WO2005114064A1

WO2005114064A1 - Engine heat pump

Info

Publication number: WO2005114064A1
Application number: PCT/JP2005/007411
Authority: WO
Inventors: Ken-Ichi Minami; Jirou Fukudome; Hiroshi Azuma; Keishi Yamanaka; Eita Kureha
Original assignee: Yanmar Co., Ltd.
Priority date: 2004-05-20
Filing date: 2005-04-18
Publication date: 2005-12-01
Also published as: EP1762792A4; CN1957211A; EP1762792A1; US20070295025A1; CN100470165C; JP4336619B2; JP2005331177A

Abstract

An engine heat pump capable of increasing operating efficiency (energy efficiency) by reducing compression work without increasing the use amount of electric power. The engine heat pump comprises a main compressor (2) and an auxiliary compressor (3) driven by an engine (4), an indoor heat exchanger (8), an outdoor heat exchanger (5), an expansion valve (23) for the indoor heat exchanger, an expansion valve (21) for the outdoor heat exchanger, and a supercooling heat exchanger (15) installed in a liquid refrigerant passing route (main route (26)) among connection routes for the indoor heat exchanger (8) and the outdoor heat exchanger (5) and supercooling a liquid refrigerant before branching by a refrigerant for supercooling branched to a branched route (27) (27a, 27b). The refrigerant discharged from the auxiliary compressor (3) is merged to the refrigerant discharged from the main compressor (2). The liquid refrigerant for supercooling is compressed by the auxiliary compressor (3), and the capacity ratio (auxiliary compressor capacity ratio R) of the capacity of the auxiliary compressor (3) to the total capacity of the main compressor (2) and the auxiliary compressor (3) is set to 20 to 29%.

Description

Specification

Engine heat pump

Technical field

The present invention relates to a device configuration of an engine heat pump, and more particularly, to a technique for reducing total compression work without newly increasing the amount of electric power used.

Background art

[0002] Regarding an engine heat pump configured to drive a compressor by an engine, a configuration described in Patent Document 1 is known. In Patent Document 1, the compression work of an engine heat pump is divided into two systems: compression work by a main compressor and compression work by an auxiliary compressor, and the evaporation pressure on one side (the auxiliary compressor side) is An invention is disclosed in which the compression work on one side is reduced by keeping the pressure higher than the evaporation pressure on the compressor side), thereby reducing the total compression work in the engine heat pump.

[0003] Patent Document 1 discloses a configuration in which compression work on the side where the evaporation pressure becomes high (auxiliary compressor side) is performed by an electrically driven compressor (electric compressor). In addition, equipment that requires new electric power (the electric compressor) will be added to the engine heat pump. In this case, although the compression work can be reduced, the amount of electric power used increases, resulting in "reduction of the amount of electric power used" and the full advantage of the engine heat pump.

Patent Document 1: JP 2004-20153

Disclosure of the invention

Problems to be solved by the invention

[0004] It is an object of the present invention to reduce the compression work without increasing the amount of electric power used in an engine heat pump, and to improve operating efficiency (energy efficiency).

Means for solving the problem

[0005] The engine heat pump of the present invention includes 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 indoor heat. The connection route between the heat exchanger and the outdoor heat exchanger is installed in the liquid refrigerant passage. A subcooling heat exchanger for subcooling the liquid refrigerant before branching with a subcooling liquid refrigerant branched to a branch path, and the refrigerant discharged from the auxiliary compressor is discharged from the main compressor. In the engine heat pump configured to be combined with the refrigerant, the supercooling liquid coolant is compressed by the auxiliary compressor, and the capacity of the auxiliary compressor with respect to the total capacity of the main compressor and the auxiliary compressor. The ratio is configured from 20% to 29%.

In the engine heat pump of the present invention, an engine waste heat recovery device is provided in parallel with the outdoor heat exchanger, and the supercooling liquid refrigerant is evaporated by the engine waste heat recovery device and compressed by an auxiliary compressor. The configuration is such that:

The invention's effect

[0007] In the engine heat pump of the present invention, the supercooling refrigerant having a higher evaporating pressure (refrigerant suction pressure) than the refrigerant compressed by the main compressor is compressed by the auxiliary compressor driven by the engine. By doing so, it is possible to reduce the total compression work in the refrigerant cycle without newly increasing the amount of power used by the auxiliary compressor, which was previously electrically driven, and to achieve the supercooling effect by supercooling heat exchange ^^ As a result, the cooling capacity can be maintained or improved.

In addition, by configuring the capacity ratio of the auxiliary compressor to the total capacity of the main compressor and the auxiliary compressor to be within a predetermined numerical range, the cooling capacity can be maintained or improved during cooling, and at the same time, heating can be performed. In this case, the performance of supercooling heat exchange can be secured. In other words, in the configuration of the present invention in which the main compressor and the auxiliary compressor are driven by the common engine, the operation efficiency (energy efficiency) during cooling and heating can be improved.

[0008] In the engine heat pump of the present invention, the total compression work during cooling is reduced by configuring the auxiliary compressor such that the capacity ratio of the auxiliary compressor to the total capacity of the main compressor and the auxiliary compressor is within a predetermined numerical range. In addition, it is possible to reduce the total compression work without newly increasing the power consumption even during heating.

