JP2001322421A - Refrigerating cycle device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To simultaneously improve both of the heating performance by gas injection to a compressor and the oil returnability to the compressor, and simplify a decompression device. SOLUTION: The high pressure refrigerant after passed through an indoor heat exchanger 12 for heating the air in a heating mode, is partially by-passed to be reduced into the medium pressure by a first pressure reducing device 26 composed of a fixed restrictor 26b, the high-pressure refrigerant after passed through the indoor heat exchanger 12 and the medium-pressure refrigerant after passed through the first decompression device 26 are heat-exchanged by a refrigerant-refrigerant heat exchanger 23, and the medium-pressure gas refrigerant gasified by the heat exchanging is injected to the compressor 22. The high-pressure refrigerant after passed through the refrigerant-refrigerant heat exchanger 23 is sucked into the compressor 22 through a second decompression device 27 formed by an electric expansion valve, an exterior heat exchanger 24 and an accumulator 25.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a heat pump type refrigeration cycle apparatus for improving a heating capacity by gas injection, and is suitable for use in, for example, an air conditioner for an electric vehicle.

[0002]

2. Description of the Related Art Conventionally, vehicles such as electric vehicles cannot heat the interior of a vehicle cabin by using engine waste heat (hot water) as a heat source. The interior of the vehicle is heated by the heat of condensation.

However, in a use environment in which the outside air temperature drops to -10 ° C. or lower, such as in a cold region in winter, the amount of heat absorbed by an outdoor heat exchanger acting as an evaporator in a heat pump cycle decreases. Since the compressor suction pressure is reduced, the specific volume of the refrigerant is increased, and the amount of circulated refrigerant is reduced, resulting in a problem that the heating capacity is reduced. For this reason, in a cold region, the heating capacity in the vehicle compartment is insufficient.

Therefore, the present applicant has disclosed in Japanese Patent Application Laid-Open No. 9-395.
In Japanese Patent No. 50, during heating, the cycle high-pressure refrigerant is reduced to an intermediate pressure, the intermediate-pressure refrigerant is separated into a gas refrigerant and a liquid refrigerant by a gas-liquid separator, and the intermediate-pressure gas refrigerant is supplied to a compressor. There has been proposed a refrigeration cycle apparatus in which the injection work increases the compression work of the compressor during heating to increase the heating capacity.

[0005]

However, in the conventional apparatus described above, during heating, the intermediate-pressure gas refrigerant separated by the gas-liquid separator is reduced to a low pressure by a temperature type expansion valve, and this low-pressure refrigerant is exchanged for outdoor heat. The superheat degree of the refrigerant sucked into the compressor is adjusted by a temperature-type expansion valve.However, when the heating load is low in the interim period, the refrigerant flowing through the outdoor heat exchanger is reduced due to a decrease in the compressor rotation speed. The flow rate (flow rate) decreases, and oil in the refrigerant easily accumulates in the outdoor heat exchanger. As a result, the return of oil to the compressor may be deteriorated.

SUMMARY OF THE INVENTION In view of the foregoing, it is an object of the present invention to achieve both improvement in heating capacity by gas injection into a compressor and improvement in oil return to a compressor.

Another object of the present invention is to simplify a pressure reducing device in a heat pump type refrigeration cycle device for improving a heating capacity by gas injection.

[0008]

According to the first aspect of the present invention, in addition to the discharge port (22a) and the suction port (22b), the compressor (22) is provided with a refrigeration cycle intermediate pressure side. Compressor (2) having a gas injection port (22c) for introducing a gas refrigerant
(2) Indoor heat exchange in which air is heated by a high-pressure gas refrigerant discharged from a discharge port (22a) of a compressor (22) in a heating mode in an air conditioning passage (2) through which air flows toward a room. In the heating mode, a part of the high-pressure refrigerant after passing through the indoor heat exchanger (12) is bypassed, and the high-pressure refrigerant is depressurized to an intermediate pressure by the first pressure reducing device (26). And an indoor heat exchanger (1
2) The high-pressure refrigerant after passing and the intermediate-pressure refrigerant after passing through the first decompression device (26) are heat-exchanged by the refrigerant-refrigerant heat exchanger (23) and cooled by the refrigerant-refrigerant heat exchanger (23). The high-pressure refrigerant thus depressurized is reduced to a low pressure by the second decompression device (27), and the low-pressure refrigerant after passing through the second decompression device (27) is heat-exchanged with the outside air by the outdoor heat exchanger (24).

Then, in the heating mode, the outdoor heat exchanger (2)
The gas-liquid of the low-pressure refrigerant that has passed through 4) is accumulated in the accumulator (2).
5), the liquid refrigerant in which oil is dissolved and the low-pressure gas refrigerant are separated from the accumulator (25) by the compressor (2).
2), the intermediate-pressure gas refrigerant gasified by heat exchange in the refrigerant-refrigerant heat exchanger (23) is discharged to the suction port (22b) during the heating mode.
The gas is introduced into the gas injection port (22c) of 2). Further, the first pressure reducing device (26) is connected to the on-off valve (26a).
And a fixed throttle (26b) formed in the refrigerant passage in the on-off valve (26a).

According to this, the refrigerant-refrigerant heat exchanger (2)
By performing gas injection into the compressor (22) using the intermediate-pressure gas refrigerant gasified by the heat exchange in 3), the compression work of the compressor (22) during heating is increased, and heating is performed. Ability to increase.

Moreover, since gas injection is performed using the refrigerant-refrigerant heat exchanger (23), a gas-liquid separator for separating gas-liquid of the intermediate-pressure refrigerant as in the prior art is not required, and the compressor (22) The cycle low-pressure refrigerant can be sent from the accumulator (25) arranged on the suction side to the suction port (22b) of the compressor (22), and the liquid refrigerant in which oil is dissolved by the accumulator (25) is reliably converted into the low-pressure gas refrigerant. Can be mixed.

As a result, the flow rate (flow velocity) of the refrigerant flowing through the outdoor heat exchanger (24) is reduced due to a decrease in the number of rotations of the compressor, and oil in the refrigerant is easily accumulated in the outdoor heat exchanger (24). Even in such a case, good oil return to the compressor (22) can be ensured, which can contribute to improvement in durability of the compressor (22).

Further, in the heat pump type refrigeration cycle apparatus having gas injection combined with such a feature, the first pressure reducing device (26) can be simply constituted by a fixed throttle (26b) integrated with the on-off valve (26a). The cost can be reduced as compared with the electric expansion valve.

Further, even with the fixed throttle (26b), when gas injection is unnecessary, the flow of the refrigerant to the gas injection port (22c) can be shut off by the on-off valve (26a), and the gas injection function can be stopped at any time.

In addition, instead of the second pressure reducing device (27),
Since the first pressure reducing device (26) is constituted by the fixed throttle (26b), there are the following advantages. That is, since the second pressure reducing device (27) serves to reduce the pressure of the cycle main refrigerant passage, the second pressure reducing device (27) is constituted by a fixed throttle, and the throttle passage cross-sectional area is set to a normal (steady) operation. When designed to be compatible, the second pressure reducing device (2
7) The cross-sectional area of the throttle passage becomes too small. As a result, as illustrated in FIG. 14 to be described later, the compressor suction pressure sharply decreases at the time of starting heating, and the compressor is formed by forming the liquid refrigerant in the accumulator (25). The problem that the oil return to (22) deteriorates occurs. In addition, since the compressor protection circuit operates due to a sudden decrease in the compressor suction pressure and the compressor rotation speed is reduced, there is also a problem that the heating start-up effect at the time of starting heating is obstructed.

On the contrary, in the first aspect of the present invention,
Since the first pressure reducing device (26) that performs the pressure reducing function for the gas injection refrigerant is a fixed throttle, the second pressure reducing device (27) can be configured by a variable throttle such as an electric expansion valve. Thereby, the opening degree of the second decompression device (27) suitable at the time of starting heating can be set, and the above problem can be avoided.

According to the second aspect of the present invention, the on-off valve of the first pressure reducing device (26) is an electromagnetic valve (26a), which is fixed by the refrigerant passage immediately after the valve port (265) of the electromagnetic valve (26a). An aperture (26b) can be formed.

