JP2010043861A - Refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a refrigerating cycle capable of consistently and uniformly demonstrating the excellent cooling performance of a plurality of evaporators. <P>SOLUTION: There are provided a plurality of evaporators 5, 11 for evaporating the low-pressure refrigerant after passing pressure-reducing means 4, 11, an accumulator 9 is provided on an outlet side of one evaporator 5, and an internal heat exchanger 3 is provided on an outlet side of the accumulator. An outlet side of the other evaporator 11 is merged with a part between the outlet side of the accumulator 9 and an inlet side of the internal heat exchanger 3. A pressure reducing means 10 on the other evaporator side is constituted of the fixed restriction. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to a refrigeration cycle including a plurality of evaporators, and is used for a supercritical refrigeration cycle using a refrigerant whose high pressure is equal to or higher than a critical pressure (supercritical state) such as a CO2 (carbon dioxide) refrigerant. Is preferred.

  Conventionally, a supercritical refrigeration cycle of this type has a configuration in which a plurality of evaporators 5 and 11 are connected in parallel as shown in FIG. 22 (see FIG. 1 of Patent Document 1). In this prior art, the first decompressor 4 that decompresses the refrigerant flowing into the first evaporator 5 among the plurality of evaporators 5 and 11 is configured by an electric valve mechanism, and the outlet refrigerant of the radiator (outdoor unit) 2. The temperature is detected by the temperature sensor 21, and the radiator outlet refrigerant pressure is detected by the pressure sensor 22.

  Then, by controlling the opening of the first pressure reducer 4 with the control output of the control device 20, the outlet refrigerant pressure of the radiator 2 is controlled to become a target pressure determined by the radiator outlet refrigerant temperature, and the cycle This improves the operating efficiency (COP). An accumulator 9 is disposed on the outlet side of the first evaporator 5 to prevent liquid refrigerant from being sucked into the compressor 1 from the flow path on the first evaporator 5 side.

  On the other hand, the second pressure reducer provided in parallel with the first pressure reducer 4 is constituted by a temperature type expansion valve 100, and the temperature type expansion valve 100 reduces the pressure of the refrigerant flowing into the second evaporator 11. It has become. The temperature type expansion valve 100 has a temperature sensing part 100 a whose internal pressure changes according to the outlet refrigerant temperature of the second evaporator 11, and controls the degree of superheat of the outlet refrigerant of the second evaporator 11. Thereby, the liquid refrigerant is prevented from being sucked into the compressor 1 from the flow path on the second evaporator 11 side.

  In Patent Document 1, as another example, as shown in FIG. 23, the second decompressor is configured by a fixed throttle 10, and the outlet side of the second evaporator 11 is the outlet side of the accumulator 9 (compressor suction side). (See FIG. 14 of Patent Document 1).

JP 2000-35250 A

  By the way, in the prior art shown in FIG. 22, the temperature type expansion valve 100 which performs the superheat degree control of the exit refrigerant | coolant of the 2nd evaporator 11 independently according to the cooling load fluctuation | variation of the 2nd evaporator 11 is used as a 2nd pressure reduction device. Therefore, when the cooling load of the second evaporator 11 increases, the cycle behavior shown in FIG. 24 occurs, the high pressure control by the first pressure reducer 4 becomes unstable, and hunting occurs. As a result, the cycle operating efficiency (COP) decreases.

  In the case of the prior art shown in FIG. 22, when the air flow rate of the second evaporator 11 is suddenly reduced and the cooling load is suddenly reduced, or when the rotational speed of the compressor 1 is suddenly increased and the low pressure is lowered, A fatal problem like this also occurs. That is, when the low pressure is lowered, in the temperature type expansion valve 100, since the responsiveness of the temperature sensing unit 100a is extremely lower than the responsiveness of the pressure sensing unit, the superheat degree is excessive.

  As a result, in the temperature type expansion valve 100, the valve opening increases to near full open, and as a result, most of the circulating refrigerant in the cycle flows to the second evaporator 11 side, and the liquid returns at the outlet of the second evaporator 11. It becomes the state of. On the other hand, an insufficient refrigerant flow rate occurs on the first evaporator 5 side, and the outlet refrigerant of the first evaporator 5 has a large degree of superheat.

  In addition, at this time, since the temperature type expansion valve 100 is in a state of being fully opened and the high pressure is lowered, the shortage of the refrigerant flow on the first evaporator 5 side is further promoted, and the cooling performance of the first evaporator 5 is improved. The phenomenon of a significant drop occurs.

  In addition, when the refrigeration cycle is started from a stopped state, when the low pressure decreases as the compressor 1 starts, the valve opening of the temperature expansion valve 100 increases to near full open as described above, and the first evaporation is performed. Since the refrigerant flow rate shortage on the side of the evaporator 5 occurs, the cooling performance of the first evaporator 5 cannot be exhibited well. And since a high pressure does not rise quickly, the start-up as the whole refrigerating cycle gets worse.

