper limit value setting units 61p and 62p set the rotation speed upper limit values of the first and second compressors 43 and 53 lower as the outgoing cold water estimated temperature is lower. As a result, it is possible to more reliably prevent the incoming cold water temperature from becoming excessively low.

  In this embodiment, in particular, the estimated temperature of the forward cold water is calculated based on the rotational speeds of the first and second compressors 43 and 53 and the valve opening degrees of the first and second expansion valves 44 and 54. Thereby, the going cold water estimated temperature can be accurately calculated.

  In this embodiment, in particular, the estimated temperature of the going cold water is calculated using the rotation speed coefficient and the valve opening coefficient, which are coefficients corresponding to the rotation speed and the valve opening degree, respectively. Thereby, it is possible to estimate the chilled water temperature quickly and accurately by calculation.

  In the present embodiment, in particular, the estimated going-out cold water temperature is calculated as the sum of the evaporator inlet refrigerant temperature, the rotation speed coefficient, and the valve opening coefficient. As a result, the outgoing chilled water is estimated with high accuracy reflecting the characteristic that the heat exchange efficiency between the refrigerants C1 and C2 and the circulating liquid L in the first and second heat exchangers 41 and 51 is proportional to the refrigerant circulation amount. The temperature can be calculated.

  In the present embodiment, in particular, the valve openings of the first and second expansion valves 44 and 54 are set so that the refrigerant temperature detected by the first and second refrigerant discharge temperature sensors 42a and 52a becomes a target temperature. Be controlled. Thereby, in the first and second heat exchangers 41 and 51, the first and second heat pump circuits 40 and 50 and the terminal circulation circuit 30 are adapted to the heat exchange amount to be exchanged between the first and second heat exchanger circuits 41 and 51. The valve opening degree of the second expansion valves 44 and 54 can be optimally controlled.

  In the above description, the case where the two heat exchange terminals 36A and 36B are connected in parallel has been described as an example. However, the present invention is not limited to this, and only one heat exchange terminal 36 is connected. However, the same problem of supercooling can occur. That is, for example, when the load pipe 31 connecting the heat exchange terminal 36 and the heat pump units 4 and 5 is relatively long, the circulating liquid L in the load pipe 31 is warmed at the start of the cooling operation. There may be a relatively hot condition. In such a case, even if the cooling operation is started, immediately after that, the temperature of the circulating fluid L detected by the return fluid temperature sensor 34 does not readily fall to the target temperature, and as a result, the compressor is controlled by the return temperature control. The rotational speeds 43 and 53 are controlled to the maximum rotational speed. Thereafter, when a certain amount of time has passed, the temperature of the circulating fluid L in the load pipe 31 is lowered to the target temperature. However, when the reduction is detected by the return liquid temperature sensor 34, the low-temperature circulating liquid L has already been supplied to the heat exchange terminal 36 (due to the time lag when the certain amount of time has elapsed). Depending on the situation, the heat exchange terminal 36 may be supercooled. Even in such a case, by performing the control of the method, the one heat exchange terminal 36 connected to the relatively long load pipe 31 causes dew condensation of the casing due to the overcooling as described above. Can be prevented.

  The present invention is not limited to the above-described embodiment, and various modifications can be made without departing from the scope of the invention. For example, in the above-described embodiment, a ground heat exchanger 23 is provided in the ground or the relatively large capacity water source, and the heat medium that exchanges heat with the ground or the water source by the ground heat exchanger 23. Although H1 was circulated in the underground heat circulation circuit 20, it is not restricted to this. That is, instead of configuring such a circulation circuit, an open type pipe line may be connected to the underground heat circulation pump 22. In this case, the upstream side (pump inflow side) and the downstream side (pump outflow side) of the geothermal circulation pump 22 are supplied with water sources (or constant temperature water) such as the aforementioned lakes, reservoirs, rivers, seas, hot springs and wells, respectively. It may be connected to a chilled water supply), and the water from the water source or the like is directly pumped by the geothermal circulation pump 22 and used. That is, water such as the water source is supplied to the underground heat circulation pump 22 through a pipe line (upstream pipe line) connected to the pump upstream side, and a pipe line (downstream pipe line) connected to the pump downstream side. ), And is then guided to the underground heat source heat exchanger 45 provided in the downstream pipe line to exchange heat with the first refrigerant C1, and further through the downstream pipe line, the water source. And so on. In this case, the water source and the like connected to the upstream pipeline and the water source and the like connected to the downstream pipeline may be the same or different.

  Further, for example, in the above-described embodiment, the case where only one underground heat exchanger 23 is provided in the ground is described as an example. However, the present invention is not limited to this, and a plurality of underground heat exchangers 23 are provided in the underground. It may be done. In that case, the plurality of underground heat exchangers 23 may be connected in parallel to each other or may be connected in series.

  Moreover, in the said embodiment, the case where this invention is applied to the composite heat source type heat pump apparatus provided with the 1st heat pump circuit 40 which used underground heat, and the 2nd heat pump circuit 50 which used air heat demonstrated as an example. However, it is not limited to this. That is, it may be applied to a single heat source type heat pump device including only the first heat pump circuit 40 using the underground heat, and is a single heat source including only the second heat pump circuit 50 using the air heat. You may apply to a heat pump apparatus of a type. Further, the present invention may be applied to a composite heat source type heat pump apparatus including the first heat pump circuit 40 and including three or more heat pump circuits. Furthermore, each heat pump circuit is not limited to the one using ground heat or air heat as the heat source as described above, and may use another appropriate heat source.