According to a third aspect of the present invention, in the heat pump type refrigeration cycle apparatus combined with gas injection having the same features as the first aspect, the first pressure reducing device (26)
Is constituted by a plurality of stages of fixed apertures (26c, 26d) that can be switched.

Thus, in addition to achieving the same effect as the first aspect, the first pressure reducing device (26) is constituted by a plurality of switchable fixed throttles (26c, 26d). The degree of opening of the fixed throttles (26c, 26d) of the stage can be switched according to the cycle conditions. Therefore, it is possible to exhibit a pressure reducing action (refrigerant flow rate adjusting action) similar to a variable throttle such as an electric expansion valve even though the fixed throttle has a simple structure.

As described in the fourth aspect of the invention, specifically, a plurality of fixed throttles (26c, 26c) are provided in accordance with the state of at least one of the refrigerants before and after the first pressure reducing device (26).
d) can be switched.

More specifically, a pressure response mechanism (268, 275) that responds to a differential pressure between the front and rear of the first pressure reducing device (26) is provided, as in the invention according to claim 5, wherein the differential pressure is a predetermined value. When it rises above the value, a plurality of fixed apertures (26c, 26d)
To increase the opening.

With this arrangement, the opening of the fixed throttles (26c, 26d) is increased when the load is high (when the outside air temperature is low) when the differential pressure across the first pressure reducing device (26) increases. In the case of a low heating load (high outside air temperature) in which the differential pressure across the first pressure reducing device (26) decreases, a plurality of fixed throttles (26c,
26d) can be reduced.

As described above, the cycle efficiency (COP) can be obtained by switching the fixed throttle opening in accordance with the heating load fluctuation.
Can be improved as shown in FIG.

According to the sixth aspect of the present invention, the second pressure reducing device (27) is an electric expansion valve, and the refrigerant-refrigerant heat exchanger (2)
The opening degree of the second pressure reducing device (27) is controlled according to the degree of supercooling of the high-pressure refrigerant cooled in 3).

[0025] Thus, the degree of supercooling of the high-pressure refrigerant can be constantly controlled to an appropriate value, which is advantageous for improving cycle efficiency.

According to the present invention, the second pressure reducing device (27) is an electric expansion valve, and the second pressure reducing device (27) is provided in accordance with the degree of superheat of the intermediate pressure gas refrigerant introduced into the gas injection port (22c). 27) The opening degree is controlled.

Thus, the degree of superheating of the gas injection refrigerant can be directly controlled to a value appropriate for the function of the compressor.

[0028] The invention described in claim 8 is characterized in that the second pressure reducing device (27) is an electric expansion valve, and the opening degree of the second pressure reducing device (27) is controlled according to the heating load.

Thus, the degree of opening of the second pressure reducing device (27) is controlled in accordance with the heating load, so that the degree of superheat of the gas injection refrigerant can be controlled to an appropriate value regardless of the heating load fluctuation.

According to a ninth aspect of the present invention, in any one of the sixth to eighth aspects, the opening degree of the second pressure reducing device (27) at the time of heating start is increased from the normal time.

As a result, it is possible to suppress a sudden drop in the compressor suction pressure at the time of heating startup, thereby suppressing a liquid refrigerant forming phenomenon inside the accumulator (25) caused by a sudden drop of the suction pressure, and thereby preventing the compressor from starting at the time of heating startup. Good return of oil to the compressor can be ensured. Further, it is possible to prevent the compressor protection circuit from being operated due to a sudden decrease in the compressor suction pressure, thereby preventing the compressor rotation speed from being reduced, and to improve the heating start-up effect at the time of startup.

Note that the reference numerals in parentheses of the above means indicate the correspondence with specific means described in the embodiments described later.

[0033]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS (First Embodiment) FIG. 1 shows a first embodiment in which the present invention is applied to an air conditioner for an electric vehicle, in which an air conditioning unit 1 is installed in a passenger compartment of an electric vehicle. The air-conditioning duct 2 forms an air-conditioning passage for guiding conditioned air into the vehicle interior. At one end of the air conditioning duct 2, suction ports 3, 4, and 5 for sucking inside and outside air are provided. The inside air intake port 4 and the outside air intake port 5 are connected to an inside / outside air switching door 6.
It is switched open and closed by.

A blower 7 for blowing air into the air-conditioning duct 2 is installed adjacent to the suction ports 3 to 5. The blower 7 includes a motor (not shown) and a fan 7a driven by the motor. 7b.

On the other hand, at the other end of the air-conditioning duct 2, there are a plurality of air outlets leading into the vehicle interior, that is, foot air outlets 8 for blowing the conditioned air toward the feet of the passengers in the vehicle interior, and toward the upper body of the passengers in the vehicle interior. A face outlet 9 for blowing out conditioned air and a defroster outlet 10 for blowing out conditioned air are formed on the inner surface of the vehicle windshield.

A cooling evaporator 11 is provided in the air conditioning duct 2 on the downstream side of the air from the blower 7. The cooling evaporator 11 is an indoor heat exchanger that constitutes a part of the refrigeration cycle 21. In a cooling operation and a dehumidifying operation mode, which will be described later, the air in the air conditioning duct 2 is absorbed by the refrigerant flowing through the inside. Functions as a cooler for dehumidification and cooling.

A heating condenser 12 is provided in the air conditioning duct 2 on the downstream side of the air from the cooling evaporator 11. The heating condenser 12 includes a refrigeration cycle 21.
And functions as a heater that heats the air in the air conditioning duct 2 by radiating the refrigerant flowing therein during a heating operation and a dehumidifying operation mode described later.

The air flow path in the air conditioning duct 2 is divided by a partition wall 13 into a first air flow path 14 on the foot outlet 8 side.
And a second air flow path 15 on the side of the face outlet 9 and the defroster outlet 10. This air channel 1
The two divisions 4 and 15 are for executing the next two-layer inside / outside air mode during heating in winter. That is, at the time of winter heating, the heating load is reduced by sucking high temperature inside air from the inside air inlet 3 into the first air flow path 14 on the side of the foot outlet 8 and blowing out the temperature to the feet, and at the same time, the defroster outlet 1
In order to implement the inside / outside air two-layer mode in which low humidity outside air is sucked into the second air passage 15 on the 0 side from the outside air suction port 5 and fogging of the front window is surely prevented, the air passage 1
4 and 15 divisions are performed.

The doors 16 and 17 are passage switching doors for switching between an air passage passing through the condenser 12 and a bypass passage 12a for bypassing the condenser 12, and one door 17 serves as a partition member for the air passages 14 and 15. Also serves as. Reference numeral 18 denotes a door disposed downstream of the air flow paths 14 and 15, and a partitioning function of the air flow paths 14 and 15 and the air flow path 1.
This is a door for switching between the communication states 4 and 15. In addition,
The outlets 8, 9, and 10 are opened and closed by an outlet switching door (not shown).

The refrigerating cycle 21 is configured as a heat pump type refrigerating cycle for performing cooling and heating with the evaporator 11 for cooling and the condenser 12 for heating. The following equipment is provided.

That is, the refrigerant compressor 22, the refrigerant-refrigerant heat exchanger 23 which exchanges heat between the gas-liquid two-phase intermediate-pressure refrigerant and the high-pressure refrigerant and gasifies, the outdoor heat exchanger 24, the cycle low-pressure refrigerant (compressor suction) Accumulator (gas-liquid separator) 25 for separating gas-liquid of refrigerant (coolant) and storing excess liquid refrigerant, condenser 12
A first decompression device 26 that bypasses a part of the high-pressure refrigerant after passing to reduce the pressure to an intermediate pressure, and a second pressure reduction device 2 that depressurizes the high-pressure refrigerant at the outlet of the refrigerant-refrigerant heat exchanger 23 to a low pressure during heating.
7. A third decompression device 29 for reducing the pressure of the high-pressure refrigerant condensed from the outdoor heat exchanger 24 to a low pressure during cooling, and electromagnetic valves (refrigerant path switching means) 28a and 28b for switching the refrigerant flow during cooling and heating are refrigerated. It is provided for cycle 21.