  On the other hand, when the cooling load increases rapidly or the compressor rotation speed decreases rapidly and the low pressure increases, the temperature expansion valve 100 generates a state of excessive superheat and the valve opening decreases rapidly. The cooling performance of the second evaporator 11 is insufficient.

  On the other hand, in the prior art shown in FIG. 23, since the second pressure reducer is the fixed throttle 10, the above-mentioned problems associated with the use of the temperature type expansion valve can be solved, but on the other hand, the following problems occur.

  That is, the hole diameter of the fixed throttle 10 constituting the second pressure reducer needs to be a hole diameter that allows the flow rate of the refrigerant to match the required capacity of the second evaporator 11 at the maximum cooling load. However, when the diameter of the fixed throttle hole is set to a required size at the maximum cooling load, the second evaporation is performed when the cooling load of the second evaporator 11 is small (when the air volume is small or the intake air temperature is low). The refrigerant flow rate on the side of the evaporator 11 becomes excessive, and the outlet refrigerant of the second evaporator 11 contains a large amount of liquid refrigerant.

  As a result, the liquid refrigerant returns to the compressor 1 and excessive stress due to liquid compression occurs, which adversely affects the life of the compressor. Further, since the liquid refrigerant cannot be effectively used for exhibiting the evaporator cooling performance, the cycle operation efficiency is deteriorated.

  On the contrary, if the fixed throttle hole diameter is set smaller than the required size at the maximum cooling load, the refrigerant flow rate of the second evaporator 11 becomes insufficient when the second evaporator 11 is heavily loaded, and the outlet refrigerant of the second evaporator 11 Has an excessive degree of superheat, resulting in insufficient cooling performance. Further, when this excessive degree of superheat occurs, a large temperature distribution is generated in the blown air of the second evaporator 11, white mist is generated in the blown air, or the indoor temperature distribution is increased to deteriorate the air conditioning feeling. Let

  The present invention has been devised in view of the above points. In a refrigeration cycle that includes a plurality of evaporators and connects an accumulator to the outlet side of at least one evaporator, the cooling performance of the plurality of evaporators is improved. The purpose is to enable both to be demonstrated stably and satisfactorily.

  Another object of the present invention is to stabilize high-pressure control in a refrigeration cycle that includes a plurality of evaporators and that includes at least decompression means that controls high-pressure.

  Another object of the present invention is to protect the compressor by reliably preventing the liquid refrigerant from returning to the compressor.

The present invention has been devised to achieve the above object, and in the invention according to claim 1, a compressor (1) for sucking and compressing refrigerant,
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
Among the plurality of evaporators (5, 11), an accumulator (9) provided on the outlet side of one evaporator (5),
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Of the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Further, the one or more decompression means is characterized in that the refrigerant supplied to at least the other evaporator (11) is decompressed by a fixed throttle (4, 10).

  According to this, the refrigerant supplied to the other evaporator (11) that joins the outlet side of the accumulator (9) among the plurality of evaporators (5, 11) is decompressed by the fixed throttle (4, 10). Thus, unlike the prior art of FIG. 22, there is no problem such as high pressure control hunting associated with superheat degree control of the refrigerant at the outlet of the evaporator, and cycle operating efficiency (COP) can be improved.

  Further, in the case of a temperature type expansion valve, when the low-pressure pressure is suddenly reduced, the valve opening becomes excessively large, resulting in insufficient refrigerant flow on one evaporator (5) side, and thus insufficient cooling performance. Then, since the refrigerant on the other evaporator (11) side is depressurized by the fixed throttle (4, 10), there is no shortage of refrigerant flow on the one evaporator (5) side when the low pressure is suddenly reduced. The cooling performance of the evaporator (5) can be ensured.

  By the way, in the present invention, the refrigerant on the other evaporator (11) side is depressurized by the fixed throttle (4, 10) that does not control the superheat degree, so when the cooling load of the other evaporator (11) is small, the other The liquid refrigerant is included in the outlet refrigerant of the evaporator (11), and the liquid refrigerant on the other evaporator (11) side merges with the outlet refrigerant on the one evaporator (5) side, Since it flows into the heat exchanger (3), the liquid refrigerant can be evaporated by heat exchange with the high-pressure side refrigerant in the internal heat exchanger (3).

  Therefore, even if the fixed throttles (4, 10) are used, the return of the liquid refrigerant to the compressor (1) can be reliably prevented, the durability of the compressor (1) can be improved, and the liquid refrigerant can be used as an internal heat exchanger. It can be effectively used to increase the amount of heat exchange in (3), and the cycle operating efficiency (COP) can be improved.

  According to a second aspect of the present invention, in the refrigeration cycle according to the first aspect, only one fixed throttle is provided and the refrigerant supplied to the plurality of evaporators (5, 11) by the one fixed throttle. The pressure is reduced.

  According to this, since the refrigerant to the plurality of evaporators (5, 11) can be depressurized by using only one fixed throttle having a simple structure, the manufacturing cost of the depressurizing means can be effectively reduced.

