JP3467989B2 - Vapor compression refrigeration cycle - Google Patents

Description

Detailed Description of the Invention

[0001]

TECHNICAL FIELD The present invention relates to a vapor compression refrigeration system.
It is related to the control of the icle, ₂)etc
Vapor compression refrigeration cycle using refrigerant in the supercritical region of
It is suitable to use.

[0002]

2. Description of the Related Art In recent years, for example, Japanese Patent Publication No.
As described in JP-A-18602, carbon dioxide (C
Vapor compression refrigeration cycle using O ₂ (hereinafter CO
Abbreviated as ₂ cycles. ) Is proposed.

The operation of this CO ₂ cycle is, in principle, the same as the operation of a conventional vapor compression refrigeration cycle using freon. That is, as shown by A-B-C-D- A in FIG. 1 (CO ₂ Mollier chart), compressing the CO ₂ in the vapor phase by a compressor (A-B), the high-temperature high-pressure super CO ₂ in a critical state is cooled by a radiator (gas cooler) (BC). Then, the pressure is reduced by a pressure reducer (C-
D), by evaporating CO ₂ in the gas-liquid two-phase state (D-
A), the latent heat of vaporization is removed from the external fluid such as air to cool the external fluid. Since CO ₂ undergoes a phase change to a gas-liquid two-phase state from when the pressure falls below the saturated liquid pressure (the pressure at the intersection of the line segment CD and the saturated liquid line SL), the state of C changes from the state of C to the state of D. When it slowly changes to, CO ₂ changes from a supercritical state to a liquid-phase state and then to a gas-liquid two-phase state.

Incidentally, the supercritical state means a state in which CO ₂ molecules move like a gas phase state while the density is substantially equal to the liquid density. However, the critical temperature of CO ₂ is about 31
C and the conventional critical temperature of CFCs (for example, R12 is 11
2 ℃), the CO on the radiator side is low in summer, etc.
_{2 The} temperature becomes higher than the critical temperature of CO ₂ . That is, CO ₂ does not condense even on the radiator outlet side (the line segment BC does not intersect the saturated liquid line).

[0005] The state of the radiator outlet side (C point) is determined by the discharge pressure of the compressor and the CO ₂ temperature at the radiator outlet side, CO ₂ temperature at the radiator outlet side, the radiator It is determined by the heat radiation capacity of the and the outside air temperature. Since the outside air temperature cannot be controlled, the CO ₂ temperature at the radiator outlet side cannot be substantially controlled. Therefore, the state on the radiator outlet side (point C) can be controlled by controlling the discharge pressure of the compressor (radiator outlet side pressure). In other words, when the outside air temperature is high, such as in summer,
To secure a sufficient refrigerating capacity (enthalpy difference),
As shown by E-F-G-H-E in FIG. 1, it is necessary to increase the radiator outlet side pressure.

However, in order to increase the pressure on the radiator outlet side, the discharge pressure of the compressor must be increased as described above, so the compression work of the compressor (the enthalpy change ΔL in the compression process) increases. . Therefore, the evaporation process (D-
From the increase of the enthalpy change Δi of A), the compression process (A
If the amount of change in enthalpy change (refrigeration capacity) ΔL in −B) is large, the coefficient of performance of the CO ₂ cycle (COP =
Δi / ΔL) deteriorates.

Therefore, for example, if the CO ₂ temperature at the radiator outlet side is 40 ° C. and the relationship between the CO ₂ pressure at the radiator outlet side and the coefficient of performance (COP) is calculated using FIG.
As shown by the solid line in FIG. 5, the coefficient of performance becomes maximum at the pressure P ₁ (about 10 MPa). Similarly, C on the radiator outlet side
When the O ₂ temperature is 30 ° C., the coefficient of performance becomes maximum at the pressure P ₂ (about 8.0 MPa) as shown by the broken line in FIG.

As described above, CO _{2 at the} radiator outlet side
The temperature and the pressure at which the coefficient of performance is maximized are calculated, and the result is plotted on FIG. 1 as shown by the thick solid line η _max (hereinafter referred to as the optimum control line). Therefore, in order to operate the CO ₂ cycle efficiently, the pressure on the radiator outlet side and the CO ₂ temperature on the radiator outlet side are set to the optimum control line η.
_A decompression device for the refrigeration cycle that is controlled as shown by _max is required.