The outdoor heat exchanger 24 is installed outside the cabin of the electric vehicle, and exchanges heat with the outside air blown by the electric outdoor fan 24a. Further, the refrigerant compressor 22 is an electric compressor, and an AC motor (not shown) is built in a sealed case integrally, and driven by this motor to perform suction, compression and discharge of the refrigerant.

An AC voltage is applied to the AC motor of the refrigerant compressor 22 by an inverter 30, and the frequency of the AC voltage is adjusted by the inverter 30 to continuously change the motor rotation speed.
Therefore, the inverter 30 serves as a means for adjusting the rotation speed of the compressor 22, and a DC voltage from the vehicle-mounted battery 31 is applied to the inverter 30. Inverter 30
Is controlled by the air-conditioning control device 40.

The refrigerant compressor 22 is provided with a discharge port 22a for discharging the compressed refrigerant, a suction port 22b for sucking the refrigerant on the low pressure side of the cycle, and a gas injection port 22c for injecting the intermediate-pressure gas refrigerant. This gas injection port 22c
Communicates with the refrigerant-refrigerant heat exchanger 23 via the gas injection passage 22d.

The first pressure reducing device 26 is composed of an electromagnetic valve 26a as an on-off valve arranged on the upstream side and a fixed throttle 26b arranged on the downstream side. First decompression device 2
A specific configuration example of No. 6 will be described later with reference to FIG.

The second pressure reducing device 27 comprises an electric expansion valve whose valve opening is electrically adjusted. The electric expansion valve has an electric driving means such as a step motor, for example. The amount of displacement of the valve body is adjusted by using the valve body, and the opening degree of the refrigerant throttle passage is adjusted by the valve body. Further, the third decompression device 29 is a fixed throttle, and in the example shown in the drawing, is composed of a combination of an upstream capillary tube 29a and a downstream orifice 29b.

The accumulator 25 is a U-shaped refrigerant outlet pipe 2
5a, which stores the excess liquid refrigerant on the bottom side and sucks the gas refrigerant from the upper end opening of the U-shaped refrigerant outlet pipe 25a to prevent liquid back to the compressor 22. Also,
At the same time, the U-shaped refrigerant outlet pipe 25a of the accumulator 25
Through a small oil return hole (not shown) at the bottom of the
The liquid refrigerant in which the oil is partially dissolved is sucked in and mixed with the gas refrigerant, so that the oil return to the compressor 22 is ensured.

The refrigerant-refrigerant pipe 32 connecting the refrigerant-refrigerant heat exchanger 23 and the outdoor heat exchanger 24 has a refrigerant
Refrigerant temperature sensor 41a and high pressure sensor 41b for detecting the temperature and pressure of the high pressure refrigerant at the outlet of refrigerant heat exchanger 23
Is installed. The output signals of the sensors 41a and 41b are input to the air-conditioning control device 40, and by controlling the opening of the second pressure reducing device (electric expansion valve) 27, the subcooling of the high-pressure refrigerant at the outlet of the refrigerant-refrigerant heat exchanger 23 is performed. (The degree of supercooling).

The air-conditioning control device 40 comprises a microcomputer and its peripheral circuits. The air-conditioning control device 40 includes, in addition to the sensors 41a and 41b, an outside air temperature sensor 41c and air just after the evaporator. An evaporator temperature sensor 41d for detecting the temperature, a discharge temperature sensor 41e for detecting the discharge gas temperature of the compressor 22, a refrigerant temperature sensor 41h at the outlet of the outdoor heat exchanger 24, and a current sensor 41i for the inverter 30.
The sensor signal is input from a sensor group 41 such as.

A signal corresponding to the setting status of each lever (operating means) operated by the occupant (user) is also input to the air-conditioning control device 40 from the air-conditioning operation panel 50 (see FIG. 2). It has become.

FIG. 1 shows only the electrical connection between the inverter 30 and the air conditioning control device 40, and does not show the electrical connection between the other devices and the air conditioning control device 40. 1 electromagnetic valve 26a of pressure reducing device 26, second pressure reducing device 2
7, solenoid valves 28a, 28b, doors 6, 16, 17, 1,
8, the operation of the air outlet switching door (not shown), the blower 7, and the outdoor fan 24a are also controlled by the control device 40.
The solenoid valve 26a and the solenoid valves 28a and 28b
, The refrigerant circulation path is switched in accordance with each of the cooling, heating, and dehumidifying operation modes.

The air-conditioning operation panel 50 shown in FIG. 2 is provided with the following operation members manually operated by an occupant. Reference numeral 51 denotes a temperature setting lever for setting a target value of the temperature of the air blown into the vehicle interior. In this example, the temperature setting lever 51 is configured to set a target value for adjusting the rotation speed of the electric compressor 22.

Further, the opening and closing of the solenoid valves 28a, 28b and the passage switching doors 16, 17 is controlled with respect to a target value set by the operation position of the temperature setting lever 51, so that the operation mode of the refrigeration cycle is switched and the condenser 12 is operated. Or evaporator 1
The amount of heat exchange in step 1 is controlled.

For switching between the operation modes, for example, the cooling mode, the dehumidification mode, and the heating mode are sequentially set by moving the lever 51 from left to right as shown in FIG. Also, as shown in FIGS. 4, 5 and 6, by moving the operating position of the temperature setting lever 51, the target evaporator outlet temperature is set during cooling, and the target high pressure is set during dehumidification and heating. I have.

The operation position signal of the temperature setting lever 51 is inputted to the control device 40, and the control device 40 makes the actual evaporator blown air temperature or high pressure detected by the sensor group 41 coincide with the target value. Then, the rotation speed of the compressor 22 is controlled to control the temperature of the blown air.

Reference numeral 52 denotes a speed switching lever of the blower 7; 53, an air conditioner switch for interrupting the operation of the compressor 22; 54, an air conditioning blowout mode switching lever for opening and closing switching doors (not shown) of the outlets 8, 9, and 10; Reference numeral 55 denotes an inside / outside air switching lever for opening and closing the inside / outside switching door 6.

On the other hand, as shown in FIG. 7, the refrigerant-refrigerant heat exchanger 23 has an internal passage 23a and an external passage 2a.
3b is concentrically formed into a cylindrical shape having a double passage structure. The main passage refrigerant (high-pressure refrigerant) flowing toward the outdoor heat exchanger 24 is located at the center of the internal passage 23a. On the other hand, the external passage 23b is formed of a number of small passages arranged in parallel in the circumferential direction on the outer peripheral side of the internal passage 23a.
The intermediate-pressure refrigerant guided to the injection port 22c flows through d.

Here, the inner passage 23a and the outer passage 2
The tubular body 23c forming 3b is molded (eg, extruded) of a metal having excellent thermal conductivity such as aluminum.
Moreover, since the heat insulating material 23d is fixed to the outer surface of the tubular body 23c, it is possible to perform good heat exchange only between the high-pressure refrigerant in the internal passage 23a and the intermediate-pressure refrigerant in the external passage 23b. it can.

When the refrigerant-refrigerant heat exchanger 23 does not require gas injection, the first pressure reducing device 2
6 by closing the electromagnetic valve 26a of the internal passage 23.
Since the high-pressure refrigerant flows only in “a”, it is used as a part of the high-pressure side pipe 32.

FIG. 8 shows a specific example of the structure of the first pressure reducing device 26. Reference numeral 260 denotes a substantially cylindrical housing member constituting a refrigerant passage, on one end of which a high-pressure refrigerant from the outlet of the condenser 12 is supplied. A refrigerant inlet 261 for inflow is formed, and a refrigerant outlet 262 connected to the inlet side of the external passage 23b of the refrigerant-refrigerant heat exchanger 23 is formed on the other end side.

An electromagnetic valve 26a is formed at an axially intermediate portion of the housing member 260. This solenoid valve 26
a has a valve body 264 integrally attached to the tip of a plunger 263 made of a magnetic material, and the valve body 264 opens and closes a valve port 265 between the refrigerant inlet 261 and the refrigerant outlet 262. .

