JP4336619B2 - Engine heat pump - Google Patents

Engine heat pump Download PDF

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Publication number
JP4336619B2
JP4336619B2 JP2004150371A JP2004150371A JP4336619B2 JP 4336619 B2 JP4336619 B2 JP 4336619B2 JP 2004150371 A JP2004150371 A JP 2004150371A JP 2004150371 A JP2004150371 A JP 2004150371A JP 4336619 B2 JP4336619 B2 JP 4336619B2
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Prior art keywords
heat exchanger
refrigerant
compressor
liquid refrigerant
auxiliary compressor
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JP2005331177A (en
Inventor
健一 南
栄太 呉服
圭史 山中
洋志 東
二朗 福留
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ヤンマー株式会社
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B41/00Fluid-circulation arrangements
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B27/00Machines, plant, or systems, using particular sources of energy
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B1/00Compression machines, plant, or systems with non-reversible cycle
    • F25B1/10Compression machines, plant, or systems with non-reversible cycle with multi-stage compression
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B2400/00General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/07Details of compressors or related parts
    • F25B2400/075Details of compressors or related parts with parallel compressors
    • F25B2400/0751Details of compressors or related parts with parallel compressors the compressors having different capacities
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25BREFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT-PUMP SYSTEMS
    • F25B2400/00General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/13Economisers

Description

  The present invention relates to an apparatus configuration of an engine heat pump, and more particularly to a technique for reducing the total compression work without newly increasing the amount of electric power used.
  Regarding an engine heat pump having a configuration in which a compressor is driven by an engine, a configuration shown in Patent Document 1 is known. In Patent Document 1, the compression work of the engine heat pump is divided into two systems: compression work by the main compressor and compression work by the auxiliary compressor, and the evaporation pressure on one side (auxiliary compressor side) is set on the other side (main compression). An invention is disclosed in which the total compression work in the engine heat pump is reduced by reducing the compression work on one side by keeping the pressure higher than the evaporation pressure on the machine side.
JP 2004-20153 A
  Patent Document 1 discloses a configuration in which the compression work on the side where the evaporation pressure becomes high (auxiliary compressor side) is performed by an electrically driven compressor (electric compressor). A device that requires electric power (the electric compressor) is additionally provided in the engine heat pump. In this case, although the compression work can be reduced, the amount of power used is increased, resulting in the fact that the original merit of the engine heat pump called “reduction of power usage” cannot be fully utilized. Therefore, an object of the present invention is to reduce the compression work and improve the operation efficiency (energy efficiency) without increasing the amount of power used in an engine heat pump.
  The problems to be solved by the present invention are as described above. Next, means for solving the problems will be described.
That is, in claim 1, a main compressor and an auxiliary compressor driven by an engine, an indoor heat exchanger, an outdoor heat exchanger, an expansion valve for an indoor heat exchanger, an expansion valve for an outdoor heat exchanger, and an indoor heat exchanger And a subcooling heat exchanger that supercools the liquid refrigerant before branching by the subcooling liquid refrigerant that is provided in the liquid refrigerant passage path of the connection path between the outdoor heat exchanger and the branch path. the refrigerant discharged from the machine in a configuration with the engine heat pump is merged with the refrigerant discharged from said main compressor, the supercooling liquid refrigerant, said after passage of the supercooling heat exchanger is compressed by the auxiliary compressor The main compression is configured such that the capacity ratio of the auxiliary compressor to the total capacity of the main compressor and the auxiliary compressor is 20% to 29%, and the driving force of the engine is connected to the main compressor and the auxiliary compressor. Machine clutch and It is provided with a clutch auxiliaries compressor.
  According to a second aspect of the present invention, in the engine heat pump according to the first aspect, an engine waste heat recovery unit is provided in parallel with the outdoor heat exchanger, and the supercooling liquid refrigerant is evaporated by the engine waste heat recovery unit and an auxiliary compressor is provided. It is set as the structure compressed by.
  As effects of the present invention, the following effects can be obtained.
According to the first aspect of the present invention, a supercooling refrigerant having an evaporation pressure (refrigerant suction pressure) higher than that of the refrigerant compressed by the main compressor is compressed by an auxiliary compressor driven by the engine. Can reduce the total compression work in the refrigerant cycle without newly increasing the amount of electric power used for the auxiliary compressor that had been electrically driven, and maintain the cooling capacity by the supercooling action of the supercooling heat exchanger Or improvement can be achieved.
In addition, by configuring the auxiliary compressor so that the capacity ratio with respect to the total capacity of the main compressor and the auxiliary compressor is within a predetermined numerical range, it is possible to maintain or improve the cooling capacity during cooling and at the time of heating. The performance of the supercooling heat exchanger can be ensured. That is, in the configuration of the present invention in which the main compressor and the auxiliary compressor are driven by a common engine, it is possible to operate with good operation efficiency (energy efficiency) during cooling and heating.
In addition, since the main compressor clutch and auxiliary compressor clutch for connecting / disconnecting the driving force of the engine to and from the main compressor and auxiliary compressor are provided, the main compressor 2 and the auxiliary compression according to the level of the required air conditioning load. Since the operation of the machine 3 can be switched, the operation state at the partial load where the combustion efficiency of the engine 4 is reduced can be reduced, so that the operation efficiency (energy efficiency) can be improved.
According to the second aspect of the present invention, it is possible to reduce the total compression work during cooling by configuring the auxiliary compressor so that the capacity ratio with respect to the total capacity of the main compressor and the auxiliary compressor falls within a predetermined numerical range, and heating. Even at times, it is possible to reduce the total compression work without newly increasing the amount of power used.
Further, by supercooling the liquid refrigerant even during heating, the heat absorption capacity from the outside air per unit mass flow rate of the refrigerant is improved by the supercooling action, so that the total amount of refrigerant flowing through the refrigerant cycle can be reduced. . As a result, the total compression work can be reduced, and the operation efficiency (energy efficiency) can be improved.
  Next, embodiments of the invention will be described with reference to the drawings. FIG. 1 is a refrigerant circuit diagram of an engine heat pump according to the present invention, FIG. 2 is a block diagram of control devices, FIG. 3 is a Mollier diagram based on the refrigerant circuit configuration, and FIG. 4 is a relationship between auxiliary compressor capacity ratio and COP. FIG. 5 is a graph showing the relationship between the auxiliary compressor capacity ratio and the supercooling heat exchanger refrigerant temperature.