  As in the third aspect of the invention, in the refrigeration cycle of the first aspect, the decompression means (4) for decompressing the refrigerant supplied to the one evaporator (5), and the other evaporator ( The fixed throttle (10) for depressurizing the refrigerant supplied to 11) may be provided independently.

  According to this, the flow rate of the refrigerant flowing into each evaporator (5, 11) can be controlled by the decompression means (4, 10) provided corresponding to each evaporator (5, 11).

In the invention according to claim 4, a compressor (1) for sucking and compressing refrigerant,
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
Among the plurality of evaporators (5, 11), an accumulator (9) provided on the outlet side of one evaporator (5),
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Of the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Further, the one or more decompression means is configured to decompress the refrigerant supplied to at least the other evaporator (11) with a variable throttle (10),
The variable throttle (10) is characterized in that when the high pressure is lowered, the opening degree is changed in a decreasing direction.

  The invention of claim 4 is different from claim 1 in that the refrigerant to the other evaporator (11) is depressurized by the variable throttle (10).

  The variable throttle (10) in claim 4 changes in a direction in which the opening degree decreases when the high pressure decreases, and does not control the degree of superheat of the outlet refrigerant of the other evaporator (11). Moreover, since the variable throttle (10) acts to suppress a decrease in the high pressure due to a decrease in the opening thereof, similarly to claim 1, there is no problem such as hunting of the high pressure control associated with the superheat control. The operating efficiency (COP) of the cycle can be improved.

  Further, since the variable throttle (10) of claim 4 does not directly respond only to the change of the low pressure, the phenomenon that the opening degree becomes excessive, such as that of the temperature type expansion valve, does not occur when the low pressure is rapidly reduced. For this reason, when the low-pressure pressure is suddenly reduced, there is no shortage of refrigerant flow on the one evaporator (5) side, and the cooling performance of one evaporator (5) can be ensured.

  And even if the liquid refrigerant is contained in the outlet refrigerant of the other evaporator (11) when the other evaporator (11) is in a cooling low load, the liquid refrigerant on the other evaporator (11) side is contained inside. Since it flows into the heat exchanger (3) and evaporates, the liquid refrigerant can be reliably prevented from returning to the compressor (1), and the liquid refrigerant can be effectively used to increase the amount of heat exchange in the internal heat exchanger (3). The cycle operating efficiency (COP) can be improved.

  Further, since the variable throttle (10) of claim 4 changes in a direction in which the opening decreases when the high pressure decreases, the opening of the variable throttle (10) increases under the condition that the high pressure increases. Thus, the refrigerant flow rate to the other evaporator (11) can be increased. Since the condition for increasing the high-pressure is when the cooling load of the entire cycle increases, the refrigerant flow rate to the other evaporator (11) can be adjusted according to the load fluctuation as compared with the fixed throttle. is there.

  As in the invention of claim 5, in the refrigeration cycle of claim 4, specifically, the variable throttle (10) changes its opening according to the differential pressure between the high pressure and the low pressure. Any one of a differential pressure valve, a high-pressure responsive valve that changes the valve opening according to the high pressure, and a high-pressure refrigerant temperature responsive valve that changes the valve opening according to the high-pressure refrigerant temperature may be used.

  As in the invention described in claim 6, in the refrigeration cycle according to claim 4 or 5, the decompression means (4) for decompressing the refrigerant supplied to the one evaporator (5), and the other evaporation The variable throttle (10) for depressurizing the refrigerant supplied to the vessel (11) may be provided independently.

  According to a seventh aspect of the present invention, in the refrigeration cycle according to the third or sixth aspect, the decompression means (4) on the one evaporator (5) side is a high-pressure refrigerant temperature on the outlet side of the radiator (2). The high pressure is controlled according to the above.

  According to this, cycle operating efficiency (COP) can be improved by high-pressure control by the decompression means (4). Moreover, even in the cycle in which the high pressure is positively controlled by the pressure reducing means (4), the pressure reducing means on the other evaporator (11) side is a fixed throttle or variable throttle that does not perform superheat control. The high pressure hunting phenomenon can be avoided reliably.

  In addition, in claim 7, the expression “high-pressure refrigerant temperature on the outlet side of the radiator (2)” represents not only the temperature of the high-pressure refrigerant immediately after the radiator (2) but also the reduced pressure on the outlet side of the internal heat exchanger (3). The high-pressure refrigerant temperature between the inlet side of the means (4) is included (see FIG. 16 and the like described later).

  The interpretation of such terms is the same in other claims adopting the expression “high-pressure refrigerant temperature on the outlet side of the radiator (2)”.

In the invention described in claim 8, a compressor (1) for sucking and compressing refrigerant,
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
One decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
Among the plurality of evaporators (5, 11), an accumulator (9) provided on the outlet side of one evaporator (5),
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Of the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Further, the one pressure reducing means controls the high pressure according to the high pressure refrigerant temperature on the outlet side of the radiator (2).