That is, when the electromagnetic coil 266b is energized, the fixed magnetic member 266a and the magnetic plunger 26
The magnetic flux flows through the magnetic circuit including the fixed magnetic member 26 and the fixed magnetic member 26.
The plunger 263 is sucked upward toward 6a, and the valve body 264 opens the valve port 265.

On the other hand, when the energization to the electromagnetic coil 266b is cut off, the plunger 263 is pushed down by the refrigerant pressure at the refrigerant inlet 261 side, that is, the cycle high pressure and the spring 266c.
4 is pressed against the end face (valve seat) of the valve port 265 to close the valve port 265. Immediately after the valve port 265, a fixed throttle 26b is formed by forming a small-diameter refrigerant passage having a smaller cross-sectional area than the valve port 265.

Next, the operation of the first embodiment in the above configuration will be described. When the air conditioner switch 53 is turned on, the signal is input to the control device 40 and the compressor 22 is started. In this state, the temperature setting lever 51 is set to P in FIG.
When the position is from H2 to PH1, the control device 40 determines that the heating mode is set, and sets the solenoid valves 26a, 28a and 28b and the passage switching doors 16 and 17 of the first pressure reducing device 26 during the heating operation in FIG. Control the state. Therefore, the passage switching doors 16 and 17 open the air passage on the side of the condenser 12, completely close the bypass passage 12 a, and fully open the air passage on the side of the condenser 12.

FIG. 1 shows the flow of the refrigerant in the heating mode.
In FIG. 1, the black arrow in FIG. 1 indicates the refrigerant flow path during the heating operation, and the high-temperature and high-pressure superheated gas refrigerant discharged from the compressor 22 is first supplied to the condenser set in the room. It flows into 12. Here, the high-pressure gas refrigerant exchanges heat (heat radiation) with the air blown by the blower 7 and condenses. The hot air heated by the condensation of the gas refrigerant is mainly blown out from the foot outlet 8 into the vehicle interior to heat the interior of the vehicle.

A part of the high-pressure two-phase refrigerant flowing out of the condenser 12 is bypassed to the injection passage 22d, flows into the first pressure reducing device 26, and is reduced to the intermediate pressure PM by the fixed throttle 26b. The two-phase refrigerant reduced to the intermediate pressure PM then passes through the external passage 23b of the refrigerant-refrigerant heat exchanger 23 and exchanges heat (heat absorption) with the high-pressure refrigerant at the outlet of the indoor condenser 12 passing through the internal passage 23a. It is gasified and flows into the injection port 22c.

On the other hand, the high-pressure refrigerant passing through the internal passage 23a of the refrigerant-refrigerant heat exchanger 23 exchanges heat (radiates heat) with the refrigerant passing through the external passage 23b, and is supercooled. Since the cooling electromagnetic valve 28b is closed during heating, the supercooled high-pressure refrigerant flows into the second pressure reducing device 27, is depressurized to the low pressure PL by the second pressure reducing device 27, and flows into the outdoor heat exchanger 24. When the low-pressure refrigerant passes through the outdoor heat exchanger 24, it absorbs heat from the air blown by the outdoor fan 24a (outside air) and evaporates.

The gas refrigerant evaporated in the outdoor heat exchanger 24 is
The liquid refrigerant that flows through the heating electromagnetic valve 28a and flows into the accumulator 25 and is generated by the fluctuation of the heating load is stored in the accumulator 25. In the accumulator 25, the gas refrigerant is sucked from the upper end opening of the U-shaped refrigerant outlet pipe 25a, and the oil partially dissolves through an oil return hole (not shown) provided at the bottom of the U-shaped refrigerant outlet pipe 25a. The liquid refrigerant is sucked and mixed with the gas refrigerant, and the gas refrigerant is supplied from the refrigerant suction passage 22f to the suction port 22 of the compressor 22.
b. This makes it possible to reliably return the oil to the compressor 22 even under a condition where the flow rate of the refrigerant is small as in the case of a low heating load in the intermediate period.

FIG. 10 is a Mollier diagram showing the state of the refrigerant in the refrigeration cycle during the heating operation described above.
Is the flow rate of the refrigerant gas injected from the injection passage 22d to the injection port 22c;
e is the flow rate of the refrigerant sucked into the compressor 22 through the outdoor heat exchanger (evaporator for heating) 24, Δi ₁ is the enthalpy difference of the intermediate pressure refrigerant on the gas injection side that absorbs heat in the refrigerant-refrigerant heat exchanger 23, Δi ₂ is the refrigerant-refrigerant heat exchanger 23
Is the enthalpy difference of the high-pressure refrigerant flowing toward the second decompression device 27.

In the conventional general heat pump system, when the outside air temperature decreases during heating, the suction pressure decreases and the specific volume of the refrigerant increases. Ability decreases. Further, since the compression ratio is increased by a decrease in the suction pressure, the discharge refrigerant temperature Td of the compressor 22 rises to T ₁ point of FIG. 10. For this reason, the compressor 22 cannot be used at the maximum capacity (maximum rotational speed) to protect the compressor 22.

On the other hand, in the first embodiment,
A part of the outlet refrigerant of the indoor condenser 12 is depressurized, heat exchanged in a refrigerant-refrigerant heat exchanger 23 and gasified, and the gas refrigerant is passed through a gas injection passage 22d to a compressor 22.
To return to the middle of the compression process (gas injection)
Compression work is performed by adding the refrigerant injection amount Gi to the refrigerant circulation amount Ge sucked by the compressor 22 and the gas injection amount Gi. As a result, the amount of compression work increases, and the amount of refrigerant radiated by the condenser 12 increases, so that the heating capacity can be improved.

At the same time, since the intermediate-pressure gas refrigerant is injected during the compression process of the compressor 22,
Gas refrigerant compressed heated halfway is cooled by the intermediate-pressure gas refrigerant, discharged refrigerant temperature Td drops to T ₂ points in FIG. 10. For this reason, it is possible to use the compressor 22 at the maximum capacity (maximum rotation speed). As described above, the heating capacity at the time of the low outside air temperature can be effectively improved.

Next, the temperature setting lever 51 is connected to the PC1 in FIG.
When the control device 40 is located at the position from to the PC2, the controller 40 determines that the air conditioner is in the cooling mode, and the solenoid valve 26a and the solenoid valve 2
8a, 28b and the air mix doors 16, 17 are shown in FIG.
Is controlled to the state of the cooling operation.

The refrigerant flow in the cooling mode will be described with reference to the cycle shown in FIG. 1. The outline arrows in FIG. 1 indicate the refrigerant flow path during the cooling operation, and the high-temperature and high-pressure superheated gas discharged from the compressor 22. The refrigerant first flows into the condenser 12 set in the room. However, the passage switching door 16,
Since 17 completely closes the air passage on the side of the condenser 12, the gas refrigerant does not exchange heat (dissipate heat) with the air blown by the blower 7 in the condenser 12.

Since all of the air blown by the blower 7 flows through the bypass passage 12a, the refrigerant discharged from the compressor 22 remains in a high-temperature, high-pressure superheated state in the refrigerant-refrigerant heat exchanger 2.
3 flows into the internal passage 23a.

At this time, since the solenoid valve 26a of the first pressure reducing device 26 installed in the injection passage 22d to the compressor 22 is controlled to be closed, the compressor 22
Is discharged from the refrigerant-refrigerant heat exchanger 23 without branching into the injection passage 22d.
Flows into the internal passage 23a. However, the external passage 23b
Since the intermediate-pressure low-temperature refrigerant does not flow, the refrigerant passing through the internal passage 22a is not cooled and flows out of the refrigerant-refrigerant heat exchanger 23 as a high-temperature and high-pressure superheated gas refrigerant, and the cooling electromagnetic valve 28b in the valve-open state is opened. The heat flows into the outdoor heat exchanger 24.

In this outdoor heat exchanger 24, the outdoor fan 2
The blown air (outside air) 4a and the high-pressure gas refrigerant exchange heat (dissipate heat) to condense the refrigerant. The refrigerant condensed in the outdoor heat exchanger 24 is discharged to the third position by closing the heating electromagnetic valve 28a.
After passing through the pressure reducing device 29, the pressure is reduced to the low pressure PL, and then flows into the evaporator 11.