[0009]

By the way, since the coefficient of performance is the ratio of the enthalpy change amount (refrigerating capacity) in the evaporator to the compression work of the compressor as described above, the coefficient of performance is the maximum. However, there are cases where the refrigerating capacity sufficient to cool the room is not obtained. Therefore, when rapid cooling is performed as in so-called cool-down, there is a problem that the cooling capacity is insufficient and rapid cooling cannot be sufficiently performed.

In view of the above points, the present invention efficiently operates a vapor compression refrigeration cycle that operates in the supercritical region, and exhibits sufficient refrigeration capacity even when the heat load increases, such as during rapid cooling. An object of the present invention is to provide a decompression device for a refrigeration cycle that can be operated.

[0011]

The present invention uses the following technical means in order to achieve the above object. Claim 1
In the invention described in 4, first, when the heat load of the evaporator (4) is less than or equal to a predetermined value, the ratio of the refrigerating capacity of the evaporator (4) to the compression work of the compressor (1) becomes large. Thus, the opening degree of the decompression device (3) is adjusted. Secondly, when the heat load of the evaporator (4) exceeds a predetermined value, the opening degree of the decompression device (3) is decreased according to the increase of the heat load.

According to the first feature, the vapor compressor refrigeration cycle can be operated while maintaining a high coefficient of performance when the heat load on the evaporator (4) is below a predetermined value. Further, according to the second feature, when the heat load of the evaporator (4) exceeds a predetermined value, it is considered that rapid cooling is being performed and the heat load of the evaporator (4) is increased as described later. Since the opening degree of the decompression device (3) is reduced, the pressure of the refrigerant at the inlet side of the decompression device (3) (outlet side of the radiator 2) rises and the heat load becomes large, such as during rapid cooling. It is possible to exert sufficient freezing capacity.

Therefore, while operating the vapor compressor type refrigeration cycle efficiently, it is possible to exhibit a sufficient refrigerating capacity even when the heat load becomes large, such as during rapid cooling.
By the way, a vapor compression refrigeration cycle (hereinafter simply referred to as a refrigeration cycle) in which a refrigerant is operated at a critical pressure or lower, such as CFC, is a compressor compared to a vapor compression refrigeration cycle in which a refrigerant is operated at a critical pressure or higher. Since the fluctuation of the discharge pressure is small and the refrigerating capacity of the refrigerating cycle is increased / decreased mainly by increasing / decreasing the mass flow rate of the refrigerant circulating in the refrigerating cycle, the capacity of the compressor is determined by the maximum mass flow rate. Therefore, if the capacity of the compressor is determined by the maximum mass flow rate, the size of the compressor will be increased.

On the other hand, in the vapor compression refrigeration cycle according to the present invention, the cycle refrigeration is mainly performed by increasing or decreasing the outlet side pressure of the radiator (2) (inlet side pressure of the pressure reducing device (3)). Since the capacity is increased or decreased, the capacity of the compressor can be determined based on the refrigerant mass flow rate when the heat load of the evaporator (4) is stable. Therefore, upsizing of the compressor (1) can be suppressed.

In the invention described in claim 2, the evaporator (4)
When the heat load on the radiator (2) is less than or equal to a predetermined value, the pressure reducing device (so that the refrigerant pressure on the outlet side of the radiator (2) becomes a first target pressure selected based on the refrigerant temperature on the outlet side of the radiator (2) ( 3) is adjusted, and when the heat load on the evaporator (4) exceeds a predetermined value, the pressure is reduced to a second target pressure selected based on the heat load and the refrigerant temperature. It is characterized in that the opening degree of the device (3) is adjusted.

In the invention described in claim 3, the radiator (2)
Pressure difference (ΔP) between the outlet side and the evaporator (4) inlet side
However, the opening of the pressure reducing device (3) is controlled so that the target pressure difference is selected based on the refrigerant temperature on the outlet side of the radiator (2). This allows the radiator (2)
As will be described later, the pressure of the refrigerant on the outlet side increases / decreases in response to the increase / decrease in the heat load of the evaporator (4). Therefore, when the heat load of the evaporator (4) increases, heat dissipation is linked with this Bowl (2)
The refrigerant pressure on the outlet side rises, and the refrigeration capacity increases. on the other hand,
When the heat load on the evaporator (4) decreases, the radiator (2)
Since the refrigerant pressure on the outlet side drops, the refrigerating capacity drops.