  First, a refrigerant circuit configuration and a refrigerant cycle of an engine heat pump according to the present invention will be described with reference to FIG. The engine heat pump according to the present invention includes a main compressor 2 and an auxiliary compressor 3 driven by the engine 4, an indoor heat exchanger 8, an outdoor heat exchanger 5, an indoor heat exchanger expansion valve 23, and an outdoor heat exchanger expansion valve. 21 and the supercooled liquid refrigerant that is provided in the main path 26 that is the liquid refrigerant passage path among the connection paths between the indoor heat exchanger 8 and the outdoor heat exchanger 5 and is branched to the branch path 27 (27a, 27b), It has a supercooling heat exchanger 15 that supercools the liquid refrigerant before branching, and uses a refrigerant cycle composed of these. The supercooling heat exchanger 15 has connection points 15 a and 15 b with the main path 26 and connection points 15 c and 15 d with the branch path 27. In this configuration, a plurality of indoor heat exchangers 8 may be provided.
  The main compressor 2 is driven by the engine 4, sucks and compresses the gas refrigerant from which the liquid refrigerant has been separated by an accumulator (not shown), and discharges the high-temperature and high-pressure gas refrigerant. The gas refrigerant discharged from the main compressor 2 is guided in a predetermined direction by the four-way valve 24. Further, since the gas refrigerant sucked into the main compressor 2 is also guided by the four-way valve 24, the refrigerant inlet of the main compressor 2 and the four-way valve 24 communicate with each other through a path 32 that constitutes a suction line of the main compressor 2. Has been. The auxiliary compressor 3 is also driven by the engine 4, and the liquid refrigerant is separated by an accumulator (not shown) out of the supercooling liquid refrigerant that is branched to the branch path 27 and passes through the supercooling heat exchanger 15. Gas refrigerant is sucked and compressed, and high-temperature and high-pressure gas refrigerant is discharged.
  The subcooling heat exchanger 15 supercools the liquid refrigerant before branching by the subcooling liquid refrigerant whose temperature has been lowered by the subcooling heat exchanger expansion valve 22 provided in the branch path 27, and this supercooling heat. The supercooled liquid refrigerant after heat exchange in the exchanger 15 is sucked into the auxiliary compressor 3. For this reason, the supercooling heat exchanger 15 and the refrigerant inlet of the auxiliary compressor 3 are communicated with each other through a path 33 constituting a suction line of the auxiliary compressor 3.
  The branch path 27 provided in the main path 26 constitutes a branch path 27 a between the indoor heat exchanger 8 and the supercooling heat exchanger 15, and the outdoor heat exchanger 5 and the supercooling heat exchanger 15. A branch path 27b is formed between them, and on-off valves 28a and 28b are provided between the branch paths 27a and 27b and the subcooling heat exchanger expansion valve 22, respectively. These on-off valves 28a and 28b are switched between open and closed so that the liquid refrigerant before branching of the main path 26 is supercooled in a cooling cycle or heating cycle to be described later.
  The refrigerant discharged from the auxiliary compressor 3 is merged with the refrigerant discharged from the main compressor 2 at a junction 65 provided in the path from the compressors 2 and 3 to the four-way valve 24. Yes. The direction in which the merged refrigerant flows in the four-way valve 24 is changed, and a cooling cycle or a heating cycle described later is performed. An oil separator (not shown) is provided between the junction 65 and the four-way valve 24 to separate the refrigeration oil contained in the high-temperature and high-pressure gas refrigerant so that the main compressor 2 and the auxiliary compressor 3 The refrigerant is refluxed to the suction side so that both the compressors 2 and 3 are lubricated satisfactorily.
  Using the refrigerant cycle configured as described above, the cooling cycle or the heating cycle is performed by switching the direction in which the refrigerant flows by the four-way valve 24. In the cooling cycle, the refrigerant compressed by the main compressor 2 and the auxiliary compressor 3 merges at the merge point 65 and is sent to the outdoor heat exchanger 5 via the four-way valve 24. The outdoor heat exchanger 5 After being dissipated and condensed, it is sent to the supercooling heat exchanger 15 and flows in from the connection point 15b and flows out from the connection point 15a. The liquid refrigerant supercooled in the supercooling heat exchanger 15 expands in the indoor heat exchanger expansion valve 23, absorbs heat in the indoor heat exchanger 8 and evaporates, and then passes through the four-way valve 24 to the main compressor. 2 is aspirated. The sucked refrigerant is compressed by the main compressor 2 and then discharged again.
  Further, a part of the liquid refrigerant sent out from the outdoor heat exchanger 5 and passing through the main path 26 is diverted to the branch path 27a as the supercooling liquid refrigerant, and is expanded and the temperature is lowered by the expansion valve 22 for the supercooling heat exchanger. Thus, the liquid refrigerant flowing through the main path 26 is supercooled in the process of becoming a low-temperature wet refrigerant and flowing into the supercooling heat exchanger 15 from the connection point 15c and flowing out to the connection point 15d. At this time, the on-off valve 28a is open and the on-off valve 28b is closed, and the liquid refrigerant passing through the main path 26 is branched to the branch path 27a without being diverted to the branch path 27b side. The total amount of liquid refrigerant before branching is supercooled by the supercooling liquid refrigerant. Thus, the refrigeration cycle efficiency is improved by supercooling the liquid refrigerant passing through the main path 26. The supercooled liquid refrigerant is sucked into the auxiliary compressor 3, compressed by the auxiliary compressor 3, and then discharged again.
  On the other hand, in the heating cycle, the refrigerant compressed by the main compressor 2 and the auxiliary compressor 3 merges at the junction 65, and is sent to the indoor heat exchanger 8 via the four-way valve 24. After the heat is dissipated and condensed in the vessel 8, it is sent to the supercooling heat exchanger 15, and flows in from the connection point 15a and flows out from the connection point 15b. The liquid refrigerant supercooled in the supercooling heat exchanger 15 expands in the outdoor heat exchanger expansion valve 21, absorbs heat in the outdoor heat exchanger 5 and evaporates, and then passes through the four-way valve 24 to the main compressor. 2 is aspirated. The sucked refrigerant is compressed by the main compressor 2 and then discharged again.