In the evaporator 11, the refrigerant absorbs heat from the air blown by the blower 7 and evaporates. The cold air absorbed and cooled by the evaporator 11 is supplied to the downstream indoor condenser 1 as described above.
2 does not pass, but passes through the bypass passage 12a as cool air, and blows out mainly from the face outlet 9 into the vehicle compartment to cool the vehicle compartment.

On the other hand, the gas refrigerant evaporated in the evaporator 11 flows into the accumulator 25, from which the gas refrigerant passes through the refrigerant suction passage 22f to the compressor 22.
Is sucked into the suction port 22b.

Finally, the temperature setting lever 51 is moved to the PD in FIG.
When the position is between 1 and PD2, the control device 40 determines that the mode is the dehumidification mode, and sets the solenoid valves 26a, 28a and 28b and the passage switching doors 16 and 17 of the first decompression device 26 to the state during the dehumidification operation in FIG. To control.

The refrigerant flow in the dehumidification mode will be described with reference to the cycle of FIG. 1. The hatched arrow in FIG. 1 indicates the refrigerant flow path during the dehumidification operation, and the high-temperature and high-pressure superheated gas discharged from the compressor 22. The refrigerant is supplied to the passage switching door 16,
By the opening of 17, it flows into the condenser 12 set in the room,
Here, heat exchange (radiation) occurs with the air blown by the blower 7, and the gas refrigerant is condensed.

In the dehumidification mode, the first pressure reducing device 26 set in the injection passage 22 d to the compressor 22
Is controlled to be closed, and the refrigerant does not flow through the injection passage 22d, so that the entire amount of the high-pressure refrigerant condensed in the condenser 12 passes through the internal passage 23a of the refrigerant-refrigerant heat exchanger 23. At this time, the refrigerant passing through the internal passage 23a is not cooled, and passes through the refrigerant-refrigerant heat exchanger 23 as it is when it exits the indoor condenser 12. The high-pressure refrigerant flows into the second pressure reducing device 27 by closing the cooling electromagnetic valve 28 b, and is decompressed to an intermediate pressure by the second pressure reducing device 27 and flows into the outdoor heat exchanger 24.

Here, in the first dehumidification mode D ₁ in which a high blowing temperature is required in the dehumidification mode, the intermediate pressure created by the second pressure reducing device 27 is set lower than the saturation pressure of the refrigerant with respect to the outside air temperature. As shown in the Mollier diagram of FIG. 11, the outdoor heat exchanger 24 can be set to the heat absorbing side by acting as an evaporator. That is, the intermediate pressure is set low by reducing the opening of the second pressure reducing device 27 and increasing the amount of pressure reduction.

Then, the intermediate-pressure refrigerant flowing out of the outdoor heat exchanger 24 flows into the third pressure reducing device 29 by closing the heating electromagnetic valve 28a, and is reduced in pressure to the low pressure PL. The decompressed low-pressure refrigerant flows into the evaporator 11, absorbs heat from the air blown by the blower 7, evaporates, and then flows into the accumulator 25. The gas refrigerant from the accumulator 25 passes through the refrigerant suction passage 22f and is sucked into the suction port 22b of the compressor 22.

In the dehumidification mode, the refrigerant flows through both the evaporator 11 and the condenser 12 set in the indoor air conditioning unit 1, and the air blown by the blower 7 is first cooled by the evaporator 11.
After being dehumidified, it is reheated in the condenser 12 and becomes hot air. This warm air is mainly blown out from the differential air outlet 10 into the vehicle compartment to prevent fogging of the window glass and to dehumidify and heat the vehicle compartment.

[0086] Incidentally, in the dehumidification mode, the high mode D ₁ outlet temperature is first dehumidification necessary, as understood from the Mollier diagram of FIG. 11, the power L, the outdoor heat exchanger 24 of the compressor 22 And the sum of the heat absorption Qe of the indoor evaporator 11 and the sum of the heat absorption Qe of the indoor evaporator 11 can be radiated from the indoor condenser 12 (radiation Qc), so that a desired high blowing temperature can be created.

On the other hand, in the second dehumidifying mode D ₂ which requires a low blowing temperature in the dehumidifying mode, the intermediate pressure generated by the second pressure reducing device 27 must be set higher than the saturation pressure of the refrigerant with respect to the outside air temperature. As a result, as shown in the Mollier diagram of FIG. 12, the outdoor heat exchanger 24 can be set as a condenser by acting as a condenser. That is, the intermediate pressure is set high by increasing the valve opening of the second pressure reducing device 27 and reducing the amount of pressure reduction.

As described above, since the outdoor heat exchanger 24 functions as a condenser and acts as a heat radiating side, the sum of the power L of the compressor 22 and the heat absorption Qe of the indoor evaporator 11 and the discharge of the outdoor heat exchanger 24 Heat amount Qeh and heat release amount Q in indoor condenser 12
and the sum of c is equal. Thus, since the heat radiation amount Qc in the indoor condenser 12 is reduced than in the first dehumidification mode D _1, it is possible to create a lower air temperature of interest.

As described above, by controlling the intermediate pressure by adjusting the valve opening of the second pressure reducing device 27, the blowout temperature during dehumidification can be linearly controlled.

Next, a specific example of the opening control of the second pressure reducing device (electric expansion valve) 27 during the heating operation will be described. In the flowchart of FIG.
At 0, the operation mode of the air conditioner is determined based on the operation position of the temperature setting lever 51 and the ON / OFF of the air conditioner switch 53. Next, in step S101, it is determined whether the operation mode is the heating mode.

If the mode is the heating mode, step S102
And the elapsed time t from the start of the heating mode is the predetermined time t
It is determined whether 0 has been exceeded. If the elapsed time t is within the predetermined time t0, the process proceeds to step S102, and control is performed immediately after the second pressure reducing device (electric expansion valve) 27 starts heating. That is, the opening of the second pressure reducing device 27 is forcibly set to a predetermined large opening θ0 larger than the opening of the normal control, and the large opening θ
Maintain 0.

When the elapsed time t after the heating is started exceeds the predetermined time t0, the process proceeds to step S103, and the normal control of the opening of the second pressure reducing device 27 is performed. That is, the second pressure reducing device 2 is controlled so that the subcool SC of the high-pressure refrigerant that has exited the internal passage 23a of the refrigerant-refrigerant heat exchanger 23 becomes the predetermined value SCO.
7 is controlled.

That is, when the subcool SC of the high-pressure refrigerant becomes larger than the predetermined value SCO, the opening degree of the second pressure reducing device 27 is increased to lower the high pressure and decrease the subcool SC. Conversely, if the subcool SC of the high-pressure refrigerant is smaller than the predetermined value SCO, the opening degree of the second pressure reducing device 27 is reduced to increase the high pressure, and the subcool SC is increased.

In the cooling mode, the dehumidifying mode, and the air blowing (when the air conditioner switch 53 is turned off), step S105 is performed.
Performs control corresponding to each operation mode.

Next, the technical significance of the opening control of the second pressure reducing device 27 will be described. FIG. 14 shows that the compressor speed and the compressor suction pressure after the heating operation is started are reduced by the first and second pressure reducing devices. It is a characteristic diagram showing how it changes with the kind of combination of 26 and 27, A is the 1st decompression device 26: fixed throttling, 2nd decompression device 27: this 1st implementation called an electric expansion valve in the figure. Shows a combination of forms. B shows the combination of Comparative Example 1 in which the first pressure reducing device: the electric expansion valve and the second pressure reducing device 27: the fixed throttle, and C shows the combination of Comparative Example 2 in which both the first and second pressure reducing devices are the electric expansion valves. .

By the way, the second pressure reducing device 27 is installed in the main refrigerant passage in the cycle at the time of heating to reduce the pressure of the main refrigerant. In Comparative Example 1, the second pressure reducing device 27 is a fixed throttle. Here, since the passage cross-sectional area (throttle diameter) of the fixed throttle is designed so as to conform to the required flow rate characteristics during normal (steady) operation, it is necessary to rapidly increase the flow rate of the refrigerant as at the start of heating. Sometimes the cross sectional area of the passage of the fixed throttle is too small. As a result, in Comparative Example 1, FIG.
As described above, when the heating is started, the compressor suction pressure sharply decreases, so that a forming phenomenon in which the liquid refrigerant evaporates rapidly inside the accumulator 25 occurs.