In addition, by supercooling the liquid refrigerant even during heating, the supercooling effect increases the heat absorption capacity of the external force per unit mass flow rate of the refrigerant, thereby reducing the total amount of refrigerant flowing through the refrigerant cycle. Can be. As a result, the total compression work can be reduced Operation efficiency (energy efficiency) can be improved.

Brief Description of 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 the control devices.

FIG. 3 is a Mollier diagram similarly showing a refrigerant circuit configuration.

FIG. 4 is a graph showing the relationship between the auxiliary compressor capacity ratio and COP.

FIG. 5 is a graph showing a relationship between a capacity ratio of an auxiliary compressor and a refrigerant temperature of a subcooling heat exchanger.

Explanation of symbols

2 Main compressor

3 Auxiliary compressor

4 Engine

5 Outdoor heat exchange

6 Engine waste heat recovery unit

8 Indoor heat exchanger

15 Subcooling heat exchanger

21 Expansion valve for outdoor heat exchanger

22 Expansion valve for subcooling heat exchanger

23 Expansion valve for indoor heat exchanger

26 main route

27a Branch route

27b Branch route

BEST MODE FOR CARRYING OUT THE INVENTION

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 an engine 4, an indoor heat exchanger 8, an outdoor heat exchanger 5, an expansion valve 23 for an indoor heat exchanger, and an expansion for an outdoor heat exchanger. The valve 21 and the connection path between the indoor heat exchanger 8 and the outdoor heat exchanger 5 It has a subcooling heat exchanger 15 that is provided in the main path 26, which is a medium passage path, and supercools the liquid refrigerant before branching with the subcooling liquid refrigerant branched to the branch path 27 (27a, 27b). And a refrigerant cycle composed of these. The supercooled heat exchanger has connection points 15a and 15b with the main path 26 and connection points 15c and 15d with the branch path 27. In this configuration, a plurality of indoor heat exchanges 8 may be provided.

[0012] 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 by the four-way valve 24 in a predetermined direction. 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 forming a suction line of the main compressor 2. Have been.

The auxiliary compressor 3 is also driven by the engine 4, and is separated by an accumulator (not shown) from the subcooling liquid refrigerant that branches into the branch path 27 and passes through the supercooling heat exchange ^ 15. Suction of compressed gas refrigerant and discharge of high temperature and high pressure gas refrigerant

[0013] The subcooling heat exchange is for supercooling the liquid refrigerant before branching by the subcooling liquid refrigerant whose temperature has been lowered by the supercooling heat exchange expansion valve 22 provided in the branch path 27. The subcooling liquid refrigerant after heat exchange by heat exchange is sucked into the auxiliary compressor 3. For this reason, the supercooling heat exchange 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.

A branch path 27 provided in the main path 26 constitutes a branch path 27a between the indoor heat exchange 8 and the supercooled heat exchange 15, and a branch path 27a between the outdoor heat exchange 5 and the supercooled heat exchange. A branch path 27b is formed between the branch 15 and the inversion valve 15, and on-off valves 28a and 28b are provided between the branch paths 27a and 27b and the supercooled heat exchange expansion valve 22, respectively. The opening and closing of these opening / closing valves 28a 'and 28b is switched so that the liquid refrigerant before branching of the main path 26 is supercooled in a cooling cycle or a heating cycle described later.

[0015] The refrigerant discharged from the auxiliary compressor 3 is combined with the refrigerant discharged from the main compressor 2 at a junction 65 provided in a path from each of the compressors 2, 3 to the four-way valve 24. It is configured to join. Here, the flowing direction of the joined refrigerant is changed at the four-way valve 24, and will be described later. A cooling cycle or a heating cycle is performed. An oil separator (not shown) is provided between the junction 65 and the four-way valve 24 to separate the refrigerating machine oil contained in the high-temperature and high-pressure gas refrigerant to separate the main compressor 2 and the auxiliary compressor 3 from each other. It is recirculated to the suction side so that both compressors 2 and 3 can be lubricated well.

[0016] Using the refrigerant cycle configured as described above, the cooling cycle or the heating cycle is performed by switching the flow direction of the refrigerant by the four-way valve 24.

In the cooling cycle, the refrigerant compressed by the main compressor 2 and the auxiliary compressor 3 joins at the junction 65, is sent to the outdoor heat exchanger 5 through the four-way valve 24, and the outdoor heat exchanger After the heat is released at 5 and condensed, it is sent to the subcooling heat exchanger 15, flows in from the connection point 15b, and flows out from the connection point 15a. The liquid refrigerant supercooled in the subcooling heat exchanger 15 expands in the indoor heat exchanger expansion valve 23, absorbs heat in the indoor heat exchanger 8, evaporates, and then passes through the four-way valve 24 to the main compressor. Sucked in 2. After the drawn refrigerant is compressed by the main compressor 2, it is discharged again.

A part of the liquid refrigerant sent from the outdoor heat exchanger 5 and passing through the main path 26 is diverted to the branch path 27a as a liquid refrigerant for supercooling, and is divided by the expansion valve 22 for the subcooling heat exchanger. Expansion • The liquid refrigerant that flows through the main path 26 is supercooled in the process of flowing into the supercooled heat exchanger from the connection point 15c and flowing out to the connection point 15d to become a low-temperature wet refrigerant due to a decrease in temperature. At this time, the on-off valve 28a is in an open state and the on-off valve 28b is in a closed state, and the liquid refrigerant passing through the main path 26 is branched to the branch path 27a without being branched to the branch path 27b. The entire amount of liquid refrigerant before branching is supercooled by the subcooling liquid refrigerant. By supercooling the liquid refrigerant passing through the main path 26 in this way, the efficiency of the refrigeration cycle is improved. The subcooling liquid refrigerant is sucked into the auxiliary compressor 3, compressed by the auxiliary compressor 3, and discharged again.