  The invention according to claim 8 corresponds to that in which the pressure reducing means is constituted by one pressure reducing means of the high pressure control type according to claim 7 with respect to claim 1 or claim 4. Therefore, according to the eighth aspect, similarly to the first and fourth aspects, since the superheat degree control by the temperature type expansion valve is not performed, the unstable cycle behavior by the superheat degree control does not occur.

  Moreover, since the return of the liquid refrigerant to the compressor (1) can be reliably prevented, the durable life of the compressor (1) can be improved. In addition, since the liquid refrigerant on the outlet side of the other evaporator (11) can be effectively used to increase the amount of heat exchange in the internal heat exchanger (3), combined with the use of the decompression means of the high pressure control type, The operating efficiency (COP) of the cycle can be effectively improved. Moreover, the structure of a refrigerating cycle can be simplified by using only one decompression means.

In the invention described in claim 9, a compressor (1) for sucking and compressing refrigerant,
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
An accumulator (9) provided between the outlet side of the plurality of evaporators (5, 11) and the suction side of the compressor (1),
Among the plurality of evaporators (5, 11), a refrigerant flow path is provided in which the refrigerant exiting at least one of the evaporators is introduced to the inlet side of the other evaporator.

  According to this, the refrigerant that has exited one evaporator can flow into the other evaporator, and the refrigerant that has passed through the other evaporator can flow into the accumulator (9). For this reason, when the cooling load of one evaporator is small, even if liquid refrigerant is contained in the outlet refrigerant of one evaporator, this liquid refrigerant flows into the other evaporator and the cooling performance of the other evaporator It can be used effectively for demonstration.

  Since the saturated gas after passing through the accumulator (9) can be sucked into the compressor (1), the liquid refrigerant can be reliably prevented from returning to the compressor (1) without having an internal heat exchanger. The durable life of (1) can be improved.

  As in the invention described in claim 10, in the refrigeration cycle according to claim 9, the refrigerant flowing into the plurality of evaporators (5, 11) is depressurized by one decompression means (4). Thus, the manufacturing cost of the decompression means can be reduced.

  As in the invention described in claim 11, in the refrigeration cycle according to claim 9, a plurality of decompression means (4, 10) corresponding to the plurality of evaporators (5, 11) as the decompression means, respectively. May be provided.

  According to a twelfth aspect of the present invention, in the refrigeration cycle according to the eleventh aspect, one of the plurality of pressure reducing means has a pressure reducing means (4) according to a high-pressure refrigerant temperature on the outlet side of the radiator (2). It is characterized by controlling high pressure.

  According to this, cycle operating efficiency (COP) can be improved by high-pressure control by the decompression means (4).

  As in the invention described in claim 13, in the refrigeration cycle according to claim 12, if the other decompression means (10) of the plurality of decompression means is a fixed throttle, the manufacturing cost of the decompression means can be reduced. . Further, since the fixed throttle does not control the degree of superheat of the evaporator outlet refrigerant, it can contribute to stabilization of high-pressure control.

In invention of Claim 14, in the refrigerating cycle of Claim 12, the other decompression means of the plurality of decompression means is a variable throttle (10),
The variable throttle (10) is characterized in that when the high pressure is lowered, the opening degree is changed in a decreasing direction.

  Thereby, the same effect as the variable aperture (10) in Claim 4 can be exhibited.

  As in the invention described in claim 15, in the refrigeration cycle according to claim 11, at least one of the plurality of decompression means (4, 10) may be constituted by a fixed throttle (10). .

As in the invention according to claim 16, in the refrigeration cycle according to claim 11, at least one of the plurality of decompression means (4, 10) is a variable throttle (10),
The variable throttle (10) may be configured to change in a direction in which the opening degree decreases when the high pressure decreases.

  As in the invention described in claim 17, in the refrigeration cycle according to claim 14 or 16, specifically, the variable throttle (10) has an opening according to a differential pressure between the high pressure and the low pressure. Any one of a differential pressure valve that changes the pressure, a high-pressure responsive valve that changes the valve opening according to the high-pressure, and a high-pressure refrigerant temperature responsive valve that changes the valve opening according to the high-pressure refrigerant temperature may be used. .

  As in the invention described in claim 18, in the refrigeration cycle according to any one of claims 9 to 17, between the outlet side of the accumulator (9) and the suction side of the compressor (1). An internal heat exchanger (3) that performs heat exchange between the low-pressure refrigerant and the high-pressure refrigerant on the outlet side of the radiator (2) may be provided.

  Thereby, the operating efficiency (COP) improvement effect of the cycle by the internal heat exchanger (3) can be exhibited.

  As in the invention described in claim 19, in the refrigeration cycle according to any one of claims 1 to 18, the refrigerant flow to at least one of the plurality of evaporators (5, 11). If the on-off valve (16) for stopping the cooling is provided, the refrigerant flow to one evaporator can be interrupted depending on whether or not the cooling operation is necessary.