The liquid refrigerant containing oil is taken out of the accumulator 25 at once by this forming phenomenon.
Thereafter, the oil return to the compressor 22 is temporarily stopped, causing poor lubrication of the compressor 22.

Further, the compressor protection circuit operates due to a sudden decrease in the compressor suction pressure at the time of starting heating, and the compressor speed is reduced as shown in FIG.
Therefore, there is also a problem that the heating start-up effect at the time of starting heating is hindered.

On the contrary, the combination A of the first embodiment
According to FIG. 13, since the second pressure reducing device 27 for reducing the flow of the main refrigerant in the cycle at the time of heating is an electric expansion valve,
, The opening of the second pressure reducing device 27 can be forcibly set to a predetermined large opening θ0 larger than the opening of the normal control within a predetermined time t0 after the heating is started. A large decrease in compressor suction pressure at the time of startup can be suppressed. Therefore, the forming phenomenon of the liquid refrigerant inside the accumulator 25 is suppressed, and the lubricity of the compressor 22 at the time of starting heating can be ensured.

Further, since the compressor suction pressure can be prevented from sharply decreasing at the time of heating startup, the compressor protection circuit can be prevented from operating at the time of heating startup, and the compressor speed can be maintained at a high speed. In addition, the heating start-up effect at the time of startup can be improved.

As compared with Comparative Example 2, in the combination A of the first embodiment, the cycle pressure reducing device can be simplified and the cost can be reduced by employing the fixed restrictor 26b in the first pressure reducing device 26.

Next, another effect of the first embodiment will be described.

In the prior art in which the low pressure refrigerant on the compressor suction side is superheated by a temperature expansion valve, the effect of improving the oil return property during a low heating load (compressor low speed range) Considering such a cycle balance at a low load of heating, oil is accumulated in the outdoor heat exchanger due to a decrease in the flow rate (flow velocity) of the low-pressure refrigerant passing through the outdoor heat exchanger due to a decrease in the number of rotations of the compressor. Oil may not be easily returned to the machine.

However, according to the first embodiment, since the refrigerant-refrigerant heat exchanger 23 does not require a gas-liquid separator for gas injection, the accumulator 25 is provided on the compressor suction side. , This accumulator 25
Thus, the superheat degree SH of the refrigerant drawn into the compressor can be controlled to be zero.
Then, the liquid refrigerant containing a certain amount of oil can be returned to the compressor 22 from the oil return hole of the outlet pipe 25a of the accumulator 25, so that the oil return to the compressor 22 can be maintained satisfactorily even at a low heating load. be able to.

Simplification of the cycle refrigerant circulation path In the cooling mode, the condenser 1
Since the air flow to 2 is cut off so that the air passes through the bypass passage 12a, the condenser 12 becomes a part of the refrigerant passage through which the high-pressure refrigerant flows. Therefore, heating, cooling,
Since the refrigerant continues to flow through the condenser 12 through all the modes of dehumidification, the refrigerant gas discharged from the compressor 22 can always flow in one direction toward the outdoor heat exchanger 24 through the condenser 12. As a result, it is possible to eliminate the four-way valve for reversing the refrigerant flow direction, or to reduce the number of valve devices such as the check valve for switching the refrigerant flow path and the solenoid valve, and to simplify the refrigerant piping configuration. .

(Second Embodiment) In the first embodiment, FIG.
Second pressure reducing device (electric expansion valve) in step S104
In the normal control of 27, the supercooling degree SC of the high-pressure refrigerant =
Although the opening degree of the second pressure reducing device 27 is controlled so as to reach the predetermined value SCO, in the second embodiment, as shown in FIG.
d so that the degree of superheat SH of the refrigerant of d is equal to the predetermined value SHO.
The opening degree of the pressure reducing device 27 is controlled.

For this reason, in the second embodiment, a refrigerant temperature sensor and a refrigerant pressure sensor are added to the gas injection passage 22d, and the superheat degree S of the refrigerant in the gas injection passage 22d is determined by the detected temperature and detected pressure of both sensors.
H is calculated, and the opening degree of the second pressure reducing device 27 is controlled so that the refrigerant superheat degree SH becomes a predetermined value SHO.

That is, when the refrigerant superheat degree SH becomes smaller than the predetermined value SHO, the opening degree of the second pressure reducing device 27 is increased. As a result, in the refrigerant-refrigerant heat exchanger 23, the flow rate of the high-pressure refrigerant on the internal passage 23a side increases, and the external passage 23b
Since the flow rate of the intermediate-pressure refrigerant on the side decreases, the degree of superheat SH of the refrigerant in the gas injection passage 22d increases.

On the contrary, the degree of superheat SH of the refrigerant is equal to the predetermined value SH.
When it becomes larger than O, the opening degree of the second pressure reducing device 27 is reduced. As a result, in the refrigerant-refrigerant heat exchanger 23, the flow rate of the high-pressure refrigerant on the side of the internal passage 23a decreases, and the external passage 2
Since the flow rate of the intermediate-pressure refrigerant on the 3b side increases, the degree of superheat SH of the refrigerant in the gas injection passage 22d decreases. In this manner, the degree of superheat SH of the refrigerant in the gas injection passage 22d can be maintained near the predetermined value SHO, and the function of gas injection into the compressor 22 can be exhibited well.

(Third Embodiment) In the third embodiment, FIG.
In step S104, the opening degree of the second pressure reducing device 27 is controlled in accordance with the change in the heating load.

More specifically, the heating load is determined by the outside air temperature sensor 41c.
Can be determined based on the outside air temperature detected by the above, and the heating load decreases as the outside air temperature increases. Then, when the heating load is small (when the outside air temperature is high), the amount of heat absorbed in the outdoor heat exchanger 24 increases and the low pressure increases, so that the intermediate pressure downstream of the first pressure reducing device 26 also increases. At this time, the compressor rotation speed is relatively low due to a small heating load, and the high pressure does not increase so much.

Therefore, in the refrigerant-refrigerant heat exchanger 23, the temperature difference between the high-pressure refrigerant and the intermediate-pressure gas injection refrigerant is reduced, and the exchange heat (radiation of the gas injection refrigerant) is reduced. May become too small, and the liquid refrigerant may return to the injection port of the compressor 22.

On the other hand, when the heating load is large (when the outside air temperature is low), the high pressure increases due to the increase in the rotation speed of the compressor 22, and the amount of heat absorbed in the outdoor heat exchanger 24 decreases to reduce the low pressure. Since the pressure decreases, the intermediate pressure downstream of the first pressure reducing device 26 also decreases. Therefore, in the refrigerant-refrigerant heat exchanger 23, the temperature difference between the high-pressure refrigerant and the intermediate-pressure gas injection refrigerant increases, and the amount of heat exchanged increases, so that the degree of superheat SH of the gas injection refrigerant becomes excessive.

Therefore, in the third embodiment, the opening of the second pressure reducing device 27 is increased at the time of a low heating load (when the outside air temperature rises). As a result, in the refrigerant-refrigerant heat exchanger 23, the flow rate of the high-pressure refrigerant on the internal passage 23a side increases, and the external passage 2
Since the intermediate-pressure refrigerant flow rate on the 3b side decreases, the degree of superheat SH of the gas injection refrigerant increases.

Conversely, the opening of the second pressure reducing device 27 is reduced when the heating is negative (when the outside air temperature is low). Thereby, in the refrigerant-refrigerant heat exchanger 23, the flow rate of the high-pressure refrigerant on the side of the internal passage 23a decreases, and the flow rate of the intermediate-pressure refrigerant on the side of the external passage 23b increases.
H can be reduced.

As described above, according to the third embodiment, the degree of superheat SH of the gas injection refrigerant can be maintained in an appropriate range by controlling the opening degree of the second pressure reducing device 27 according to the change in the heating load. In addition, the function of gas injection into the compressor 22 can be favorably exhibited.