  Further, a part of the liquid refrigerant sent out from the indoor heat exchanger 8 and passing through the main path 26 is diverted to the branch path 27b as the supercooling liquid refrigerant, and is expanded and the temperature is lowered by the expansion valve 22 for the supercooling heat exchanger. Thus, the liquid refrigerant flowing through the main path 26 is supercooled in the process of becoming a low-temperature wet refrigerant and flowing into the supercooling heat exchanger 15 from the connection point 15c and flowing out to the connection point 15d. At this time, the on-off valve 28a is in a closed state and the on-off valve 28b is in an open state, and the liquid refrigerant passing through the main path 26 is branched to the branch path 27b without being diverted to the branch path 27a side. The total amount of liquid refrigerant before branching is supercooled by the supercooling liquid refrigerant. The supercooled liquid refrigerant that has passed through the supercooling heat exchanger 15 absorbs heat and evaporates in the engine waste heat recovery unit 6, is sucked into the auxiliary compressor 3, and is compressed in the auxiliary compressor 3. It is discharged again later.
  Next, an apparatus configuration relating to operation control of the engine heat pump according to the present invention will be described with reference to FIG. The controller 25 which is a control device provided in the engine heat pump according to the present invention is connected to the expansion valve 21 for the outdoor heat exchanger, the expansion valve 22 for the supercooling heat exchanger, and the expansion valve 23 for the indoor heat exchanger. 25 controls the opening of each expansion valve.
  Similarly, the controller 25 is connected to the on-off valves 28a and 28b provided on the branch paths 27a and 27b, respectively, and controls the opening and closing thereof. Here, each on-off valve 28a, 28b is specifically controlled as follows. That is, the on-off valve 28a is opened when the liquid refrigerant is supercooled in the above-described cooling cycle, and is closed otherwise. The on-off valve 28b is opened when the liquid refrigerant is supercooled in the heating cycle described above, and is closed otherwise. By controlling the on-off valves 28a and 28b as described above, the liquid refrigerant is branched on the downstream side of the supercooling heat exchanger 15 in each of the cooling cycle and the heating cycle, and is branched to the branch path 27. The previous total amount of liquid refrigerant is supercooled by the supercooling heat exchanger 15. Further, the controller 25 is connected to the engine 4 (control circuit thereof), and controls the operation of the main compressor 2 and the auxiliary compressor 3 by performing start / stop / control of the engine 4.
  In the above configuration, the controller 25 has the supercooling heat exchanger expansion valve 22 so that the wet refrigerant expanded by the supercooling heat exchanger expansion valve 22 is superheated in the path 33 that is the suction line of the auxiliary compressor 3. To control the opening degree. Then, as will be described later, by selecting (constructing) the auxiliary compressor 3, the refrigerant suction pressure of the auxiliary compressor 3 becomes higher than the refrigerant suction pressure of the main compressor 2, and is shown by the Mollier diagram of FIG. Thus, the compression work ΔWs by the auxiliary compressor 3 can be made smaller than the compression work ΔWm by the main compressor 2. In this way, the total compression work is reduced compared to the case where the entire amount of refrigerant is compressed with a single compression work ΔWm.
  Next, the Mollier diagram (FIG. 3) of the refrigeration cycle in the refrigerant circuit configuration as described above will be described according to the refrigerant flow in the refrigerant circuit configuration. In this Mollier diagram, the state change of the refrigerant per unit mass flow rate is represented, the horizontal axis indicates the specific enthalpy (kJ / kg) which is the energy per kg of the refrigerant mass, and the vertical axis Indicates (absolute) pressure (MPa abs).
  Regarding the refrigeration cycle on the Mollier diagram, the case of the cooling cycle will be described. A point Am in the Mollier diagram indicates a state in which the refrigerant flows through the path 32 constituting the suction line of the main compressor 2, and the specific enthalpy and the pressure value in this state are h2 (kJ / kg) and p2, respectively. (MPa abs). The flow rate of the refrigerant in the refrigerant circuit here is Gm. Point As indicates the state in which the refrigerant flows through the path 33 constituting the suction line of the auxiliary compressor 3, and the specific enthalpy and the pressure value in this state are h1 (kJ / kg) and p1 (MPa abs), respectively. ). And let Gs be the flow rate of the refrigerant in the refrigerant circuit here. The refrigerant in these states is sucked into the compressors 2 and 3 from the respective suction lines, and compression work is performed in the compressors 2 and 3. At this time, the main compressor 2 performs compression work ΔWm on the refrigerant per unit mass flow rate (compression section AmB), and the auxiliary compressor 3 performs compression work ΔWs on the refrigerant per unit mass flow rate. (Compression section AsB).
  The refrigerant (gas refrigerant) compressed to high pressure by the compressors 2 and 3 joins at a junction 65. Here, the flow rate of the combined refrigerant in the refrigerant circuit is defined as the total amount Go (= Gm + Gs). The merged refrigerant is sent to the outdoor heat exchanger 5. In the outdoor heat exchanger 5, heat is dissipated by condensing the refrigerant that has become high-pressure gas, and the refrigerant is cooled to become a liquid refrigerant (condensing section BC). That is, the state at point B indicates a state in which the refrigerant is in the path from the junction 65 to the outdoor heat exchanger 5, and the specific enthalpy value in this state is h0 (kJ / kg).
  The refrigerant sent out as the liquid refrigerant from the outdoor heat exchanger 5 is supercooled by the supercooling heat exchanger 15 by the supercooling liquid refrigerant branched into the branch path 27 a on the downstream side of the supercooling heat exchanger 15. (Supercooling section CD). Here, T1, T2, and T3 in the figure indicate isotherms (t1> t2> t3) of temperatures t1 (° C.), t2 (° C.), and t3 (° C.), respectively, and the liquid refrigerant flowing through the main path 26 Indicates that the supercooling heat exchanger 15 is supercooled from t1 (° C.) to t2 (° C.). Let the pressure value in the state at the point D of the liquid refrigerant after this supercooling be p0 (MPa abs).