The opening of the pressure reducing device (3) is controlled so that the pressure difference (ΔP) at this time becomes a target pressure difference selected based on the refrigerant temperature at the outlet side of the radiator (2). Therefore, the vapor compressor refrigeration cycle can be efficiently operated, as in the case of the first aspect of the invention. Therefore, the same effect as that of the invention according to claim 1 can be obtained without requiring any special means for detecting the heat load of the evaporator (4), so that the structure of the vapor compression refrigeration cycle is simple. As a result, the manufacturing cost of the vapor compression refrigeration cycle can be reduced.

The invention according to claim 4 is characterized in that carbon dioxide is used as the refrigerant. The reference numerals in parentheses of the above-mentioned means indicate the correspondence with the specific means described in the embodiments to be described later.

[0019]

DETAILED DESCRIPTION OF THE INVENTION

Hereinafter, the present invention will be described with reference to the embodiments shown in the drawings.
(First Embodiment) FIG. 3 shows a vapor compression refrigeration cycle (hereinafter abbreviated as CO ₂ cycle) using carbon dioxide (CO ₂ ) as a refrigerant, which is applied to a vehicle air conditioner.
Receives a driving force from a vehicle drive source such as an engine to generate gas-phase CO ₂
Is a compressor that compresses. 2 is C compressed by the compressor 1
A radiator (gas cooler) for exchanging heat between O ₂ and the outside air to cool it, and a decompression device ₃ for decompressing CO ₂ flowing out from the radiator 2.

The pressure reducing device 3 is provided on the outlet side of the radiator 2.
CO₂Pressure that controls the radiator 2 outlet side pressure according to the temperature
Also serves as force control means, and CO on the radiator 2 side₂Pressure is
The temperature is raised to the prescribed pressure described below and then reduced to low temperature.
Low-pressure gas-liquid two-phase CO₂Becomes 4 is the sky inside the vehicle
An evaporator (heat absorber) that serves as a gas cooling means,
CO ₂Is empty in the vehicle compartment when vaporizing (evaporating) in the evaporator 4.
The latent heat of vaporization is removed from the air to cool the air inside the vehicle. 5 is
CO in vapor phase₂And liquid phase CO₂And separate and surplus
CO₂Is stored temporarily and liquid phase CO₂Is pressure
With an accumulator that prevents it from being sucked into the compressor 1.
is there.

The compressor 1, the radiator 2, the decompression device 3, the evaporator 4 and the accumulator 5 are connected by pipes to form a closed circuit. By the way,
The radiator 2 is arranged in front of the vehicle in order to maximize the temperature difference between CO 2 inside the radiator ₂ and the outside air,
Reference numerals 6 and 7 denote fans that promote heat exchange between the radiator 2 and the evaporator 4.

In this embodiment, the refrigerating capacity of the CO ₂ cycle is controlled by adjusting the opening degree of the pressure reducing device 3, which will be described in detail later. FIG. 4 shows the structure of the decompression device 3 according to this embodiment, where 31 is an inflow port 32 communicating with the radiator 2 and an outflow port 33 communicating with the evaporator 4.
And a valve opening 34 for communicating the space 32a on the inflow port 32 side with the space 33a on the outflow port 33 side is formed in the housing 31. The needle-shaped valve element 35 for adjusting

Further, numeral 36 indicates that the valve body 35 is moved to move the valve mouth 3
4 is a step motor 36 for adjusting the opening degree,
The female rotor 3 is attached to the magnet rotor 36a of the drive motor 36.
6b is formed, and in the valve body 35, the female screw portion 36b is formed.
A male screw portion 35a that is screwed is formed. Also,
The step motor 36 is controlled by the controller 8.
As shown in FIG. 3, the control device 8 includes a radiator 2
On the outlet side (the inlet side of the pressure reducing device 3) of CO ₂Detect temperature
Temperature sensor 9 and the radiator 2 outlet side (the pressure reducing device 3
Mouth side) CO₂Pressure sensor 10 for detecting pressure and evaporation
The temperature sensor 11 disposed on the air downstream side of the container 4 and the indoor
Indoor temperature sensor 12 that detects the temperature of air, and outdoor air
Outdoor temperature sensor 13 to detect the temperature of the
The temperature setting means 14 for setting and inputting the room temperature
No. is input and the CO shown in FIG.₂Temperature and
Correspondence to the target pressure corresponding to this (hereinafter, temperature-pressure
Call it up. ) Is stored.