  As in the invention described in claim 20, in the refrigeration cycle according to any one of claims 1 to 18, a switching valve (17) for switching the refrigerant flow to the plurality of evaporators (5, 11) is provided. If provided, the refrigerant flow to the plurality of evaporators can be interrupted depending on whether or not the cooling action is necessary.

  The invention according to claim 21 is the refrigeration cycle according to any one of claims 1 to 20, wherein the refrigeration cycle is a supercritical refrigeration cycle using a refrigerant whose high pressure is equal to or higher than the critical pressure.

  By the way, since the high-pressure refrigerant does not condense in the supercritical refrigeration cycle, the fluctuation of the high-pressure pressure is larger than that of a subcritical refrigeration cycle using a normal chlorofluorocarbon refrigerant. In a supercritical refrigeration cycle having such characteristics, the present invention can contribute to the stabilization of high pressure by adopting a pressure reducing means such as a temperature type expansion valve that does not have a superheat degree control function.

According to a twenty-second aspect of the present invention, there is provided a vehicle air conditioner including the refrigeration cycle according to any one of the first to twenty-first aspects,
Among the plurality of evaporators (5, 11), one of the evaporators (5) is disposed in the air conditioning unit (6) on the front seat side of the vehicle interior,
Of the plurality of evaporators (5, 11), the other evaporator (11) is disposed in the air conditioning unit (12) on the rear seat side of the vehicle interior.

  Thus, the present invention can be suitably implemented as a dual air conditioner type vehicle air conditioner including the air conditioning unit (6) on the front seat side of the vehicle interior and the air conditioning unit (12) on the rear seat side of the vehicle interior.

  In addition, the code | symbol in the bracket | parenthesis of each said means shows the correspondence with the specific means as described in embodiment mentioned later.

It is a block diagram of the refrigerating cycle which shows 1st Embodiment of this invention. It is a Mollier diagram explaining the effect of the internal heat exchanger of FIG. It is a block diagram of the refrigerating cycle which shows 2nd Embodiment. It is a block diagram of the refrigerating cycle which shows 3rd Embodiment. It is a block diagram of the refrigerating cycle which shows 4th Embodiment. It is a block diagram of the refrigerating cycle which shows 5th Embodiment. It is a block diagram of the refrigerating cycle which shows 6th Embodiment. It is a block diagram of the refrigerating cycle which shows 7th Embodiment. It is a block diagram of the refrigerating cycle which shows 8th Embodiment. It is a block diagram of the refrigerating cycle which shows 9th embodiment. It is a block diagram of the refrigerating cycle which shows 10th Embodiment. It is a block diagram of the refrigerating cycle which shows 11th Embodiment. It is a block diagram of the refrigerating cycle which shows 12th Embodiment. It is a block diagram of the refrigerating cycle which shows 13th Embodiment. It is a block diagram of the refrigerating cycle which shows 14th Embodiment. It is a block diagram of the refrigerating cycle which shows 15th Embodiment. It is a block diagram of the refrigerating cycle which shows 16th Embodiment. (A) is a whole sectional view of a differential pressure valve which constitutes the 2nd decompressor in a 17th embodiment, (b) is a sectional view of a valve element, and (c) is a front view of a valve element. It is a refrigerant | coolant flow rate adjustment characteristic figure of the differential pressure | voltage valve which comprises the 2nd pressure reduction device in 17th Embodiment. It is a block diagram of the refrigerating cycle which shows 18th Embodiment. It is a block diagram of the refrigerating cycle which shows 19th Embodiment. It is a block diagram of the refrigerating cycle which shows a prior art. It is a block diagram of the refrigerating cycle which shows another example of a prior art. It is explanatory drawing of the cycle behavior (problem) in the prior art of FIG.

(First embodiment)
FIG. 1 is a configuration diagram of a refrigeration cycle for vehicle air conditioning showing a first embodiment of the present invention, and this refrigeration cycle uses CO2 whose high pressure is not less than a critical pressure (supercritical state) as a refrigerant. Therefore, this refrigeration cycle constitutes a supercritical refrigeration cycle.

  The compressor 1 obtains driving force from a vehicle travel engine (not shown) via an electromagnetic clutch (not shown) and sucks and compresses the refrigerant. The compressor 1 may be either a fixed displacement compressor whose operation is interrupted by the engagement of an electromagnetic clutch or a variable displacement compressor whose discharge capacity can be changed.

  A radiator 2 serving as an outdoor unit is provided on the discharge side of the compressor 1. The radiator 2 cools the refrigerant by exchanging heat between the refrigerant discharged from the compressor 1 in a high-temperature and high-pressure supercritical state and the outside air (outdoor air). Outside air is blown to the radiator 2 by an electric cooling fan (not shown).