  Then, the liquid refrigerant after being supercooled is partly branched in the main path 26 and then expanded by the indoor heat exchanger expansion valve 23, and the liquid refrigerant having a lower temperature and lower pressure than the indoor air for cooling. (Expansion section DEm). The pressure value at the point Em of the liquid refrigerant that has become low temperature and low pressure is defined as p2 (MPa abs). The liquid refrigerant in the state of point Em is sent to the indoor heat exchanger 8, where the refrigerant is evaporated by heat absorption from the indoor air (evaporation section EmAm). Then, the refrigerant that has become the gas refrigerant flows through the path 32 constituting the suction line of the main compressor 2 and is sucked into the main compressor 2 again. That is, here, the refrigerant pressure (value p2) in the evaporation section EmAm is equal to the refrigerant suction pressure Pm of the refrigerant of the main compressor 2 described above, and the flow rate of refrigerant sucked into the main compressor 2 in the refrigerant circuit is Gm. Become.
  On the other hand, the supercooled liquid refrigerant branched into the branch path 27a is expanded by the expansion valve 22 for the supercooling heat exchanger and has a lower pressure and temperature than the liquid refrigerant in the state at point C (expansion section DEs). . At this time, the supercooled liquid refrigerant is decreased from the temperature t2 (° C.) to t3 (° C.) of the liquid refrigerant after being supercooled by the expansion valve 22 for the supercooling heat exchanger. In this way, among the liquid refrigerant supercooled by the supercooling heat exchanger 15, the liquid refrigerant branched to the branch path 27a becomes the supercooling liquid refrigerant. And the flow volume in the refrigerant circuit of the liquid refrigerant branched to the branch path 27a becomes Gs. Here, the expansion of the branched liquid refrigerant (expansion section DEs) by the expansion valve 22 for the supercooling heat exchanger is suppressed more than the expansion of the liquid refrigerant (expansion section DEm) by the expansion valve 23 for the indoor heat exchanger. The reason is as follows. That is, in order to supercool the liquid refrigerant flowing through the main path 26 by the supercooling liquid refrigerant branched into the branch path 27a, the liquid refrigerant before the supercooling liquid refrigerant is sent to the supercooling heat exchanger 15 (point) C state), and the expansion of the supercooling liquid refrigerant in the supercooling heat exchanger expansion valve 22 is caused until the refrigerant pressure value p0 in the state of point D drops to the pressure value p1. This is because supercooling can be performed even if it is limited.
  Then, the supercooled liquid refrigerant in the state of point Es absorbs heat from the liquid refrigerant flowing through the main path 26 in the supercooling heat exchanger 15, thereby supercooling the liquid refrigerant flowing through the main path 26 (evaporation). Section EsAs). The refrigerant that has been supercooled flows through the path 33 that forms the suction line of the auxiliary compressor 3 and is sucked into the auxiliary compressor 3 again. Here, in the refrigerant circuit, the liquid refrigerant flowing through the main path 26 is partially branched (flow rate Gs) into the branch path 27a, and the flow rate Gm of the liquid refrigerant sent to the indoor heat exchanger 8 is reduced compared to the total amount Go. However, since the liquid refrigerant before branching is supercooled in the supercooling heat exchanger 15, the heat absorption capacity (cooling capacity) per unit mass flow rate (kJ / kg) of the liquid refrigerant is increased. The cooling capacity in the indoor heat exchanger 8 is maintained or improved.
  In this way, the expansion of the subcooled liquid refrigerant at the flow rate Gs branched into the branch path 27a by the subcooling heat exchanger expansion valve 22 is expanded by the expansion of the branch flow rate Gm of the liquid refrigerant by the indoor heat exchanger expansion valve 23. By suppressing the pressure drop of the supercooled liquid refrigerant from the pressure value p0 to the pressure value p1, the evaporation pressure in the evaporation section EsAs can be increased. That is, since the evaporation pressure of the refrigerant for subcooling at the flow rate Gs to be branched can be increased compared with the evaporation pressure of the refrigerant at the remaining flow rate Gm after the branching, the compression work ΔWs required in the compression section AsB is Compared with the required compression work ΔWm in the compression section AmB, it can be greatly reduced. Thereby, the compression work in the auxiliary compressor 3 can be significantly reduced as compared with the compression work in the main compressor 2, and the total compression work in the engine heat pump can be reduced.
  The specific amount of reduction in compression work is expressed as follows. The comparison target here is the total compression work in the case where the entire amount of refrigerant is compressed with a single compression work ΔWm. In other words, it is the total compression work in the case where the refrigerant of the total amount Go is compressed with the compression work ΔWm in the refrigerant circuit including the single compressor without including the auxiliary compressor. This is equivalent to the total compression work when the pressure drop in the expansion section DEs of the supercooled liquid refrigerant with the flow rate Gs branched to the branch path 27a is changed from the pressure value p0 to the pressure value p2. First, the total compression work in the case of compressing the refrigerant of the total amount Go, which is the comparison target here, with a single compression work ΔWm is Go × ΔWm = Go × (h0−h2) (Go: Gm + Gs). It is represented by (1). On the other hand, as described above, the compression work of the engine heat pump as a whole in the present invention keeps the pressure drop of the supercooled liquid refrigerant at the flow rate Gs branched to the branch path 27a from p0 to p1, so (Gm × ΔWm) + (Gs × ΔWs) = {Gm × (h0−h2)} + {Gs × (h0−h1)} (2) That is, the pressure drop of the supercooled liquid refrigerant at the flow rate Gs branched into the branch path 27a is kept from p0 to p1, and the amount of reduction in compression work by increasing the evaporation pressure of the refrigerant at the flow rate Gs is expressed by the above formula (1 ) And the expression (2), that is, compression work corresponding to Gs × (ΔWm−ΔWs) = (Gs × (h1−h2)) is reduced.
  As described above, the subcooling refrigerant having a higher evaporation pressure (the refrigerant suction pressure) than the refrigerant compressed by the main compressor 2 is compressed by the auxiliary compressor 3 driven by the engine 4. In addition, it is possible to reduce the total compression work in the refrigerant cycle without newly increasing the amount of electric power used for the auxiliary compressor, which was conventionally electrically driven, and to reduce the cooling by the supercooling action of the supercooling heat exchanger 15. Capability can be maintained or improved.