Then, the control device 8 calculates the target pressure corresponding to the CO ₂ temperature detected by the temperature sensor 9 from the temperature-pressure map (FIG. 5) and rotates the step motor 36 to open the valve opening 34. And the temperature setting means 1
Based on the set temperature set in step 4, the room temperature detected by the room temperature sensor 12, and the like, a well-known air mix door and a well-known air conditioning control means such as a fan 7 for controlling the temperature of the air blown into the vehicle interior are controlled. ing.

By the way, the pressure on the outlet side of the radiator 2 indicated by the optimum control line η _max described in the section of "Prior Art" is uniquely determined from the CO ₂ temperature on the outlet side of the radiator 2. Rather, strictly speaking, it is necessary to determine in consideration of the pressure fluctuation on the evaporator 4 side, that is, the fluctuation of the amount of heat given to the evaporator 4 (heat load of the evaporator 4). However, if the vapor compression refrigeration cycle is not limited to the CO ₂ cycle and is continuously operated,
The indoor temperature gradually stabilizes at a temperature at which the evaporation temperature of the refrigerant, the heat exchange capacity of the evaporator 4 and the amount of heat given to the room from the outside are balanced. Therefore, the pressure fluctuation on the evaporator 4 side also gradually stabilizes.

Therefore, when the refrigeration cycle is continuously operating, the pressure on the evaporator 4 side fluctuates abruptly unless the outside air temperature changes abruptly, except during rapid cooling such as during cool down. There is no such thing. Therefore, the optimum control line η _max may be determined based on the pressure on the evaporator 4 side corresponding to the CO ₂ evaporation temperature in the evaporator 4. The optimum control line η _max in FIG. 5 is the CO ₂ evaporation temperature (0
Based on the pressure (3.5 MPa) on the evaporator 4 side corresponding to (° C.), the target pressure (hereinafter referred to as the first target pressure) according to the CO ₂ temperature on the outlet side of the radiator 2 is shown. . In addition,
According to various tests conducted by the inventors, when the pressure on the radiator outlet side is equal to or lower than the critical pressure, a high coefficient of performance is maintained and CO ₂
In order to operate the cycle satisfactorily, it has been obtained that the degree of subcooling (subcool) on the inlet side of the decompression device 3 is preferably about 1 ° C to 10 ° C, and the optimum control line η in FIG.
_In consideration of this point, _max is calculated so that the degree of supercooling is about 3 ° C.

By the way, during rapid cooling, the heat load on the evaporator 4 increases, so that the pressure on the evaporator 4 side increases as the CO ₂ temperature in the evaporator 4 increases. Therefore, with the first target pressure calculated so that the coefficient of performance of the CO ₂ cycle is maximized when the pressure on the evaporator 4 side is 3.5 MPa,
As mentioned in the section “Problems to be solved by the invention”,
There are times when the refrigeration capacity sufficient to cool the room is not obtained.

Therefore, in the present embodiment, when the heat load of the evaporator 4 becomes larger than that when the heat load of the evaporator 4 is stable as in the case of rapid cooling, the heat load of the evaporator 4 becomes larger. The refrigerating capacity is enhanced by increasing the target pressure above the first target pressure (decreasing the opening degree of the pressure reducing device 3) in response to the increase. Next, the operation of the decompression device 3 will be described based on the flowchart shown in FIG.

The start switch (not shown) causes CO
_{When two} cycles are started, the temperature detected by the temperature sensor 9 (CO ₂ temperature at the inlet side of the decompression device 3) Tv, the room temperature (set temperature) desired by the personnel set in the temperature setting means 14
The detected temperature (indoor temperature) Tr is fetched from Tset and the indoor temperature sensor 12 (steps 100 to 12).
0).

Then, the room temperature Tr and the set temperature Tset
And the temperature difference δT is calculated (step 130), and the temperature difference δ
When T is equal to or lower than the predetermined value To, it is considered that the heat load of the evaporator 4 is equal to or lower than the predetermined value and is stable, and the temperature-
From the pressure map (FIG. 5), the target pressure (first target pressure) corresponding to the CO ₂ temperature Tv on the inlet side of the pressure reducing device 3 is set (steps 140 and 150).

On the other hand, when the temperature difference δT exceeds the predetermined value To, it is considered that the heat load of the evaporator 4 exceeds the predetermined value and rapid cooling is being performed, and the pressure-temperature difference δT map (FIG. 7) is used. Select the pressure rise δP corresponding to the temperature difference δT,
The target pressure (hereinafter referred to as the second target pressure) is obtained by the sum of the selected pressure increase δP and the target pressure (first target pressure) corresponding to the CO ₂ temperature Tv selected from the temperature-pressure map (FIG. 5). .) Is set (steps 140 and 16).
0).