[0018] On the other hand, in the heating cycle, the refrigerant compressed by the main compressor 2 and the auxiliary compressor 3 joins at a junction 65 and is sent to the indoor heat exchange 8 via the four-way valve 24. After being radiated and condensed in the indoor heat exchanger 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 and absorbs heat in the outdoor heat exchanger 5 to evaporate. Is sucked into the main compressor 2. Then, the drawn refrigerant is compressed by the main compressor 2 and then discharged again.

A part of the liquid refrigerant sent from the indoor heat exchange 8 and passing through the main path 26 is diverted to the branch path 27b as a liquid refrigerant for subcooling, and is divided by the expansion valve 22 for the subcooling heat exchanger. Expansion • The liquid refrigerant that flows through the main path 26 is supercooled in the process of flowing into the supercooled heat exchanger from the connection point 15c and flowing out to the connection point 15d to become a low-temperature wet refrigerant due to a decrease in temperature. 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 into a branch path 27b that is not diverted to the branch path 27a side. The entire amount of liquid refrigerant before branching is supercooled by the supercooling liquid refrigerant. The supercooling liquid refrigerant that has passed through the supercooling heat exchanger 5 absorbs heat in the engine waste heat recovery unit 6, evaporates, is sucked by the auxiliary compressor 3, and is compressed by the auxiliary compressor 3. Is discharged again.

Next, a device 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 outdoor heat exchanger expansion valve 21, the supercooling heat exchanger expansion valve 22, and the indoor heat exchanger expansion valve 23, The controller 25 controls the opening of each expansion valve.

[0021] Similarly, the controller 25 is connected to on-off valves 28a and 28b provided in the branch paths 27a and 27b, respectively, and controls the opening and closing of these valves. Here, the on-off valves 28a and 28b are specifically controlled as follows. That is, the on-off valve 28a is opened when the liquid refrigerant is supercooled in the cooling cycle described above, and is closed otherwise. Further, the opening / closing valve 28b is opened when the liquid refrigerant is supercooled in the above-described heating cycle, and is closed otherwise. By controlling the on-off valves 28a and 28b in this manner, the liquid refrigerant is branched off downstream of the subcooling heat exchanger 15 in each of the cooling cycle and the heating cycle, and The entire amount of liquid refrigerant before branching is supercooled by supercooling heat exchange 15.

Further, the controller 25 is connected to (the control circuit of) the engine 4 and controls the operation of the main compressor 2 and the auxiliary compressor 3 by performing start / stop control of the engine 4. [0022] In the above configuration, the controller 25 controls the subcooling heat exchanger so that the wet refrigerant expanded by the subcooling heat exchanger expansion valve 22 has a degree of superheating in the path 33 that is the suction line of the auxiliary compressor 3. The opening of the expansion valve 22 is controlled. By selecting (composing) the auxiliary compressor 3 as described later, the refrigerant suction pressure of the auxiliary compressor 3 becomes higher than the refrigerant suction pressure of the main compressor 2, and the Mollier diagram in FIG. As shown, the compression work AWs by the auxiliary compressor 3 can be made smaller than the compression work AWm by the main compressor 2. Thus, the total compression work is reduced as compared with the case where all the refrigerant is compressed by a single compression work AWm.

Next, a Mollier diagram (FIG. 3) of the refrigeration cycle in the above-described refrigerant circuit configuration will be described according to the flow of the refrigerant in the refrigerant circuit configuration. In this Mollier diagram, the state change of the refrigerant per unit mass flow rate is shown, and the horizontal axis shows the specific enthalpy (kj / kg), which is the energy per 1 kg of the mass of the refrigerant. The axis indicates (absolute) pressure (MPa abs).

With respect to the refrigeration cycle on the Mollier diagram, a description will be given of a case of a cooling cycle. A point Am in the Mollier diagram is a state where the refrigerant flows through the path 32 constituting the suction line of the main compressor 2. Where h2 (kj / kg) and p2 (MPa abs) are the specific enthalpy and pressure value in this state, respectively. Then, the flow rate of the refrigerant in the refrigerant circuit is Gm. The point As indicates the state in which the refrigerant is flowing 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 hi (kj / kg) and pi (MPa abs, respectively). ). Then, the flow rate of the refrigerant in the refrigerant circuit is set to Gs.

The refrigerant in these states is sucked into the respective compressors 2 and 3 at the respective suction line forces, and the respective compressors 2 and 3 perform compression work. At this time, the compression work AWm is performed on the refrigerant per unit mass flow rate in the main compressor 2 (compression section AmB), and the compression work AWs on the refrigerant per unit mass flow rate is performed in the auxiliary compressor 3. (Compression section AsB).

[0025] The refrigerant (gas refrigerant) which has been compressed by each of the compressors 2 and 3 to have a high pressure joins at a junction 65. To do. Here, the total flow rate of the combined refrigerant in the refrigerant circuit is defined as Go (= Gm + Gs). The combined refrigerant is sent to the outdoor heat exchanger 5. In the outdoor heat exchanger 5, heat is released by condensation of the refrigerant that has become high-pressure gas, and is cooled to become a liquid refrigerant (condensation 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 outdoor heat exchanger 5, and the value of the specific enthalpy in this state is defined as hO (kjZkg).