DESCRIPTION OF SYMBOLS 1 Heat pump apparatus 20 Geothermal circulation circuit 21 Geothermal piping 22 Geothermal circulation pump 23 Geothermal heat exchanger 30 Terminal circulation circuit (load side circuit)
31 Load piping (circulating fluid piping)
32 Circulating fluid circulation pump (load-side circulation pump)
34 Return fluid temperature sensor (circulating fluid temperature detection means)
36 Heat exchange terminal (load terminal)
40 1st heat pump circuit (heat pump circuit)
41 1st heat exchanger (load side heat exchanger)
42 First refrigerant piping (refrigerant piping)
42a First refrigerant discharge temperature sensor (discharge temperature detection means)
42c 1st refrigerant | coolant temperature sensor (refrigerant temperature detection means)
43 First compressor (compressor)
44 1st expansion valve (pressure reduction means)
45 Ground heat source heat exchanger (heat source side heat exchanger)
50 Second heat pump circuit (heat pump circuit)
51 2nd heat exchanger (load side heat exchanger)
52 Second refrigerant piping (refrigerant piping)
52a Second refrigerant discharge temperature sensor (discharge temperature detection means)
52c Second refrigerant temperature sensor (refrigerant temperature detection means)
53 Second compressor (compressor)
54 Second expansion valve (pressure reduction means)
55 Air heat source heat exchanger (heat source side heat exchanger)
57 Outside air temperature sensor 61 Geothermal control device 61A Compressor control unit (compressor control means)
61B Expansion valve controller (pressure reduction control means)
61C Pump control unit 61D Temperature estimation unit (estimated temperature calculation means, state quantity detection means)
61p Upper limit setting unit (upper limit setting means)
62 Air Heat Control Device 62A Compressor Control Unit (Compressor Control Unit)
62B Expansion valve controller (pressure reduction control means)
62C Fan control unit 62D Temperature estimation unit (estimated temperature calculation means, state quantity detection means)
62p Upper limit setting unit (upper limit setting means)
C1 First refrigerant C2 Second refrigerant H1 Heat medium L Circulating fluid

Claims (7)

A compressor, a load side heat exchanger as an evaporator, a decompression means, and a heat source side heat exchanger as a condenser capable of exchanging heat with the heat source side, connected by refrigerant piping to form a heat pump circuit;
Connecting the load-side heat exchanger and at least one load terminal with a circulating fluid pipe to form a load-side circuit;
In the load circuit, circulating fluid temperature detection means for detecting the circulating fluid temperature returning from the load terminal to the load side heat exchanger;
Compressor control means for controlling the rotational speed of the compressor so that the circulating fluid temperature detected by the circulating fluid temperature detection means becomes a target circulating fluid temperature;
In the heat pump device provided with
In the heat pump circuit, a refrigerant temperature detecting means for detecting a refrigerant temperature flowing into the load side heat exchanger,
State quantity detection means for detecting a refrigerant state quantity related to the refrigerant circulation quantity in the heat pump circuit;
Corresponding to the refrigerant temperature detected by the refrigerant temperature detection means and the refrigerant state quantity detected by the state quantity detection means, an upper limit value of the rotation speed of the compressor in the control of the compressor control means is set. An upper limit value setting means for setting is provided. Estimation of the circulating fluid supplied from the load-side heat exchanger to the load terminal based on the refrigerant temperature detected by the refrigerant temperature detection means and the refrigerant state quantity detected by the state quantity detection means It further has estimated temperature calculation means for calculating temperature,
The upper limit value setting means includes:
The heat pump apparatus according to claim 1, wherein an upper limit value of the rotation speed of the compressor in the control of the compressor control means is set according to the estimated temperature calculated by the estimated temperature calculation means. The upper limit value setting means includes:
The higher the estimated temperature, the higher the upper limit value,
The heat pump device according to claim 2, wherein the upper limit value is set lower as the estimated temperature is lower. The state quantity detection means includes
Detecting the rotational speed of the compressor and the valve opening of the pressure reducing means as the refrigerant state quantity;
The estimated temperature calculating means includes
3. The estimated temperature is calculated based on the refrigerant temperature detected by the refrigerant temperature detecting means and the rotation speed and the valve opening detected by the state quantity detecting means. Item 4. The heat pump device according to Item 3. The estimated temperature calculating means includes
5. The heat pump according to claim 4, wherein the estimated temperature is calculated using the refrigerant temperature, a rotation speed coefficient corresponding to the rotation speed, and a valve opening coefficient corresponding to the valve opening degree. apparatus. The estimated temperature calculating means includes
6. The heat pump apparatus according to claim 5, wherein the estimated temperature is calculated from a sum of the refrigerant temperature, the rotation speed coefficient, and the valve opening coefficient. In the heat pump circuit, discharge temperature detection means for detecting a refrigerant temperature discharged from the compressor,
The pressure reducing control means for controlling the valve opening degree of the pressure reducing means according to the refrigerant temperature detected by the discharge temperature detecting means. The heat pump apparatus as described.

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