Next, the detected pressure from the pressure sensor 10 (CO ₂ pressure on the inlet side of the pressure reducing device 3) is taken in (step 170), and the target pressure and the pressure taken in at step 170 (hereinafter referred to as pressure reducing device inlet pressure). .) Is compared (step 180). Then, when the target pressure exceeds the pressure reducing device inlet pressure, the opening degree of the pressure reducing device 3 is reduced (step 190), and when the target pressure is equal to or lower than the pressure reducing device inlet pressure, the opening degree of the pressure reducing device 3 is decreased. Increase (Step 2
00), the opening of the decompression device 3 is adjusted so that the target pressure and the decompression device inlet pressure are substantially equal.

Then, the process returns to step 100, and the subsequent steps
Repeat from 100 to 200. In addition, δP is uncontrolled
If you set it to the limit, the coefficient of performance will drop extremely.
In this embodiment, 2.7 MPa is set as the upper limit of δP.
It was Next, the features of this embodiment will be described. I mentioned above
As described above, according to this embodiment, the heat load of the evaporator 4 is a predetermined value.
It is considered that the heat load of the evaporator 4 is stable when:
Then, CO on the inlet side of the decompression device 3 ₂The pressure is the first target pressure
Since the opening of the decompression device 3 is adjusted so that
CO while maintaining high₂Can drive the cycle
Wear.

Further, when the heat load of the evaporator 4 exceeds a predetermined value, it is considered that it is during rapid cooling, and the opening degree of the decompression device 3 is reduced in accordance with the increase of the heat load of the evaporator 4. ,
The CO ₂ pressure on the inlet side of the decompression device 3 rises. Therefore, a sufficient refrigerating capacity can be exhibited even when the heat load becomes large, such as during rapid cooling. That is, according to the present embodiment, it is possible to efficiently operate the CO ₂ cycle and to exert a sufficient refrigerating capacity even when the heat load becomes large such as during rapid cooling.

By the way, in a vapor compression refrigeration cycle (hereinafter simply referred to as a refrigeration cycle) in which a refrigerant is operated at a pressure below the critical pressure such as CFC, the discharge pressure fluctuation of the compressor is smaller than that in the CO ₂ cycle. Since the refrigerating capacity of the refrigerating cycle is increased or decreased mainly by increasing or decreasing the mass flow rate of the refrigerant circulating in the refrigerating cycle, the capacity of the compressor is determined by the maximum mass flow rate. Therefore, if the capacity of the compressor is determined by the maximum mass flow rate, the size of the compressor will be increased.

On the other hand, in the present embodiment, in the CO ₂ cycle according to the present embodiment, by mainly increasing / decreasing the outlet side pressure of the radiator 2 (inlet side pressure of the pressure reducing device 3), the C 2
Since the refrigerating capacity of the O ₂ cycle is increased or decreased, the capacity of the compressor can be determined based on the CO ₂ mass flow rate when the heat load of the evaporator 4 is stable. Therefore, the compressor 1
Can be prevented from increasing in size.

(Second Embodiment) In the above-described embodiment, the temperature difference δT between the room temperature Tr and the set temperature Tset is used as means for detecting the heat load of the evaporator 4, but in the present embodiment, FIG. As shown in (1), the room temperature sensor 12 for detecting the room temperature Tr and the temperature setting means 14 for setting the set temperature Tset can be eliminated to reduce the manufacturing cost of the entire CO ₂ cycle.

For this reason, in this embodiment, instead of the pressure reducing device 3 shown in the first embodiment, the radiator 2 outlet side (the inlet side of the pressure reducing device 30) and the evaporator 4 inlet side (the pressure reducing device 30).
The opening of the pressure reducing device is adjusted so that the pressure difference ΔP from the outlet side of the pressure reducing device becomes a target pressure difference selected based on the CO ₂ temperature at the radiator 2 outlet side (the inlet side of the pressure reducing device 30). The decompression device 30 is used. The decompression device 30 according to this embodiment will be described below with reference to FIG. 9.

Reference numeral 36 is a coil spring for generating an elastic force that pushes the valve element 35 toward the valve opening 34, and one end of this coil spring 36 is pressed by the elastic force of the coil spring 36 from the outlet 33 side. First to act on body 35
A pressing plate 40 is provided, and the coil spring 36 is provided on the other end side.
The second pressing plate 41 for adjusting the elastic force of the is provided.