  Next, the capacity ratio between the main compressor 2 and the auxiliary compressor 3 in the engine heat pump according to the present invention will be described. The capacity ratio of the main compressor 2 and the auxiliary compressor 3 here is the ratio of the discharge capacities of the compressors 2 and 3, and the discharge capacities of the compressors 2 and 3 are the volume capacity and rotation of the compressors 2 and 3, respectively. Derived from the number. The volume capacity is the refrigerant suction volume (cc / cycle) per one cycle (one rotation) of the rotating body included in each of the compressors 2 and 3. The rotation speeds of the compressors 2 and 3 are such that the main compressor 2 and the auxiliary compressor 3 are driven by the common engine 4 as described above. It depends on the pulley ratio (transmission ratio) with respect to the engine pulley. From these things, the discharge capacity of each compressor 2 and 3 is calculated | required from the product of volume capacity | capacitance and pulley ratio, the volume capacity and pulley ratio of the main compressor 2 are set to Vm and Um, respectively, and auxiliary compressor 3's When the volume capacity and the pulley ratio are Vs and Us, respectively, the discharge capacity of the main compressor 2 is Vm × Um, and the discharge capacity of the auxiliary compressor 3 is Vs × Us. That is, the capacity ratio of the auxiliary compressor 3 to the total capacity (total discharge capacity) of the main compressor 2 and the auxiliary compressor 3 (hereinafter referred to as “auxiliary compressor capacity ratio R (%)”) is expressed by the following equation: R = (Vs × Us) / {(Vm × Um) + (Vs × Us)}. From this, the auxiliary compressor capacity ratio R is determined by the pulley ratios Um and Us for the respective engines 4 when the volume capacities Vm and Vs of the respective compressors 2 and 3 are equal, and the engine of each compressor 2 and 3 is determined. When the pulley ratios Um and Us with respect to 4 are equal, they are determined by the respective volume capacities Vm and Vs. In the present invention, the discharge capacity of the auxiliary compressor 3 is smaller than the discharge capacity of the main compressor 2. In the engine heat pump according to the present invention, the auxiliary compressor capacity ratio R (%) is configured from 20% to 29%. Hereinafter, the configuration of the auxiliary compressor capacity ratio R in the numerical range will be described.
  In the refrigerant circuit of the engine heat pump, the effect of the change in the auxiliary compressor capacity ratio R is that the supercooling liquid with the flow rate Gs branched in the main path 26 into the branch path 27a (during the cooling cycle) or 27b (during the heating cycle). That is, the ratio of the refrigerant to the total amount Go changes. That is, the ratio of the branched flow rate Gs to the total amount of liquid refrigerant Go increases as the auxiliary compressor capacity ratio R increases, and the ratio of the branched flow rate Gs to the total amount of liquid refrigerant Go decreases as the auxiliary compressor capacity ratio R decreases. Decrease. Based on this, the numerical range 20% to 29% of the auxiliary compressor capacity ratio R in the present invention will be described. In the following description, the supercooled liquid refrigerant (flow rate Gs) branched into the branch path 27a or 27b in the main path 26 is “branched liquid refrigerant”, and the liquid refrigerant (flow rate Gm) that flows through the main path 26 after branching. ) Is defined as “main liquid refrigerant”.
  First, regarding the numerical range 20% to 29% of the auxiliary compressor capacity ratio R, an explanation will be given of setting the upper limit value to 29%. The upper limit 29% of the auxiliary compressor capacity ratio R is derived from a change in operating efficiency (energy efficiency) during the cooling cycle (cooling). That is, at the time of cooling, by increasing the auxiliary compressor capacity ratio R, the flow rate Gs of the branch liquid refrigerant to the branch path 27a, that is, the supercooled liquid refrigerant that supercools the total amount of liquid refrigerant flowing through the main path 26. Therefore, the supercooling effect in the supercooling heat exchanger 15 is increased, and the cooling capacity per unit mass flow rate of the main liquid refrigerant is also increased. However, as the flow rate Gs of the branch liquid refrigerant increases, the flow rate Gm of the main liquid refrigerant decreases, and sufficient cooling capacity in the indoor heat exchanger 8 cannot be obtained. The upper limit value of the auxiliary compressor capacity ratio R is determined from a change in operating efficiency (energy efficiency) based on such a phenomenon.
  And in this invention, it is the graph shown in FIG. 4 which shows the specific measurement data used as the foundation about setting the upper limit of the auxiliary compressor capacity ratio R to 29%. In the graph shown in FIG. 4, the horizontal axis represents the auxiliary compressor capacity ratio R (%), and the vertical axis represents the coefficient of performance (COP) in the refrigerant cycle. The COP is expressed by cooling / heating capacity / fuel consumption, and the larger the COP value, the better the operation efficiency (energy efficiency). Moreover, the graph shown with a broken line shows COP in a refrigerant circuit structure in the case of providing a single compressor without providing an auxiliary compressor. As can be seen from this graph, the COP during cooling is flat at a higher value than in the case of a single compressor when the auxiliary compressor capacity ratio R is around 10%, but the auxiliary compressor capacity ratio R is 15%. As the auxiliary compressor capacity ratio R increases, the COP decreases. The COP at the time of cooling is lower than the COP in the case of a single compressor from the time when the auxiliary compressor capacity ratio R becomes about 30%. In other words, the value (about 30%) of the auxiliary compressor capacity ratio R at this point is a critical value (COP) that can improve the operating efficiency (COP) by reducing the total compression work during cooling in the present invention described above. If the auxiliary compressor capacity ratio R is less than about 30%, the COP at the time of cooling can maintain a high value as compared with the conventional case. Therefore, the upper limit value of the auxiliary compressor capacity ratio R in the present invention is set to 29%. As can be seen from the graph, the COP during the heating cycle always shows a higher value than before, regardless of the value of the auxiliary compressor capacity ratio R.