[0026] The refrigerant sent out as a liquid refrigerant from the outdoor heat exchanger 5 is supplied to the subcooling heat exchanger 15 at a downstream side of the subcooling heat exchanger 15 to a subcooling liquid refrigerant branched to a branch path 27a. Is supercooled (supercooling section CD). Here, Tl, Τ2, and Τ3 in the figure indicate isotherms (tl> t2> t3) of the temperatures tl (° C), t2 (° C), and t3 (° C), respectively, and indicate the main route 26. This indicates that the flowing liquid refrigerant is subcooled in the subcooling heat exchanger 15 from tl (° C) to t2 (° C). The pressure value of the liquid refrigerant after supercooling in the state at point D is defined as pO (MPa abs).

[0027] Then, the liquid refrigerant after being supercooled is partially branched in the main path 26 and then expanded by the indoor heat exchanger expansion valve 23 to have a lower temperature and a lower pressure than the indoor air for cooling. It becomes liquid refrigerant (expansion section DEm). Let _P 2 (MPa abs) be the pressure value of the liquid refrigerant at the low temperature and low pressure at the point Em. The liquid refrigerant in the state of the point Em is sent to the indoor heat exchanger 8, and in the indoor heat exchange 8, the refrigerant is evaporated by absorbing heat 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, the refrigerant pressure (value p2) in the evaporation section E mAm is equal to the refrigerant suction pressure Pm of the refrigerant of the main compressor 2 described above, and the flow force SGm of the refrigerant sucked into the main compressor 2 in the refrigerant circuit SGm It becomes.

On the other hand, the supercooling liquid refrigerant branched to the branch path 27a is expanded by the subcooling heat exchanger expansion valve 22 and has a pressure * temperature lower than that of the liquid refrigerant in the state at the point C ( Expansion zone DEs). At this time, the temperature of the liquid refrigerant for supercooling drops from the temperature t2 (° C) of the liquid refrigerant after supercooling described above by the supercooling heat exchange expansion valve 22 to t3 (° C). As described above, of the liquid refrigerant supercooled by the subcooling heat exchanger 15, the liquid refrigerant branched to the branch path 27a becomes the liquid refrigerant for supercooling. Then, the flow rate of the liquid refrigerant branched into the branch path 27a in the refrigerant circuit becomes Gs. Here, the expansion of the branched liquid refrigerant by the supercooling heat exchange expansion valve 22 (expansion section DEs) is suppressed more than the expansion of the liquid refrigerant by the indoor heat exchange expansion valve 23 (expansion section DEm). The reasons are as follows. That is, in order to supercool the liquid refrigerant flowing through the main path 26 by the subcooling liquid refrigerant branched to the branch path 27a, the liquid refrigerant (point) before the supercooling liquid refrigerant is sent to the subcooling heat exchanger 15 (State C), the expansion of the subcooling liquid refrigerant in the subcooling heat exchanger expansion valve 22 until the pressure value ρθ of the refrigerant at the state of point D drops to the pressure value pi. It is a cara that can perform supercooling even if stopped.

Then, the supercooling liquid refrigerant in the state at the point Es absorbs heat from the liquid refrigerant flowing through the main path 26 in the supercooling heat exchanger 15, thereby superposing the liquid refrigerant flowing through the main path 26. Cool (EsAs evaporation section). The refrigerant that has been supercooled flows through the path 33 constituting the suction line of the auxiliary compressor 3 and is sucked into the auxiliary compressor 3 again.

Here, in the refrigerant circuit, a part (flow rate Gs) of the liquid refrigerant flowing through the main path 26 is branched to 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 exchange 15, the heat absorption capacity (cooling capacity) (kj / kg) per unit mass flow rate of the liquid refrigerant increases. In addition, the cooling capacity of the indoor heat exchanger 8 is maintained or improved.

As described above, the expansion of the subcooling liquid refrigerant at the flow rate Gs branched to the branch path 27a by the expansion valve 22 for the subcooling heat exchanger is performed by the expansion valve for the indoor heat exchanger at the flow rate Gm of the branched refrigerant. By suppressing the pressure drop of the subcooling liquid refrigerant from the pressure value ρθ to the pressure value pi, the evaporation pressure in the evaporation section EsAs can be increased. In other words, the evaporating pressure of the subcooling refrigerant having the flow rate Gs to be branched can be increased as compared with the evaporating pressure of the refrigerant having the remaining flow rate Gm after branching. Ws can be significantly reduced compared to the required compression work AWm in the compression section AmB. As a result, 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.

[0031] A specific reduction amount of the compression work is expressed as follows. Note that the comparison here The elephant is the total compression work when all the Go refrigerant is compressed by a single compression work AWm. In other words, in a refrigerant circuit having a single compressor without an auxiliary compressor, it is the total compression work in the case where all the Go refrigerant is compressed by the compression work AWm. This is equivalent to the total compression work when the pressure drop in the expansion section DEs of the subcooling liquid refrigerant having the flow rate Gs branched into the branch path 27a is changed from the pressure value ρθ to the pressure value p2. First, the total compression work when compressing the entire amount of Go refrigerant to be compared with a single compression work AWm is Go X AWm = Go X (hO-h2) (Go: Gm + Gs) · · Represented by (1).