The second pressing plate 41 has a female screw portion 41.
a is formed, and a male screw portion 42a formed on an adjusting shaft 42 for moving the second pressing plate 41 in the axial direction of the coil spring 36 is screwed to the female screw portion 41a. Therefore, since the second pressing plate 41 moves in the axial direction of the coil spring 36 in association with the rotation of the adjustment shaft 42, the elastic force acting on the valve body 35 can be adjusted.

Reference numeral 43 is a key for preventing the second pressing plate 41 from rotating together with the adjusting shaft 42.
Is an O-ring made of nitrile rubber that forms a seal member that seals the gap between the adjustment shaft 42 and the housing 31.
Incidentally, the operating pressure of the CO ₂ cycle is higher than that of a normal vapor compression refrigeration cycle using freon (about 8
Therefore, when a seal member such as an O-ring is arranged in the movable part, it is desirable to dispose the seal member on the side of the outflow port 33 which has been decompressed as described above.

Reference numeral 45 is a step motor which serves as an actuator for rotating the adjusting shaft 42. The step motor 45 has a target pressure difference selected based on the CO ₂ temperature Tv on the inlet side of the pressure reducing device 30. It is controlled by the controller 8. Next, the operation of the decompression device 30 according to this embodiment will be described.

As is apparent from FIG. 9, the acting force F ₁ due to the outlet side pressure of the radiator 2 acts on the inflow port 32 side of the valve body 35, so that the valve body 35 is pressed toward the outflow port 33 side. To be done. On the other hand, the inlet side pressure of the evaporator 2 and the acting force F ₂ due to the elastic force of the coil spring 36 act on the outflow port 33 side, so that the valve element 35 is pressed toward the inflow port 32 side. That is, when the acting force F ₂ is larger than the acting force F ₁ , the valve body 35 moves so that the opening degree of the valve opening 34 becomes smaller, and when the acting force F ₁ is larger than the acting force F _2. , Valve body 35
Moves so that the opening degree of the valve opening 34 increases.

Therefore, the valve body 35 stops at a position where the acting force F ₁ and the acting force F ₂ are in balance (or a position where the acting force F ₂ contacts the valve opening 34), so that the opening degree of the valve opening 34 depends on the coil spring 36. Is determined by the elastic force exerted on the valve element 35. That is, the pressure difference ΔP between the spaces 32 a and 33 a corresponds to the elastic force exerted by the coil spring 36 on the valve body 35. Since the movement amount (lift amount) of the valve body 35 is small, the change in the elastic force exerted by the coil spring 36 on the valve body 35 can be almost ignored, so that the pressure difference ΔP between the spaces 32a and 33a is It becomes almost constant.

Next, the pressure difference ΔP will be described. When the temperature inside the evaporator 4 becomes below freezing (0 ° C. or lower), frost is generated on the evaporator 4 and the refrigerating capacity of the evaporator 4 is deteriorated. Therefore, the temperature inside the evaporator 4 should be higher than the freezing point. Is desirable.
However, if the temperature inside the evaporator 4 is unnecessarily raised,
There is a problem that the air blown into the vehicle compartment cannot be cooled sufficiently.

Therefore, in this embodiment, as in the first embodiment, the pressure on the inlet side of the evaporator 4 is set to the evaporation pressure (3.5 MPa) corresponding to the evaporation temperature of CO ₂ , and the pressure on the outlet side of the radiator 2 is set. _Is a pressure (first target pressure) indicated by the optimum control line η _max , and a pressure difference ΔP between the spaces 32a and 33a is defined as. Therefore, the control device 8 detects C detected by the temperature sensor 9 from the temperature-pressure map (FIG. 5).
The first target pressure corresponding to the O ₂ temperature is calculated, and steps are performed so that the difference between the first target pressure and the evaporation pressure (3.5 MPa at 0 ° C.) corresponding to the evaporation temperature of CO ₂ is the pressure difference ΔP. The motor 45 is driven to adjust the elastic force of the coil spring 36.