  Next, regarding the numerical range 20% to 29% of the auxiliary compressor capacity ratio R, setting the lower limit to 20% will be described. The lower limit value 20% of the auxiliary compressor capacity ratio R is the refrigerant temperature (hereinafter simply referred to as “inlet temperature”) at the connection point 15a serving as the refrigerant inlet on the main path 26 side of the supercooling heat exchanger 15 during the heating cycle (heating). And the refrigerant temperature at the connection point 15b serving as the refrigerant outlet on the main path 26 side of the supercooling heat exchanger 15 (hereinafter simply referred to as “outlet temperature”). That is, during heating, by reducing the auxiliary compressor capacity ratio R, the subcooling liquid that supercools the flow rate Gs of the branch liquid refrigerant that branches into the branch path 27b, that is, the total amount of liquid refrigerant that flows through the main path 26. Since the amount of the refrigerant is reduced, the supercooling action in the supercooling heat exchanger 15 is reduced, and the branch liquid refrigerant is easily evaporated. However, as the flow rate Gs of the branch liquid refrigerant decreases, the flow rate Gm of the main liquid refrigerant increases, so that the total amount of liquid refrigerant is not sufficiently subcooled by the subcooling heat exchanger 15 and the subcooling heat exchange is performed. In the vessel 15, the outlet temperature rises with respect to the substantially constant inlet temperature. Such an increase in the outlet temperature with respect to the inlet temperature in the supercooling heat exchanger 15 hinders obtaining a sufficient degree of supercooling in the supercooling heat exchanger 15 during heating. That is, in order to ensure the performance of the supercooling heat exchanger 15 during heating, a temperature difference of a certain level (for example, 5 ° C.) between the inlet temperature of the supercooled liquid refrigerant and the outlet temperature after supercooling. In other words, it is necessary to select (configure) the capacity of the auxiliary compressor 3 so that the degree of supercooling occurs. For this reason, the lower limit value of the auxiliary compressor capacity ratio R is determined.
  And in this invention, it is the graph shown in FIG. 5 which shows the specific measurement data used as the basis about setting the lower limit of auxiliary compressor capacity ratio R to 20%. In the graph shown in FIG. 5, the horizontal axis is the auxiliary compressor capacity ratio R (%), and the vertical axis is the inlet temperature or outlet temperature (° C.) of the supercooling heat exchanger 15. Yes. As can be seen from this graph, the inlet temperature of the subcooling heat exchanger 15 is substantially constant (32 to 33 ° C.) regardless of the value of the auxiliary compressor capacity ratio R. On the other hand, the outlet temperature of the supercooling heat exchanger 15 increases from a temperature lower than the inlet temperature to a higher temperature as the auxiliary compressor capacity ratio R decreases. That is, the outlet temperature becomes higher than the inlet temperature from the time when the auxiliary compressor capacity ratio R becomes a certain value. In the present invention, the relationship between the inlet temperature and the outlet temperature that can ensure the performance of the supercooling heat exchanger 15 during heating is preferably such that the outlet temperature is lower by about 5 ° C. or more than the inlet temperature. The critical value (lower limit value) of the auxiliary compressor capacity ratio R at which the outlet temperature is about 5 ° C. lower than the inlet temperature is 20%. For this reason, the lower limit value of the auxiliary compressor capacity ratio R in the present invention is set to 20%.
  As described above, the auxiliary compressor capacity ratio R in the engine heat pump according to the present invention is configured such that the numerical range is 20% to 29% from the upper limit value determined from the cooling time and the lower limit value determined from the heating time. As a result, the cooling capacity can be maintained or improved during cooling, and the performance of the supercooling heat exchanger 15 can be ensured during heating. In other words, in the configuration of the present invention in which the main compressor 2 and the auxiliary compressor 3 are driven by the common engine 4, the auxiliary compressor capacity ratio R is configured in the range of 20% to 29%, so that the cooling and heating can be performed. Operation with good operation efficiency (energy efficiency) at the time is possible.
  In the refrigerant circuit configuration of the engine heat pump according to the present invention, a continuously variable transmission (CVT) is adopted to transmit driving force from the engine 4 to the main compressor 2 and the auxiliary compressor 3. You can also. In this case, the gear ratios of the main compressor 2 and the auxiliary compressor 3 are changed by CVT in consideration of the critical value of the auxiliary compressor capacity ratio R during cooling and heating as described above.
  Specifically, in the engine heat pump according to the present invention, it is sufficient that the value of the auxiliary compressor capacity ratio R is smaller than the above-described upper limit value during cooling, and the value of the auxiliary compressor capacity ratio R is described above during heating. It only needs to be larger than the lower limit. That is, the CVT is controlled so that the auxiliary compressor capacity ratio R is less than about 30% during cooling, and the auxiliary compressor capacity ratio R is 20% or more during heating, and the speed is changed during cooling and heating. The ratio is changed. As described above, by using the CVT, the degree of freedom of the volume capacity Vs and the pulley ratio Us of the auxiliary compressor 3 set with respect to the volume capacity Vm and the pulley ratio Um of the main compressor 2 is improved. Can do. Further, only the upper limit value needs to be determined in the cooling cycle, and only the upper limit value needs to be determined in the heating cycle. Therefore, the auxiliary compressor capacity ratio R is more suitable for each of the cooling time and the heating time. The value can be set, and the operation efficiency (energy efficiency) in each cycle can be improved.
  Incidentally, in the engine heat pump according to the present invention, an engine waste heat recovery unit 6 is provided in parallel with the outdoor heat exchanger 5. The supercooled liquid refrigerant branched in the main path 26 is evaporated by the engine waste heat recovery unit 6 and compressed by the auxiliary compressor 3.
  As described above, the engine waste heat recovery unit 6 is for absorbing and evaporating the branched liquid refrigerant that has passed through the supercooling heat exchanger 15 during heating. In the engine waste heat recovery unit 6, The branch liquid refrigerant absorbs heat and evaporates by heat exchange between the branch liquid refrigerant and the engine coolant CW, which is higher in temperature than the branch liquid refrigerant.
  Next, regarding the refrigeration cycle on the Mollier diagram (FIG. 3), the case of the heating cycle will be described. In addition, the description which overlaps with the case of the cooling cycle mentioned above is abbreviate | omitted. First, the refrigerant (gas refrigerant) compressed to high pressure by the main compressor 2 and the auxiliary compressor 3 joins at a junction 65. The merged refrigerant is sent to the indoor heat exchanger 8. In the indoor heat exchanger 8, heat is dissipated by condensing the refrigerant that has become high-pressure gas, and the heat is dissipated into the room to be heated and cooled to become a liquid refrigerant (condensing section BC). That is, the state at the point B indicates a state where the refrigerant is on the path from the junction 65 to the indoor heat exchanger 8.
  The refrigerant sent out from the indoor heat exchanger 8 as the liquid refrigerant is supercooled by the supercooling heat exchanger 15 by the supercooling liquid refrigerant branched into the branch path 27 b on the downstream side of the supercooling heat exchanger 15. (Supercooling section CD).