  A high-pressure side refrigerant flow path 3 a of the internal heat exchanger 3 is provided on the outlet side of the radiator 2. Here, the internal heat exchanger 3 performs heat exchange between the high-temperature high-pressure refrigerant in the high-pressure side refrigerant flow path 3a and the low-temperature low-pressure refrigerant in the low-pressure side refrigerant flow path 3b.

  By this heat exchange, the enthalpy of the refrigerant flowing into the evaporators 5 and 11, which will be described later, is reduced, and the enthalpy difference (refrigeration capacity) of the refrigerant between the refrigerant inlet and outlet of the evaporators 5 and 11 is increased. When the internal heat exchanger 3 is installed in this way, the enthalpy difference (refrigeration capacity) of the refrigerant between the refrigerant inlets and outlets of the evaporators 5 and 11 can be increased, and the cycle operation efficiency (COP) can be improved.

  A first refrigerant channel A and a second refrigerant channel B are provided in parallel on the outlet side of the high-pressure side refrigerant channel 3 a of the internal heat exchanger 3. C is a branch point of both the flow paths A and B, and D is a junction of both the flow paths A and B. In the present embodiment, the first refrigerant passage A constitutes a refrigerant passage for front seat air conditioning, and the second refrigerant passage B constitutes a refrigerant passage for rear seat air conditioning.

  A first decompressor 4 that serves as a pressure control valve for controlling the high pressure is provided in the first refrigerant flow path A. The opening of the first pressure reducer 4 is adjusted by a mechanical mechanism so that the high pressure of the cycle becomes the target high pressure.

  The decompressor 4 has a temperature sensing part 4a provided between the outlet side of the radiator 2 and the inlet side of the high-pressure side refrigerant flow path 3a of the internal heat exchanger 3, and the inside of the temperature sensing part 4a. In addition, a pressure corresponding to the temperature of the high-pressure refrigerant on the outlet side of the radiator 2 is generated.

  The throttle opening (valve) of the first pressure reducer 4 is balanced by the balance between the internal pressure of the temperature sensing portion 4a and the high pressure (specifically, the outlet side refrigerant pressure of the high pressure side refrigerant flow path 3a of the internal heat exchanger 3). By adjusting the body opening degree, the high pressure is adjusted to a target high pressure determined by the high-pressure refrigerant temperature on the outlet side of the radiator 2. A decompressor 4 having such a high-pressure control function is known in Japanese Patent Application Laid-Open No. 2000-81157.

  A first evaporator 5 is provided on the outlet side of the first decompressor 4 in the first refrigerant flow path A. The first evaporator 5 is disposed in a case 7 that forms an air passage of the front seat side air conditioning unit 6 of the vehicle air conditioner, and constitutes a cooling means for cooling the air in the case 7. The front seat side indoor air conditioning unit 6 is normally disposed in the inner space of the instrument panel (instrument panel) in the front part of the vehicle interior. An electric blower 8 is arranged on the upstream side of the air flow of the first evaporator 5 so that the inside air or outside air introduced through an inside / outside air switching box (not shown) is blown into the case 7.

  In addition, a heater core (not shown) or the like that forms heating means for heating air is disposed in the case 7 on the downstream side of the air flow of the first evaporator 5, and the conditioned air whose temperature is adjusted according to the degree of heating of the heater core. It blows off from the blower outlet (not shown) of the air flow downstream edge part of case 7 to the front seat side area | region in a vehicle interior.

  An accumulator 9 is provided on the outlet side of the first evaporator 5. This accumulator 9 is a gas-liquid separation means that separates the liquid refrigerant and gas refrigerant of the outlet refrigerant of the first evaporator 5 and stores excess refrigerant in the cycle, and the outlet side of the accumulator 9 is the internal heat exchanger 3. Is connected to the inlet side of the low-pressure side refrigerant flow path 3b. The outlet side of the low-pressure side refrigerant flow path 3 b of the internal heat exchanger 3 is connected to the suction side of the compressor 1.

  On the other hand, the second decompressor 10 and the second evaporator 11 are also provided in series in the second refrigerant channel B. The second pressure reducer 10 is a fixed throttle with a constant throttle opening. This fixed throttle is specifically composed of an orifice, a capillary tube, and the like.

  The outlet side of the second evaporator 11 joins between the outlet side of the accumulator 9 and the inlet side of the low-pressure side refrigerant flow path 3b of the internal heat exchanger 3. Therefore, the refrigerant after the outlet refrigerant of the accumulator 9 and the outlet refrigerant of the second evaporator 11 join together flows into the low-pressure side refrigerant flow path 3 b of the internal heat exchanger 3.

  The 2nd evaporator 11 is arrange | positioned in the case 13 which forms the air path of the backseat side indoor air conditioning unit 12 of a vehicle air conditioner, and comprises the cooling means which cools the air in this case 13. FIG. The rear seat side indoor air conditioning unit 12 is arranged in a region on the rear seat side of the vehicle interior (for example, a vehicle body side plate portion on the rear seat side). An electric blower 14 is disposed on the upstream side of the air flow of the second evaporator 11 so that the inside air (vehicle compartment air) is blown into the case 13.