Next, the operation and characteristics of the CO ₂ cycle according to this embodiment will be described. FIG. 10 is a Mollier diagram showing the operation of the CO ₂ cycle according to the present embodiment. In FIG. 10, the dashed-dotted line diagram A shows the case where the refrigerating capacity (heat load) is small, and the solid line diagram B. Indicates the case where the heat load is large. When the heat load on the evaporator 4 is large, the temperature of the air cooled by the evaporator 4 also rises, so the temperature inside the evaporator 4 rises as the heat load on the evaporator 4 rises, and the temperature inside the evaporator 4 rises. The pressure and the vaporization pressure of CO ₂ rise (see FIG. 11).
reference).

Therefore, the acting force F ₂ becomes large and the opening of the valve port 34 becomes small as described above.
The outlet side pressure rises (state of diagram B). Therefore, the difference in specific enthalpy between the outlet and the inlet of the evaporator 4 becomes large (h ₁ > h ₂ ), so that the refrigerating capacity is increased. Then, the heat load of the evaporator 4, as described above, gradually CO ₂
Since it stabilizes toward the evaporation pressure (3.5 MPa) corresponding to the evaporation temperature of, the acting force F ₂ becomes small and the opening of the valve port 34 becomes large as described above. Therefore, as the outlet side pressure of the radiator 2 decreases (state of the line A), the outlet side pressure of the radiator 2 approaches the pressure (first target pressure) indicated by the optimum control line η _max. Go.

Therefore, according to this embodiment, CO ₂
While operating the cycle efficiently, it is possible to exert sufficient refrigerating capacity even when the heat load increases, such as during rapid cooling. Further, the room temperature sensor 12 and the temperature setting means 14 are not required as means for detecting the heat load of the evaporator 4, and the pressure difference ΔP between the radiator 2 outlet side and the evaporator 4 inlet side.
Since the above effect can be obtained by controlling the CO ₂ cycle by a simple means such as adjusting the opening degree of the valve opening 34 so that is a predetermined value, the manufacturing cost of the CO ₂ cycle can be reduced. be able to. (Third Embodiment) In the above-described embodiment, the temperature sensor 9 detects the CO ₂ temperature on the outlet side of the radiator 2, but in the present embodiment, as shown in FIG. It is abolished and the CO ₂ cycle according to the second embodiment is controlled by estimating the CO ₂ temperature at the outlet side of the radiator 2 from the temperature detected by the outdoor temperature sensor 13.

That is, since the CO ₂ temperature on the outlet side of the radiator 2 is substantially determined by the outside air temperature and the capacity of the radiator 2, the outside air temperature and the CO ₂ on the outlet side of the radiator 2 are determined.
_If the relationship with the _two temperatures is obtained in advance by a test or the like, the CO ₂ temperature on the outlet side of the radiator 2 can be calculated from the outside air temperature. By the way, when the CO ₂ temperature on the outlet side of the radiator 2 is detected by the temperature sensor 9, the temperature sensor 9 and the outside air are thermally coupled with the temperature sensor 9 assembled near the outlet of the radiator 2. Insulation that shuts off is needed. For this reason,
The material cost of the heat insulating material, the number of steps for assembling the temperature sensor 9 and the heat insulating material are required, which causes an increase in the manufacturing cost of the CO ₂ cycle.

On the other hand, according to the present embodiment, since it is not necessary to provide the temperature sensor 9, it is possible to reduce the material cost of the heat insulating material and the number of steps for assembling the temperature sensor 9 and the heat insulating material. Therefore, it is possible to suppress an increase in manufacturing cost in the CO ₂ cycle. Further, since the outdoor temperature sensor 13 provided for controlling the automatic air conditioner (auto air conditioner) can be used, it is not necessary to newly provide the outdoor temperature sensor 13. Therefore, it is possible to further suppress the manufacturing cost increase in the CO ₂ cycle.

By the way, in the second and third embodiments, the elastic force generated by the coil spring 36 is controlled by the step motor 45. However, like the so-called thermal expansion valve, the pressure on the evaporator 4 side is controlled by a thin tube or the like. The elastic force of the coil spring 36 may be controlled by guiding the pressure reducing device 3 and actuating a pressure responsive member such as a diaphragm by the pressure thus guided.

Further, in the second and third embodiments, the pressure difference ΔP is mechanically controlled by using the coil spring 36, but the pressure inside the evaporator 4 is detected by the pressure sensor and based on this detected value. The present invention can be implemented by using an electric actuator such as a proportional control solenoid valve that can adjust the valve opening. Further, the pressure sensor 10 may detect the pressure on the inlet side of the radiator 2. However, when the pressure loss in the radiator 2 is large, it is necessary to compensate for the pressure loss.