In the third embodiment, the case where the heating load is determined based on the outside air temperature has been described.
The heating load may be determined based on the refrigerant pressure or the refrigerant temperature on the high pressure side of the cycle or the refrigerant pressure or the refrigerant temperature on the low pressure side of the cycle.

(Fourth Embodiment) FIG. 17 shows a fourth embodiment. In the first embodiment, the gas injection passage 22 is used.
Although the first decompression device 26 provided in d is constituted by the fixed throttle 26b formed in the refrigerant passage in the solenoid valve 26, in the fourth embodiment, the first decompression device 26 is formed separately from the solenoid valve 26a. The first and second fixed apertures 26c and 26d are provided.

In FIG. 17, reference numeral 267 denotes a cylindrical housing member constituting a refrigerant passage on the downstream side of the electromagnetic valve 26a, and a valve body 268 and a holding member 269 are disposed inside the cylindrical housing member. This holding member 269 is also a cylindrical member,
It is fixed to the inner wall of the housing member 267 by press fitting or the like. An O-ring 270 seals between the holding member 269 and the inner wall of the housing member 267.

The valve body 268 has a stepped cylindrical shape in which an upstream small-diameter portion 268a and a downstream large-diameter portion 268b are integrally formed, and is held inside a holding member 269 so as to be displaceable in the axial direction. The first fixed throttle 26c is provided at the center of the valve body 268.
Is formed. The first fixed throttle 26c is connected to the valve element 268.
And a passage hole having a small cross-sectional area (small diameter) penetrating the entire length in the axial direction. Also, an O-ring 271 is disposed near the most downstream end of the large diameter portion 268b of the valve body 268,
The space between the holding member 68 and the holding member 269 is sealed.

On the other hand, of the outer peripheral surface of the holding member 269, O
The housing member 26 is located downstream of the ring 270 in the axial direction.
7 is formed with a stepped small diameter portion 273 for forming a gap 272, and a second fixed throttle 26 d is formed at the portion of the stepped small diameter portion 273 so as to communicate with the gap 272. is there.

A radial communication hole 268c is formed in the axially intermediate portion of the large diameter portion 268b of the valve body 268, and when the valve body 268 is displaced to the position shown in FIG.
The fixed aperture 26d is configured to communicate with an intermediate portion of the first fixed aperture 26c via the communication hole 268c.

A spring holding plate 274 is fixed in the vicinity of the most upstream end of the small diameter portion 268a of the valve body 268 by a screw or the like so as to be adjustable in position, and elasticity is provided between the spring holding plate 274 and the upstream end surface of the holding member 269. Compression coil spring 275 as means
Is arranged.

When the load of the heating is low, the high pressure PH on the upstream side of the valve 268 decreases and the low pressure increases due to the decrease in the rotation speed of the compressor 22.
The intermediate pressure PM on the downstream side of is also increased. Therefore, the differential pressure ΔP between the high pressure PH and the intermediate pressure PM becomes smaller than the set pressure P0 set by the mounting load of the spring 275.

As a result, the valve body 268 contacts the stopper 276 of the holding member 269 as shown in FIG. 17A by the spring force of the spring 275, and the second fixed throttle 26d is closed. Therefore, at the time of heating low load, the refrigerant flows through only the first fixed throttle 26c in the first pressure reducing device 26 as shown by the arrow a.

On the other hand, when the heating load is high, the compressor 22
Since the high pressure PH rises and the low pressure falls due to the increase in the rotation speed of, the intermediate pressure PM also falls. Therefore, the differential pressure ΔP between the high pressure PH and the intermediate pressure PM becomes larger than the set pressure P0 set by the mounting load of the spring 275. As a result, the valve element 268 moves downstream by overcoming the spring force of the spring 275, and thereby the valve element 268
The communication hole 268c of the holding member 26 as shown in FIG.
9 communicates with the second fixed aperture 26d.

Therefore, when the heating is under a high load, the refrigerant flows through both the first fixed throttle 26c and the second fixed throttle 26d as indicated by arrows a and b in the first pressure reducing device 26. Can be increased more than when the heating load is low. That is, in the fourth embodiment, the first and second fixed throttles 26c and 26d constitute a plurality of stages of fixed throttles that can be switched by the differential pressure ΔP, and the following effects can be exerted.

FIG. 18 shows the external temperature indicating the heating load on the horizontal axis, and the vertical axis shows the degree of supercooling SC of the high-pressure refrigerant at the inlet side of the second pressure reducing device 27 and the COP (coefficient of performance) of the heat pump. . In the drawing, A indicates a combination of the first embodiment described above, in which the first pressure reducing device 26: fixed throttle, and the second pressure reducing device 27: electric expansion valve. B shows a combination of the fourth embodiment in which the first pressure reducing device 26 is a switching type two-stage fixed throttle and the second pressure reducing device is an electric expansion valve. C shows a combination of a comparative example in which both the first and second pressure reducing devices are electric expansion valves.

FIG. 18 shows that the supercooling degrees SC and COP of the combination A of the first embodiment and the combination B of the fourth embodiment are obtained when the outside air temperature is −20 ° C. (when the heating is under a high load).
This is a characteristic in a case where the fixed throttle opening characteristics of the combinations A and B are set so as to be the same as the comparative example C. According to the combination A of the fourth embodiment, the outside air temperature = 0 ° C. (at the time of a low heating load). Thus, the cycle efficiency (COP) can be improved as compared with the combination A of the first embodiment, and good characteristics closer to those of the comparative example C can be obtained.

Next, the reason why the cycle efficiency (COP) is improved by the combination A of the fourth embodiment will be described in comparison with the first embodiment. In the first embodiment, the opening degree of the one-stage fixed throttle 26b is set to be smaller. If the heating pressure is set to match the high load (low outside temperature), the intermediate pressure PM tends to increase due to the increase of the cycle low pressure when the heating is low (high outside temperature). In addition, the opening degree of the fixed throttle 26b becomes excessive under a low load condition, which further promotes an increase in the intermediate pressure PM.

The broken line in FIG. 19 (b) is a Mollier diagram in the case where the first pressure reducing device 26 is composed of only one stage fixed throttle 26b as described above. When the pressure PM rises, the refrigerant-refrigerant heat exchanger 2
3: decrease of temperature difference ΔT → decrease of heat exchange → supercooling degree S
C (reduction of the degree of superheat SH of the gas injection refrigerant) is caused, and as a result, COP is reduced.

On the other hand, according to the fourth embodiment, when the heating load is high (at low outside air temperature), the high pressure PH and the intermediate pressure PM
As a result, the first fixed throttle 26c and the second fixed throttle 26d of the first pressure reducing device 26 are both opened, and the opening degree of the first pressure reducing device 26 is increased. On the other hand, when the heating load is low (high outside air temperature), the first difference is caused by the decrease in the differential pressure ΔP.
Of the second fixed apertures 26c and 26d, the second fixed aperture 26
d is in the closed state, and the opening of the first pressure reducing device 26 can be reduced to a level suitable for a low heating load.

As a result, as shown by the solid line in FIG. 19B, in the fourth embodiment, the intermediate pressure PM at the time of the low heating load (at the time of the high outside air temperature) is changed to the first pressure reducing device 26 by the one-stage fixed throttle. 26
b, the degree of supercooling S
C can be suppressed from decreasing. Therefore, the COP at the time of a low heating load (at the time of high outside air temperature) can be increased.

In the fourth embodiment, the first pressure reducing device 2
Although the first pressure reducing device 26 is composed of three or more switchable fixed throttles, the first pressure reducing device 26 may be configured with two switchable fixed throttles 26c and 26d.

Further, in the fourth embodiment, the first pressure reducing device 2
6, the two-stage fixed throttles 26c and 26 are switched by a pressure response mechanism (constant differential pressure valve mechanism) that responds to the differential pressure ΔP. A temperature-sensitive mechanism (for example, a mechanism using a temperature-sensitive member such as a bimetal or a shape memory alloy) that responds to the intermediate-pressure refrigerant temperature is provided. May be switched.