On the other hand, the compression work of the engine heat pump as a whole according to the present invention is, as described above, because the pressure drop of the subcooling liquid refrigerant having the flow rate Gs branched to the branch path 27a is limited from ρθ to pi. Is represented by (GmX AWm) + (Gs X AWs) = {Gm X (hO—h2)} + {Gs X (hO—1ι1)}... (2)

In other words, the pressure drop of the subcooling liquid refrigerant having the flow rate Gs branched into the branch path 27a is limited from ρθ to pi, and the amount of reduction of the compression work by increasing the evaporation pressure of the refrigerant having the flow rate Gs is expressed by the above equation The difference between (1) and (2), that is, Gs X (AWm-AWs) = (GsX (hi-h2)) compression work is reduced.

As described above, a configuration in which 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. As a result, it is possible to reduce the total compression work in the refrigerant cycle without newly increasing the amount of power used by the auxiliary compressor, which was previously electrically driven, and to supercool by the supercooling heat exchanger 15. By the action, the cooling capacity 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 between the main compressor 2 and the auxiliary compressor 3 is the ratio of the discharge capacity of each of the compressors 2 and 3, and the discharge capacity of each of the compressors 2 and 3 is the volume capacity and rotation of each. Derived from numbers. The volume capacity is the suction volume (ccZ cycle) of the refrigerant per cycle (1 rotation) of the rotating body provided in each of the compressors 2 and 3. The rotation speed of each of the compressors 2 and 3 depends on the main compressor 2 and the auxiliary compressor 3, which are driven by the common engine 4 as described above. Each of the compressor 2 and the auxiliary compressor 3 is determined by a pulley ratio (speed ratio) of the engine 4 to the engine pulley.

From these facts, the discharge capacity of each of the compressors 2 and 3 is determined by the product force of the volume capacity and the pulley ratio, and the volume capacity and the pulley ratio of the main compressor 2 are set to Vm and Um, respectively. Assuming that the volume capacity and the pulley ratio are Vs and Us, respectively, the discharge capacity of the main compressor 2 is VmX Um, and the discharge capacity of the auxiliary compressor 3 is Vs X 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 “auxiliary compressor capacity ratio R (%)” and! ヽぅ) is It is expressed by the following equation, R = (Vs X Us) / {(Vm X Um) + (Vs X Us)}. From this, the auxiliary compressor capacity ratio R is determined by the pulley ratio Um and Us for each engine 4 when the volume capacities Vm and Vs of the compressors 2 and 3 are equivalent, and the capacity of each compressor 2 and 3 When the pulley ratios Um and Us for the engine 4 are equal, they are determined by their 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 to be 20% to 29%. Hereinafter, the configuration of the auxiliary compressor capacity ratio R within the above numerical range will be described.

[0034] In the refrigerant circuit of the engine heat pump, the effect of the change in the auxiliary compressor capacity ratio R affects the flow rate of the main path 26 that is branched to the branch path 27a (during the cooling cycle) or 27b (during the heating cycle). This means that the ratio of Gs to the total amount of liquid refrigerant for supercooling, Go, changes. That is, when the auxiliary compressor capacity ratio R increases, the ratio of the branched flow rate Gs to the total amount of liquid refrigerant Go increases, and when the auxiliary compressor capacity ratio R decreases, the ratio of the branched flow rate Gs to the total amount of liquid refrigerant Go relative to Go Decrease.

Based on these facts, the numerical range of the auxiliary compressor capacity ratio R in the present invention in the range of 20% to 29% will be described. In the following description, the subcooling liquid refrigerant (flow rate Gs) branched to the branch path 27a or 27b in the main path 26 is referred to as a `` branch liquid refrigerant, '' and the liquid refrigerant flowing through the main path 26 after the branch (flow rate Gm) is defined and described as “main liquid refrigerant”.

[0035] First, regarding the numerical range of the auxiliary compressor capacity ratio R of 20% to 29%, setting the upper limit to 29% will be described. The upper limit of 29% of the auxiliary compressor capacity ratio R is derived from the changing power of the operating efficiency (energy efficiency) during the cooling cycle (cooling). In other words, during 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 supercooling liquid for supercooling the entire amount of liquid refrigerant flowing through the main path 26, Go. Since the amount of the refrigerant increases, the supercooling action in the supercooling heat exchanger 15 increases, and the cooling capacity per unit mass flow of the main liquid refrigerant also increases. However, as the flow rate Gs of the branch liquid refrigerant increases, the flow rate Gm of the main liquid refrigerant decreases, and it becomes impossible to obtain sufficient cooling capacity in the indoor heat exchanger 8. The upper limit of the auxiliary compressor capacity ratio R is determined from changes in operating efficiency (energy efficiency) based on these phenomena.

[0036] Fig. 4 is a graph showing specific measurement data as a basis for setting the upper limit of the auxiliary compressor capacity ratio R to 29% in the present invention.

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. This COP is represented by cooling / heating capacity Z fuel consumption, and the larger the value of COP, the higher the operating efficiency (energy-efficient). The graph indicated by the broken line shows the COP in the refrigerant circuit configuration when a single compressor is provided without the auxiliary compressor.