  Then, after the liquid refrigerant after being supercooled is partially branched in the main path 26, the liquid refrigerant is expanded by the outdoor heat exchanger expansion valve 21 to become a low-temperature / low-pressure liquid refrigerant (expansion section DEm). The liquid refrigerant in the state of point Em is sent to the outdoor heat exchanger 5 where the refrigerant is evaporated by heat absorption from the outside air (evaporation section EmAm). Then, the refrigerant that has become the gas refrigerant flows through the path 32 constituting the suction line of the main compressor 2 and is sucked into the main compressor 2 again.
  On the other hand, the supercooled liquid refrigerant branched to the branch path 27b is expanded by the expansion valve 22 for the supercooling heat exchanger, and the pressure and temperature are lower than the liquid refrigerant in the state of point C (expansion section DEs). In this way, among the liquid refrigerant supercooled by the supercooling heat exchanger 15, the liquid refrigerant branched to the branch path 27b becomes the supercooling liquid refrigerant. And the flow volume in the refrigerant circuit of the liquid refrigerant branched to the branch path 27b becomes Gs.
  The supercooled liquid refrigerant in the state of point Es absorbs heat from the liquid refrigerant flowing through the main path 26 by the supercooling heat exchanger 15, thereby supercooling the liquid refrigerant flowing through the main path 26. The supercooled liquid refrigerant that has passed through the supercooling heat exchanger 15 is sent to the engine waste heat recovery unit 6. In the engine waste heat recovery unit 6, heat exchange between the supercooling liquid refrigerant and the engine cooling water CW is performed, and the supercooling liquid refrigerant absorbs heat and evaporates (evaporation section EsAs). The evaporated refrigerant flows through the path 33 constituting the suction line of the auxiliary compressor 3 and is sucked again into the auxiliary compressor 3.
  Thus, by performing supercooling even during heating, the operation efficiency (energy efficiency) is improved by the following action. The total amount of the liquid refrigerant flowing through the main path 26 is supercooled by the supercooling heat exchanger 15 as described above. Here, when the liquid refrigerant is supercooled, the heat absorption capacity (kJ / kg) per unit mass flow rate of the refrigerant is increased. That is, in the outdoor heat exchanger 5 after being supercooled, the heat absorption capacity from the outside air per unit mass flow rate of the liquid refrigerant is increased, and a small amount of liquid refrigerant is used as compared with the liquid refrigerant in the case of not being supercooled. It is possible to absorb the same amount of heat. Thereby, the flow rate Gm of the main liquid refrigerant sent to the outdoor heat exchanger 5 during heating can be reduced, and the total amount Go of the refrigerant circulating in the refrigerant cycle can be reduced. As a result, the total compression work in the refrigerant cycle can be reduced, and the operation efficiency (energy efficiency) can be improved.
  As described above, the engine waste heat recovery unit 6 is provided in parallel with the outdoor heat exchanger 5, and the supercooled branch liquid refrigerant is evaporated by the engine waste heat recovery unit 6 and compressed by the auxiliary compressor 3. As a result, the total compression work during cooling can be reduced by setting the auxiliary compressor capacity ratio R within the above-described range, and the total compression work can be reduced without increasing the amount of power used even during heating. Reduction can be achieved. Further, by supercooling the liquid refrigerant even during heating, the heat absorption capacity from the outside air per unit mass flow rate of the refrigerant is improved by the supercooling action, so that the total amount of refrigerant flowing through the refrigerant cycle can be reduced. . As a result, the total compression work can be reduced, and the operation efficiency (energy efficiency) can be improved.
  By the way, in the engine heat pump demonstrated above, it can also be set as the structure which drives independently the main compressor 2 and the auxiliary compressor 3 which are driven with the engine 4, respectively. By adopting such a configuration, it becomes possible to operate / stop the main compressor 2 and the auxiliary compressor 3 according to the magnitude of the air conditioning load, and it is possible to improve the operation efficiency (energy efficiency).
  In this case, as a specific configuration, as shown in FIG. 1, the driving force of the engine 4 is connected or disconnected between the engine 4 and the main compressor 2 and the auxiliary compressor 3 (switching between connected and unconnected). The main compressor clutch 42 and the auxiliary compressor clutch 43 are provided. The passage 32 constituting the suction line of the main compressor 2 and the passage 33 constituting the suction line of the auxiliary compressor 3 are communicated with each other by a communication route 34, and an opening / closing valve 35 is provided in the communication route 34. That is, by switching the opening / closing of the communication path 34 by opening / closing the opening / closing valve 35, the communication circuit 34 can be switched between communication / non-communication between the path 32 and the path 33, and the refrigerant circuit is reduced in air conditioning load. Operate in each load state corresponding to medium and high load state. Here, as shown in FIG. 2, the controller 25 described above is connected to the main compressor clutch 42 and the auxiliary compressor clutch 43, and the controller 25 receives each clutch from the engine 4 according to each load state. Controls connection / disconnection of driving force to / from. Similarly, the controller 25 is connected to the opening / closing valve 35 and controls the opening / closing of the opening / closing valve 35.
  With such a configuration, control according to each load state is performed as follows, for example, at the time of both cooling and heating. That is, during cooling, the auxiliary compressor 3 is operated alone when the air conditioning load is low, and the main compressor 2 is operated independently when the load is medium. When the load is high, the operation is performed by both the main compressor 2 and the auxiliary compressor 3 as described above, and the supercooling heat exchanger 15 performs supercooling. On the other hand, during heating, when the air conditioning load is low, the auxiliary compressor 3 is operated alone, and when the load is medium, the main compressor 2 is operated alone and the engine waste heat recovery unit 6 performs heat exchange. Do. In the case of a high load, the operation is performed by both the main compressor 2 and the auxiliary compressor 3 as described above, and the supercooling in the supercooling heat exchanger 15 and the heat exchange in the engine waste heat recovery unit 6 are performed. In addition, the level of the air-conditioning load referred to here is such that the range in which the air-conditioning load (%) of the engine heat pump is 0% to 15% is low, the range in which 15% to 60% is medium, and 60% to 60%. The range of 100% is a high load.