  The cold air cooled by the second evaporator 11 is blown out from a blowout port (not shown) at the downstream end of the air flow of the case 13 to the rear seat side region of the vehicle interior. Note that a heater core or the like may be installed in the rear seat side indoor air conditioning unit 12 as in the front seat side air conditioning unit 6.

  Next, the operation of this embodiment in the above configuration will be described. When the compressor 1 is rotationally driven by the driving force of the vehicle engine, the high-temperature and high-pressure refrigerant (CO2) compressed by the compressor 1 flows into the radiator 2 in a supercritical state where the pressure is higher than the critical pressure. To do. Here, the high-temperature and high-pressure supercritical refrigerant exchanges heat with the outside air to dissipate heat into the outside air, thereby reducing enthalpy.

  The high-pressure refrigerant at the outlet of the radiator 2 flows into the high-pressure side passage 3a of the internal heat exchanger 3, and exchanges heat with the low-temperature low-pressure refrigerant (outlet-side refrigerant of the accumulator 9) passing through the low-pressure side passage 3b. Since it is cooled, enthalpy is further reduced.

  The high-pressure refrigerant that has passed through the high-pressure side passage 3a of the internal heat exchanger 3 then flows in parallel through the first refrigerant passage A and the second refrigerant passage B that are connected in parallel. The pressure is reduced by the pressure reducer 10. In the first evaporator 5, the low-pressure refrigerant after passing through the first pressure reducer 4 absorbs heat from the blown air of the electric blower 8 and evaporates. As a result, the air blown from the electric blower 8 is cooled, and the cold air is blown out to the front seat side of the passenger compartment to cool the front seat side area of the passenger compartment.

  At the same time, in the second evaporator 11, the low-pressure refrigerant after passing through the second decompressor 10 absorbs heat from the blown air of the electric blower 14 and evaporates. As a result, the air blown from the electric blower 14 is cooled, and the cold air is blown out to the rear seat side of the vehicle interior to cool the rear seat side region of the vehicle interior.

  Here, the first pressure reducer 4 adjusts the throttle opening so that the actual high pressure becomes a target high pressure determined by the radiator outlet side refrigerant temperature sensed by the temperature sensing unit 4a. That is, when the actual high pressure falls below the target high pressure, the first pressure reducer 4 decreases its throttle opening, and conversely, when the actual high pressure rises above the target high pressure, the first pressure reducer 4 Increase the throttle opening. By adjusting the throttle opening of the first pressure reducer 4, the actual high pressure is maintained at the target high pressure, and the cycle COP is improved.

  In particular, in the present embodiment, since the second pressure reducer 10 is configured with a fixed throttle, the opening degree of the second pressure reducer 10 does not change even when the cycle operation condition fluctuates. Only by this, the high pressure can be stably controlled.

  Incidentally, since the second pressure reducer 10 is a fixed throttle, it does not have a function of controlling the phase state of the outlet refrigerant of the second evaporator 11. For this reason, the state of the refrigerant at the outlet of the second evaporator 11 has a degree of superheat or a wet steam with a predetermined dryness due to a change in the cooling load of the second evaporator 11.

  Therefore, when the outlet refrigerant of the second evaporator 11 is merged with the inlet side of the accumulator 9, the outlet of the first evaporator 5 is canceled so that the degree of superheat is canceled when the outlet refrigerant of the second evaporator 11 is overheated. Since the refrigerant balances in the state of wet steam having a predetermined dryness, the enthalpy difference between the inlet and the outlet of the first evaporator 5 is reduced, and the cooling ability of the first evaporator 5 is reduced.

  However, in this embodiment, since the outlet side of the second evaporator 11 is merged between the outlet side of the accumulator 9 and the inlet side of the low-pressure side refrigerant flow path 3b of the internal heat exchanger 3, the second evaporation is performed. Without being influenced by the phase state of the outlet refrigerant of the evaporator 11, the state of the outlet refrigerant of the first evaporator 5 is always maintained in the vicinity of a saturated gas with a dryness of 1 due to the presence of the gas-liquid interface in the accumulator 9. The Thereby, the cooling performance of the 1st evaporator 5 is securable stably.

  By the way, the first evaporator 5 is a main cooling means for cooling the front seat side area of the vehicle interior, whereas the second evaporator 11 is a sub cooling means for cooling the rear seat side area of the vehicle interior. The required cooling performance of the first evaporator 5 is larger than the required cooling performance of the second evaporator 11. For this reason, it is necessary to secure the cooling performance of the first evaporator 5 with priority over the second evaporator 11. From this point of view, the above effect that the cooling performance of the first evaporator 5 can be ensured is extremely useful in practice.