As can be seen from this graph, the COP during cooling becomes higher and flatter than the single compressor when the auxiliary compressor capacity ratio R is around 10%! From around the compressor capacity ratio R approaching 15%, the COP decreases as the auxiliary compressor capacity ratio R increases. Also, the instantaneous force at which the auxiliary compressor capacity ratio R becomes about 30% is lower than the COP for a single compressor during cooling. In other words, the value of the auxiliary compressor capacity ratio R at this point (approximately 30%) 1S The critical value (upper limit) at which the operation efficiency (COP) can be improved 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 during cooling can be maintained at a higher value than before. For this reason, the upper limit 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 of the auxiliary compressor capacity ratio R of 20% to 29%, the lower limit is set to 20%. This will be described.

The lower limit of 20% of the auxiliary compressor capacity ratio R is the refrigerant temperature at the connection point 15a, which is the refrigerant inlet on the main path 26 side of the supercooling heat exchange 15 during the heating cycle (heating) (hereinafter simply referred to as “ Temperature)) and the refrigerant temperature at the connection point 15b, which is the refrigerant outlet on the main path 26 side of the supercooling heat exchange 15 (hereinafter simply referred to as "outlet temperature"). In other words, during heating, by reducing the auxiliary compressor capacity ratio R, the flow rate Gs of the branched liquid refrigerant branched into the branch path 27b, that is, the supercooling for supercooling the total amount of liquid refrigerant flowing through the main path 26, Go. Since the amount of the liquid refrigerant is reduced, the supercooling effect in the supercooling heat exchange 5 is reduced, and the branched 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, and the liquid refrigerant of the total amount Go is not sufficiently supercooled by the supercooling heat exchanger 15, and becomes supercooled. In heat exchange 15, the outlet temperature rises with respect to the inlet temperature that is substantially constant. Such an increase in the outlet temperature with respect to the inlet temperature in the subcooling heat exchanger 15 hinders obtaining a sufficient degree of subcooling in the subcooling heat exchanger 15 during heating. In other words, in order to ensure the performance of the subcooling heat exchange 15 during heating, a temperature difference between the inlet temperature of the supercooled liquid refrigerant and the outlet temperature after the supercooling exceeds a certain level (for example, 5 °). C or more), that is, the capacity of the auxiliary compressor 3 needs to be selected (configured) so that the degree of supercooling occurs. Therefore, the lower limit of the auxiliary compressor capacity ratio R is determined.

FIG. 5 is a graph showing specific measurement data as a basis for setting the lower limit of the auxiliary compressor capacity ratio R to 20% in the present invention.

In the graph shown in Fig. 5, the horizontal axis represents the auxiliary compressor capacity ratio R (%), and the vertical axis represents the inlet or outlet temperature (° C) of the subcooling heat exchanger 15, and the respective values during heating are shown. Is shown.

As can be seen from this graph, the inlet temperature of the subcooling heat exchanger 15 is a substantially constant temperature (32 to 33 ° C) regardless of the value of the auxiliary compressor capacity ratio R. On the other hand, the outlet temperature of the subcooling heat exchanger 15 increases from a temperature lower than the inlet temperature to a higher temperature as the auxiliary compressor capacity ratio R decreases. In other words, the outlet temperature becomes higher than the inlet temperature from the point in time when the auxiliary compressor capacity ratio R reaches a certain value. And, in the present invention, when heating The relationship between the inlet temperature and the outlet temperature that can ensure the performance of supercooling heat exchange 15 is that the outlet temperature is preferably about 5 ° C or more lower than the inlet temperature. The critical value (lower limit) of the auxiliary compressor capacity ratio R, which is lower than the temperature by about 5 ° C or more, is 20%. For this reason, the lower limit of the auxiliary compressor capacity ratio R in the present invention is set to 20%.

As described above, the numerical value range of the auxiliary compressor capacity ratio R in the engine heat pump according to the present invention ranges from 20% to 29% from the upper limit determined from cooling and the lower limit determined from heating power. With this configuration, the cooling capacity can be maintained or improved during cooling, and the performance of supercooling heat exchange can be ensured during heating. That is, 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, by setting the auxiliary compressor capacity ratio R within the range of 20% to 29%, cooling and heating are performed. Operation with good operation efficiency (energy efficiency) at the time is possible.

[0040] In the refrigerant circuit configuration of the engine heat pump according to the present invention, a continuously variable transmission (CVT) is used to transmit the driving force from the engine 4 to the main compressor 2 and the auxiliary compressor 3. It can also be configured to be.

In this case, the gear ratio of the main compressor 2 and the auxiliary compressor 3 is changed by CVT in consideration of the critical value of the auxiliary compressor capacity ratio R at the time of cooling and at the time of heating as described above.

[0041] 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-mentioned upper limit during cooling, and the auxiliary compressor capacity ratio during heating. It is sufficient that the value of R is larger than the lower limit described above. That is, during cooling, the CVT is controlled so that the auxiliary compressor capacity ratio R is less than about 30%, and during heating, the auxiliary compressor capacity ratio is 20% or more. To change the gear ratio.

As described above, by using the CVT, the degree of freedom of the volume capacity Vs of the auxiliary compressor 3 and the pulley ratio Us set with respect to the volume capacity Vm of the main compressor 2 and the pulley ratio Um is improved. Can be done. In the cooling cycle, only the upper limit should be determined. In the heating cycle, only the lower limit needs to be determined. In each of heating and heating, the auxiliary compressor capacity ratio R can be set to a more suitable value, and the operation efficiency (energy efficiency) in each cycle can be improved.