  First, the operation during cooling will be described. When the air conditioning load is low, the auxiliary compressor 3 is operated alone. In this case, the controller 25 disengages the main compressor clutch 42 and opens the on-off valve 35. That is, the driving force of the engine 4 is transmitted only to the auxiliary compressor 3 and the passage 32 that is the suction line of the main compressor 2 is connected to the passage 33 that is the suction line of the auxiliary compressor 3, so that the total amount of Go The refrigerant is compressed by the auxiliary compressor 3. In this case, whether or not the supercooling heat exchanger 15 performs supercooling is controlled by controlling the opening and closing of the supercooling heat exchanger expansion valve 22. When the supercooling by the supercooling heat exchanger 15 is performed, the controller 25 considers the pressure relationship in order to reduce the pressure loss at the junction 64 (FIG. 1), and the controller 25 determines the refrigerant pressure from the path 32 and the path. From 33, the opening degree of the expansion valve 22 for the supercooling heat exchanger and the expansion valve 23 for the indoor heat exchanger is controlled so that the refrigerant pressure becomes substantially the same.
  When the air conditioning load is medium load, the main compressor 2 is operated alone. In this case, the controller 25 turns off the auxiliary compressor clutch 43, transmits the driving force of the engine 4 only to the main compressor 2, and compresses the entire amount of refrigerant by the main compressor 2. In this case, when the supercooling by the supercooling heat exchanger 15 is performed, the controller 25 opens the on-off valve 35 and the refrigerant pressure from the path 32 and the refrigerant pressure from the path 33 at the junction 63 (FIG. 1). Are controlled so that the opening degree of the expansion valve 22 for the supercooling heat exchanger and the expansion valve 23 for the indoor heat exchanger is controlled to be substantially the same.
  When the air conditioning load is high, the operation is performed by both the main compressor 2 and the auxiliary compressor 3 and the supercooling heat exchanger 15 performs supercooling. In this case, the controller 25 turns on both the main compressor clutch 42 and the auxiliary compressor clutch 43 and closes the on-off valve 35. That is, the driving force of the engine 4 is transmitted to each of the compressors 2 and 3, and the communication between the passage 32 and the passage 33 is cut off, and the refrigerant of the flow rate Gm is compressed by the main compressor 2, and the supercooling of the flow rate Gs is performed. The refrigerant is compressed by the auxiliary compressor 3.
  Next, operation during heating will be described. When the air conditioning load is low, the auxiliary compressor 3 is operated alone. That is, in this case, the control mode by the controller 25 is the same as in the case of the low load in the cooling operation described above.
  When the air conditioning load is a medium load, the main compressor 2 is operated alone and the engine waste heat recovery unit 6 performs heat exchange. In this case, the controller 25 disengages the auxiliary compressor clutch 43 and opens the on-off valve 35. That is, the driving force of the engine 4 is transmitted only to the main compressor 2, and heat exchange is performed by the engine waste heat recovery unit 6, and the total amount of Go refrigerant that merges at the junction 63 is compressed by the main compressor 2. . In this case, when performing the supercooling by the supercooling heat exchanger 15, the controller 25 opens the on-off valve 35 and the refrigerant pressure from the path 32 and the refrigerant pressure from the path 33 become substantially the same at the junction 63. Thus, the opening degree of the expansion valve 22 for the supercooling heat exchanger and the expansion valve 21 for the outdoor heat exchanger is controlled.
  When the air conditioning load is high, the operation is performed by both the main compressor 2 and the auxiliary compressor 3, and the supercooling in the supercooling heat exchanger 15 and the heat exchange in the engine waste heat recovery unit 6 are performed. In this case, the controller 25 turns on both the main compressor clutch 42 and the auxiliary compressor clutch 43 and closes the on-off valve 35. That is, the driving force of the engine 4 is transmitted to each of the compressors 2 and 3, the communication between the path 32 and the path 33 is cut off, the refrigerant having a flow rate Gm is compressed by the main compressor 2, and the engine waste heat recovery unit 6 The sub-cooling refrigerant with the flow rate Gs to be heat-exchanged is compressed by the auxiliary compressor 3.
  Thus, the operation state at the partial load in which the combustion efficiency of the engine 4 is reduced by adopting a configuration in which the operation of the main compressor 2 and the auxiliary compressor 3 can be switched according to the level of the required air conditioning load. Therefore, it is possible to improve the operation efficiency (energy efficiency).
The refrigerant circuit figure of the engine heat pump which concerns on this invention. The block diagram of control equipments similarly. The Mollier diagram by a refrigerant circuit structure similarly. The graph which shows the relationship between auxiliary compressor capacity ratio and COP. The graph which shows the relationship between auxiliary compressor capacity ratio and a supercooling heat exchanger refrigerant temperature.
2 Main Compressor 3 Auxiliary Compressor 4 Engine 5 Outdoor Heat Exchanger 6 Engine Waste Heat Recovery Unit 8 Indoor Heat Exchanger 15 Supercooling Heat Exchanger 21 Outdoor Heat Exchanger Expansion Valve 22 Supercooling Heat Exchanger Expansion Valve 23 Indoor Heat Exchange expansion valve 26 Main path 27a Branch path 27b Branch path

Claims (2)

  1. Main compressor and auxiliary compressor driven by engine, indoor heat exchanger, outdoor heat exchanger, expansion valve for indoor heat exchanger, expansion valve for outdoor heat exchanger, and connection path between indoor heat exchanger and outdoor heat exchanger A subcooling heat exchanger that supercools the liquid refrigerant before branching by the subcooling liquid refrigerant that is provided in the liquid refrigerant passage path and branches into the branch path, and the refrigerant discharged from the auxiliary compressor is in the configuration and the engine heat pump is merged with the refrigerant discharged from the main compressor, the supercooling liquid refrigerant, said after passage of the supercooling heat exchanger, with a configuration that is compressed by the auxiliary compressor, an auxiliary compressor The main compressor and auxiliary compressor have a capacity ratio of 20% to 29% with respect to the total capacity of the main compressor and the auxiliary compressor, and the main compressor clutch and the auxiliary compressor are connected to the main compressor and the auxiliary compressor. Provide a clutch Engine heat pump, characterized in that.
  2.   The engine heat pump according to claim 1, wherein an engine waste heat recovery unit is provided in parallel with the outdoor heat exchanger, and the supercooled liquid refrigerant is evaporated by the engine waste heat recovery unit and compressed by an auxiliary compressor; An engine heat pump characterized by
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