Further, the fixed throttles 26c and 26 of the plurality of stages may be switched according to the state of only one of the high-pressure refrigerant and the intermediate-pressure refrigerant.

Also, in the fourth embodiment, the control of FIGS. 13, 15 and 16 can be similarly applied to the opening control of the second pressure reducing device 27 composed of an electric expansion valve.
Thereby, the effect of the control of FIGS. 13, 15, and 16 can be exerted also in the fourth embodiment.

[Brief description of the drawings]

FIG. 1 is a refrigeration cycle diagram showing a first embodiment of the present invention.

FIG. 2 is a front view of an air conditioning control panel used in the first embodiment.

3 is a characteristic diagram of an operation area and an operation mode of a temperature control lever in the air conditioning control panel of FIG.

FIG. 4 is a characteristic diagram of a cooling region of the temperature control lever.

FIG. 5 is a characteristic diagram of a dehumidification region of the temperature control lever.

FIG. 6 is a characteristic diagram of a heating region of the temperature control lever.

FIG. 7 is a sectional view showing a specific example of a refrigerant-refrigerant heat exchanger used in the first embodiment.

FIG. 8 is a sectional view showing a specific example of a first decompression device according to the first embodiment.

FIG. 9 is a chart for explaining the operation of the valve / door used in the first embodiment.

FIG. 10 is a Mollier chart showing the operation of the refrigeration cycle in the heating mode in the first embodiment in comparison with the conventional art.

11 is a Mollier diagram showing the operation of the first dehumidification mode D ₁ of the refrigeration cycle in the first embodiment.

12 is a Mollier diagram showing the operation of the second dehumidification mode D ₂ of the refrigeration cycle in the first embodiment.

FIG. 13 is a control flowchart of a second pressure reducing device (electric expansion valve) according to the first embodiment.

FIG. 14 is an explanatory diagram of an effect of the first embodiment.

FIG. 15 is a control flowchart of a second pressure reducing device (electric expansion valve) according to the second embodiment.

FIG. 16 is a control flowchart of a second pressure reducing device (electric expansion valve) according to the third embodiment.

FIG. 17 is a sectional view showing a specific example of a first decompression device according to a fourth embodiment.

FIG. 18 is an explanatory diagram of an effect of the fourth embodiment.

FIG. 19 is a Mollier diagram for describing effects of the fourth embodiment.

[Explanation of symbols]

11 evaporator, 12 condenser, 16, 17 passage switching door, 22 compressor, 22c gas injection port, 23 refrigerant-refrigerant heat exchanger, 24 outdoor heat exchanger
25: accumulator, 26: first decompression device, 27: second decompression device, 29: third decompression device.

Claims (9)

[Claims] 1. An air conditioning passage (2) through which air flows into a room, a discharge port (22a) for discharging compressed refrigerant, a suction port (22b) for sucking refrigerant on a low pressure side of the refrigeration cycle,
And a compressor (2) having a gas injection port (22c) for introducing a gas refrigerant on the refrigeration cycle intermediate pressure side.
(2) an indoor heat exchanger (1) installed in the air-conditioning passage (2) for heating air with a high-pressure gas refrigerant discharged from a discharge port (22a) of the compressor (22) in a heating mode.
2) a first pressure reducing device (26) that bypasses a part of the high-pressure refrigerant after passing through the indoor heat exchanger (12) in the heating mode, and depressurizes a part of the high-pressure refrigerant to an intermediate pressure; A refrigerant-refrigerant heat exchanger (23) that exchanges heat between the high-pressure refrigerant after passing through the indoor heat exchanger (12) and the intermediate-pressure refrigerant after passing through the first pressure reducing device (26) in the heating mode; A second pressure reducing device (27) for reducing the pressure of the high-pressure refrigerant cooled by the refrigerant-refrigerant heat exchanger (23) to a low pressure, and a low-pressure refrigerant having passed through the second pressure reducing device (27) in the heating mode. Outdoor heat exchanger that exchanges heat with the outside air (24)
And separating the gas-liquid of the low-pressure refrigerant that has passed through the outdoor heat exchanger (24) in the heating mode, and causing the liquid refrigerant in which the oil is dissolved and the low-pressure gas refrigerant to flow toward the suction port (22b). An accumulator (25), wherein the intermediate-pressure gas refrigerant gasified by heat exchange in the refrigerant-refrigerant heat exchanger (23) is introduced into the gas injection port (22c) in the heating mode. Further, the first pressure reducing device (26) is connected to an on-off valve (26a).
And a fixed throttle (26b) formed in a refrigerant passage in the on-off valve (26a). 2. The on-off valve is a solenoid valve (26a),
The refrigeration cycle apparatus according to claim 1, wherein the fixed throttle (26b) is formed by a refrigerant passage immediately after a valve port (265) of the solenoid valve (26a). 3. An air-conditioning passage (2) through which air flows toward a room, a discharge port (22a) for discharging compressed refrigerant, a suction port (22b) for sucking refrigerant on a low-pressure side of the refrigeration cycle,
And a compressor (2) having a gas injection port (22c) for introducing a gas refrigerant on the refrigeration cycle intermediate pressure side.
(2) an indoor heat exchanger (1) installed in the air-conditioning passage (2) for heating air with a high-pressure gas refrigerant discharged from a discharge port (22a) of the compressor (22) in a heating mode.
2) a first pressure reducing device (26) that bypasses a part of the high-pressure refrigerant after passing through the indoor heat exchanger (12) in the heating mode, and depressurizes a part of the high-pressure refrigerant to an intermediate pressure; A refrigerant-refrigerant heat exchanger (23) that exchanges heat between the high-pressure refrigerant after passing through the indoor heat exchanger (12) and the intermediate-pressure refrigerant after passing through the first pressure reducing device (26) in the heating mode; A second pressure reducing device (27) for reducing the pressure of the high-pressure refrigerant cooled by the refrigerant-refrigerant heat exchanger (23) to a low pressure, and a low-pressure refrigerant having passed through the second pressure reducing device (27) in the heating mode. Outdoor heat exchanger that exchanges heat with the outside air (24)
And separating the gas-liquid of the low-pressure refrigerant that has passed through the outdoor heat exchanger (24) in the heating mode, and causing the liquid refrigerant in which the oil is dissolved and the low-pressure gas refrigerant to flow toward the suction port (22b). An accumulator (25), wherein the intermediate-pressure gas refrigerant gasified by heat exchange in the refrigerant-refrigerant heat exchanger (23) is introduced into the gas injection port (22c) in the heating mode. The refrigeration cycle apparatus further comprises a plurality of switchable fixed throttles (26c, 26d) for the first pressure reducing device (26). 4. The fixed throttle (26c, 26d) of the plurality of stages is switched according to a state of at least one of refrigerants before and after the first pressure reducing device (26). A refrigeration cycle apparatus according to item 1. 5. A pressure response mechanism (268, 275) which responds to a pressure difference between before and after the first pressure reducing device (26), and when the pressure difference rises to a predetermined value or more, the plurality of fixed throttles ( The refrigeration cycle apparatus according to claim 4, wherein the opening degree of each of (26c, 26d) is increased. 6. The second decompression device (27) is an electric expansion valve, and according to the degree of supercooling of the high-pressure refrigerant cooled by the refrigerant-refrigerant heat exchanger (23). 2
The refrigeration cycle apparatus according to any one of claims 1 to 5, wherein the opening degree of (7) is controlled. 7. The second depressurizing device (27) is an electric expansion valve, and the second depressurizing device (27) is operated in accordance with the degree of superheat of the intermediate-pressure gas refrigerant introduced into the gas injection port (22c). The refrigeration cycle apparatus according to any one of claims 1 to 5, wherein the opening degree is controlled. 8. The method according to claim 1, wherein the second pressure reducing device is an electric expansion valve, and controls an opening degree of the second pressure reducing device in accordance with a heating load. The refrigeration cycle apparatus according to any one of the above. 9. An opening degree of the second pressure reducing device (27) at the time of starting heating is increased from a normal time.
9. The refrigeration cycle apparatus according to any one of items 8 to 8.