Incidentally, in the engine heat pump according to the present invention, the engine waste heat recovery device 6 is provided in parallel with the outdoor heat exchange 5. The supercooling liquid refrigerant branched off 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 device 6 is for absorbing and evaporating the branched liquid refrigerant that has passed through the supercooling heat exchanger 15 during heating. In, the branch liquid refrigerant absorbs heat and evaporates by heat exchange between the branch liquid refrigerant and the engine cooling water CW having a higher temperature than the branch liquid refrigerant.

Next, the refrigeration cycle on the Mollier diagram (FIG. 3) will be described in the case of a heating cycle. The description of the same parts as those in the cooling cycle described above is omitted.

First, the refrigerant (gas refrigerant) compressed by the main compressor 2 and the auxiliary compressor 3 to have a high pressure joins at a junction 65. The combined refrigerant is sent to indoor heat exchange 8. In the indoor heat exchange 8, heat is dissipated by condensation of the refrigerant that has become high-pressure gas, and is dissipated in the room where heating is performed and cooled to become a liquid refrigerant (condensation 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.

[0045] The refrigerant sent out as liquid refrigerant from the indoor heat exchanger 8 is supplied to the subcooling heat exchanger 15 at a downstream side of the subcooling heat exchanger 15 to a subcooling liquid refrigerant branched to a branch path 27b. Is supercooled (supercooling section CD).

Then, the liquid refrigerant after being supercooled is partially branched in the main path 26 and then expanded by the outdoor heat exchange expansion valve 21 to become a low-temperature and low-pressure liquid refrigerant (expansion section). D Em). The liquid refrigerant in the state of the point Em is sent to the outdoor heat exchanger 5, where the refrigerant is evaporated by absorbing heat 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 subcooling liquid refrigerant branched to the branch path 27b is the expansion valve 2 for the subcooling heat exchanger. The pressure * temperature is lower than that of the liquid refrigerant in the state at the point C when expanded at 2 (expansion section DEs). As described above, of the liquid refrigerant supercooled by the subcooling heat exchanger 15, the liquid refrigerant branched to the branch path 27b becomes the subcooling liquid refrigerant. Then, the flow rate of the liquid refrigerant branched in the branch path 27b in the refrigerant circuit becomes Gs.

[0048] Then, the subcooling liquid refrigerant in the state of point Es absorbs heat from the liquid refrigerant flowing through main path 26 in subcooling heat exchanger 15, thereby superposing the liquid refrigerant flowing through main path 26. Cooling. The subcooling liquid refrigerant that has passed through the subcooling heat exchanger 15 is sent to the engine waste heat recovery unit 6. In this 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 forming the suction line of the auxiliary compressor 3 and is sucked into the auxiliary compressor 3 again.

[0049] By performing supercooling even during heating in this manner, the operation efficiency (energy efficiency) is improved by the following operation.

The liquid refrigerant of the total amount Go flowing through the main path 26 is supercooled by the supercooling heat exchange 15 as described above. Here, the subcooling of the liquid refrigerant increases the heat absorption capacity (kjZkg) per unit mass flow rate of the refrigerant. 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 smaller amount of the liquid refrigerant is used as compared with the liquid refrigerant that is not supercooled. It is possible to absorb the same amount of heat. Accordingly, 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 branch liquid refrigerant for supercooling is evaporated by the engine waste heat recovery unit 6 and compressed by the auxiliary compressor 3. With this configuration, by setting the auxiliary compressor capacity ratio R within the above range, it is possible to reduce the total compression work during cooling, and also to increase the amount of electric power used during heating without newly increasing Thus, the total compression work can be reduced.

Furthermore, even during heating, by supercooling the liquid refrigerant, Since the heat absorbing capacity of the external force per unit mass flow rate of the medium is improved, the total amount of the 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.

Meanwhile, in the engine heat pump described above, the main compressor 2 and the auxiliary compressor 3 driven by the engine 4 may be configured to be driven independently. With such a configuration, the main compressor 2 and the auxiliary compressor 3 can be operated and stopped according to the magnitude of the air conditioning load, and operation efficiency (energy efficiency) can be improved.

In this case, as a specific configuration, as shown in FIG. 1, connection and disconnection (connection / non-connection) of the driving force of the engine 4 between the engine 4 and the main compressor 2 and the auxiliary compressor 3 respectively. A clutch 42 for the main compressor and a clutch 43 for the auxiliary compressor for switching the connection) are provided.

Then, a path 32 forming a suction line of the main compressor 2 and a path 33 forming a suction line of the auxiliary compressor 3 are connected by a communication path 34, and an on-off valve 35 is provided in the communication path 34. In other words, by opening / closing the on-off valve 35 to switch the communication path 34 between open and non-open, the communication between the paths 32 and 33 can be switched between open and closed, and the refrigerant circuit is connected to the air conditioning load. Operate at each load condition corresponding to low, medium and high load conditions.

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. Control the connection and disconnection of the driving force to the motor. Similarly, the controller 25 is connected to the on-off valve 35 and controls opening and closing of the on-off valve 35.

With such a configuration, control according to each load state is performed, for example, as follows in each of cooling and heating. That is, during cooling, the auxiliary compressor 3 is operated independently when the air conditioning load is low, and the main compressor 2 is operated independently when the air conditioning 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 independently, and when the air conditioning load is medium, the main compressor 2 is operated independently, and heat is exchanged by the engine waste heat recovery unit 6. I do. And high In the case of a 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 exchange 15 and the heat exchange in the engine waste heat recovery unit 6 are performed.