In addition, in the second embodiment, CO
_Although the pressure of CO ₂ circulating in the _two cycles was directly applied, the pressure inside the evaporator 4 or at the outlet side or the inlet side of the evaporator 4 is taken out by a thin tube or the like, and the valve element 35 is operated via a diaphragm or the like. You may let me. Further, the decompression device 3 according to the present invention is not limited to use in a vapor compression refrigeration cycle using CO _2, and uses, for example, a refrigerant used in a supercritical region such as ethylene, ethane, and nitric oxide. It can also be applied to a conventional vapor compression refrigeration cycle.

Further, even if the accumulator 5 is eliminated, the vapor compression refrigeration cycle can be carried out.
In this case, the refrigerant remaining in the evaporator 4 is sucked, and the same operation as in the CO ₂ cycle having the accumulator 5 can be obtained.

[Brief description of drawings]

FIG. 1 is a Mollier diagram of CO ₂ .

FIG. 2 is a graph showing a relationship between a coefficient of performance (COP) and a radiator outlet side pressure.

FIG. 3 is a schematic diagram showing a vapor compression refrigeration cycle according to the first embodiment.

FIG. 4 is a cross-sectional view of the pressure reducing device according to the first embodiment.

FIG. 5 is a map showing the relationship between the target pressure and the CO ₂ temperature on the inlet side of the pressure reducing device.

FIG. 6 is a flowchart showing an operation of the pressure reducing device according to the first embodiment.

FIG. 7 is a map showing a relationship between a temperature difference δT and a pressure increase δP.

FIG. 8 is a schematic diagram showing a vapor compression refrigeration cycle according to a second embodiment.

FIG. 9 is a sectional view of a decompression device according to a second embodiment.

FIG. 10 is a Mollier diagram showing the operation of the vapor compression refrigeration cycle according to the second embodiment.

FIG. 11 is a map showing the relationship between evaporation temperature and pressure change.

FIG. 12 is a schematic diagram showing a vapor compression refrigeration cycle according to a third embodiment.

[Explanation of symbols]

1 ... Compressor, 2 ... Radiator, 3 ... Pressure reducing device, 4 ... Evaporator,
5 ... Accumulator (tank means), 6, 7 ... Fan, 8 ... Control device, 9 ... Temperature sensor, 10 ... Pressure sensor.

─────────────────────────────────────────────────── --Continued from the front page (56) References JP-A-7-294033 (JP, A) JP-A-6-185814 (JP, A) JP-A-5-133618 (JP, A) (58) Field (Int.Cl. ⁷ , DB name) F25B 1/00 395 F25B 1/00 304

Claims (4)

(57) [Claims] 1. A compressor (1) for compressing a refrigerant, and a radiator (2) for cooling the refrigerant compressed by the compressor (1), the internal pressure of which exceeds the critical pressure of the refrigerant. A pressure reducing device (3) for reducing the pressure of the refrigerant flowing out from the radiator (2); and an evaporator (4) for evaporating the refrigerant reduced in pressure by the pressure reducing device (3). When the heat load of 4) is less than a predetermined value, the opening degree of the decompression device (3) is adjusted so that the ratio of the refrigerating capacity of the evaporator (4) to the compression work of the compressor (1) becomes large. Further, when the heat load of the evaporator (4) exceeds a predetermined value, the opening degree of the decompression device (3) is reduced according to the increase of the heat load. Refrigeration cycle. 2. When the heat load of the evaporator (4) is below a predetermined value, the refrigerant pressure on the outlet side of the radiator (2) is selected based on the refrigerant temperature on the outlet side of the radiator (2). The opening degree of the decompression device (3) is adjusted so as to reach the first target pressure, and when the heat load of the evaporator (4) exceeds a predetermined value, the heat load and the refrigerant temperature are The vapor compression refrigeration cycle according to claim 1, wherein the opening degree of the pressure reducing device (3) is adjusted so as to reach the second target pressure selected based on the above. 3. The pressure difference (ΔP) between the radiator (2) outlet side and the evaporator (4) inlet side is the radiator (2).
The vapor compression refrigeration cycle according to claim 1, wherein the opening degree of the pressure reducing device (3) is controlled so that a target pressure difference selected based on a refrigerant temperature on the outlet side is obtained. 4. The vapor compression refrigeration cycle according to any one of claims 1 to 3, wherein carbon dioxide is used as the refrigerant.