The level of the air-conditioning load referred to here means that the air-conditioning load (%) of the engine heat pump is approximately 0% to 15%, low load, 15% to 60%, medium load, and 60% to 60%. High load is set in the range of 100%.

First, the operation during cooling will be described.

When the air conditioning load is low, the auxiliary compressor 3 is operated independently. In this case, the controller 25 sets the clutch 42 for the main compressor to the disengaged state and opens the on-off valve 35. In other words, by transmitting the driving force of the engine 4 only to the auxiliary compressor 3, and by connecting the path 32, which is the suction line of the main compressor 2, with the path 33, which is the suction line of the auxiliary compressor 3, the total amount is reduced. The Go refrigerant is compressed by the auxiliary compressor 3. Further, in this case, by controlling the opening and closing of the expansion valve 22 for the supercooling heat exchanger, it is controlled whether or not the supercooling heat exchanger 15 performs the supercooling. When supercooling is performed by the supercooling heat exchanger 15, the controller 25 considers the pressure relationship in order to reduce the pressure loss at the junction 64 (FIG. 1) and the like. The opening degree of the supercooling heat exchange expansion valve 22 and the indoor heat exchanger expansion valve 23 is controlled so that the refrigerant pressure from the path 33 becomes substantially the same.

When the air-conditioning load is a medium load, the main compressor 2 is operated independently. In this case, the controller 25 sets the clutch 43 for the auxiliary compressor to the disengaged state, transmits the driving force of the engine 4 to only the main compressor 2, and compresses the entire amount of the refrigerant in the main compressor 2. Further, in this case, when performing supercooling by the supercooling heat exchange 15, the controller 25 opens the on-off valve 35, and at the junction 63 (FIG. 1), the refrigerant pressure from the path 32 and the refrigerant pressure from the path 33. The opening degrees of the supercooling heat exchange expansion valve 22 and the indoor heat exchange expansion valve 23 are controlled so that the values are 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 is performed by the supercooling heat exchange 5. In this case, the controller 25 turns on the clutch 42 for the main compressor and the clutch 43 for the auxiliary compressor, and closes the on-off valve 35. That is, while transmitting the driving force of the engine 4 to each of the compressors 2 and 3, The communication between the path 32 and the path 33 is cut off, the refrigerant having the flow rate Gm is compressed by the main compressor 2, and the subcooling refrigerant having the flow rate Gs is compressed by the auxiliary compressor 3.

Next, the operation during heating will be described.

When the air conditioning load is low, the auxiliary compressor 3 is operated independently. 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 operation during cooling described above.

When the air conditioning load is a medium load, the main compressor 2 is operated independently, and heat is exchanged in the engine waste heat recovery unit 6. In this case, the controller 25 turns off the auxiliary compressor clutch 43 and opens the on-off valve 35. In other words, the driving force of the engine 4 is transmitted only to the main compressor 2 and heat is exchanged in 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. I do. In this case, when performing supercooling by the supercooling heat exchanger 15, the controller 25 opens the on-off valve 35, and at the junction 63, the refrigerant pressure from the path 32 and the refrigerant pressure from the path 33 are substantially the same. The opening degree of the expansion valve 22 for the supercooling heat exchanger and the expansion valve 21 for the outdoor heat exchanger are controlled so that

[0059] 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 exchange 15 and the heat in the engine waste heat recovery unit 6 are performed. Make a replacement. In this case, the controller 25 turns on the clutch 42 for the main compressor and the clutch 43 for the auxiliary compressor, and closes the on-off valve 35. In other words, 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 of Gm is compressed by the main compressor 2, and the engine waste heat collector 6 The auxiliary compressor 3 compresses a subcooling refrigerant having a flow rate of Gs, which is heat-exchanged at.

[0060] As described above, the configuration in which the operation of the main compressor 2 and the auxiliary compressor 3 can be switched in accordance with the level of the required air conditioning load allows the partial load in which the combustion efficiency of the engine 4 is reduced. Therefore, the operation state (energy efficiency) can be improved.

Industrial applicability

[0061] The engine heat pump of the present invention is widely applied to an engine heat pump having a configuration in which a compressor is driven by an engine, thereby reducing the compression work without increasing the amount of electric power used. The operation efficiency (energy efficiency) can be improved.

Claims

The scope of the claims

[1] main and auxiliary compressors driven by the engine, indoor heat exchangers, outdoor heat exchangers, expansion valves for indoor heat exchangers, expansion valves for outdoor heat exchangers, and indoor and outdoor heat exchangers A supercooling heat exchanger that is provided in the liquid refrigerant passage path of the connection path, and that supercools the liquid refrigerant before branching with the subcooling liquid refrigerant branched into the branch path, and is discharged from the auxiliary compressor. An engine heat pump configured to combine the refrigerant discharged from the main compressor with the refrigerant discharged from the main compressor.

The supercooling liquid refrigerant is compressed by an auxiliary compressor, and the capacity ratio of the auxiliary compressor to the total capacity of the main compressor and the auxiliary compressor is set to 20% to 29%. Engine heat pump.

[2] The engine heat pump according to claim 1!

An engine heat pump characterized in that an engine waste heat recovery device is provided in parallel with an outdoor heat exchanger, and the supercooling liquid refrigerant is evaporated by the engine waste heat recovery device and compressed by an auxiliary compressor.