  Further, since the outlet refrigerant of the second evaporator 11 merges with the outlet refrigerant (saturated gas) of the accumulator 9, it flows into the low-pressure side refrigerant flow path 3b of the internal heat exchanger 3, so when the second evaporator 11 is under low load Even when the outlet refrigerant of the second evaporator 11 is in the state of wet steam having a predetermined dryness, the liquid phase refrigerant of the wet steam can be effectively used for cooling the high-pressure refrigerant in the internal heat exchanger 3.

  That is, the high-pressure refrigerant can be further cooled by evaporating the liquid-phase refrigerant contained in the outlet refrigerant of the second evaporator 11 in the internal heat exchanger 3, thereby further reducing the enthalpy of the high-pressure refrigerant, The cooling capacity of the vessels 5 and 11 can be improved. At the same time, the return of the liquid refrigerant to the compressor 1 can be prevented, so that the compressor 1 can be protected by avoiding the liquid compression of the compressor 1.

  FIG. 2 is a Mollier diagram showing the effect of this embodiment. A solid line a indicates the refrigerant state in the cycle with the internal heat exchanger according to the present embodiment, and a broken line b indicates the refrigerant state in the cycle without the internal heat exchanger. Show.

  In FIG. 2, L1 is the state of the inlet refrigerant of the low-pressure side refrigerant flow path 3b of the internal heat exchanger 3, that is, the outlet refrigerant (saturated gas) of the accumulator 9 and the outlet refrigerant of the second evaporator 11 (wetness of a predetermined dryness). L2 indicates the state of the refrigerant exiting from the low-pressure side refrigerant flow path 3b of the internal heat exchanger 3. H1 indicates the state of the inlet refrigerant (radiator 2 outlet refrigerant) of the high-pressure side refrigerant flow path 3a of the internal heat exchanger 3, and H2 indicates the state of the outlet refrigerant of the high-pressure side refrigerant flow path 3a of the internal heat exchanger 3. .

  As can be understood from FIG. 2, the enthalpy of the high-pressure refrigerant is reduced by Δi by heat exchange in the internal heat exchanger 3, and the cooling capacity corresponding to the decrease Δi can be improved, and the low pressure of the internal heat exchanger 3 can be improved. Even when the inlet refrigerant of the side refrigerant flow path 3b is in the state of wet steam containing liquid phase refrigerant, the outlet refrigerant is in an overheated state, so that return of the liquid refrigerant to the compressor 1 can be reliably prevented.

  In the present embodiment, the throttle opening (throttle diameter) of the fixed throttle constituting the second pressure reducer 10 can satisfy the required refrigerant flow rate at the maximum cooling load under the normal operation condition of the second evaporator 11. It is preferable to set to.

(Second Embodiment)
FIG. 3 shows a second embodiment, in which the first pressure reducer 4 of the first embodiment is arranged upstream of the branch point C of the first refrigerant flow path A and the second refrigerant flow path B.

  As a result, the first pressure reducer 4 functions to depressurize the refrigerant flowing into both the first refrigerant flow path A and the second refrigerant flow path B. As a result, the second decompressor 10 is disposed in the flow path of the low-pressure refrigerant after passing through the first decompressor 4, so that the throttle opening degree of the second decompressor 10 is the second decompressor of the first embodiment. The second decompressor 10 can be easily manufactured.

  The first decompressor 4 of the second embodiment is the same as the first embodiment in that it is a pressure control valve that controls the high pressure to a target high pressure determined by the radiator outlet refrigerant temperature.

(Third embodiment)
FIG. 4 shows a third embodiment, in which the second pressure reducer 10 in the second embodiment is abolished and only one first pressure reducer 4 is provided as a pressure reducing means. Thereby, the cost reduction effect can be further improved.

  According to the third embodiment, the flow resistance ratio between the first refrigerant flow path A and the second refrigerant flow path B, that is, the flow path A between the branch point C and the merge point D of both flow paths A and B. The ratio of the refrigerant flow rate in the first refrigerant flow path A and the refrigerant flow rate in the second refrigerant flow path B can be determined by the ratio of the resistance of the first flow path and the resistance of the flow path B.

(Fourth embodiment)
FIG. 5A shows a fourth embodiment. In the third embodiment, an auxiliary throttle mechanism 15 for adjusting the flow resistance is added to the inlet side of the first evaporator 5. The auxiliary throttle mechanism 15 can be composed of a fixed throttle such as an orifice or a capillary tube.

  By this auxiliary throttle mechanism 15, the distribution ratio between the refrigerant flow rate in the first refrigerant flow path A and the refrigerant flow rate in the second refrigerant flow path B is appropriately set in advance according to the cooling load ratio of the first evaporator 5 and the second evaporator 11. Can be set to

  Note that, in the fourth embodiment, the auxiliary throttle mechanism according to the change of the cooling load ratio of the first and second evaporators 5 and 11, the change of the flow resistance ratio of the first and second refrigerant flow paths A and B, and the like. 15 may be provided not in the first refrigerant channel A but in the second refrigerant channel B. Therefore, it can be said that the second decompressor 10 of the second embodiment (FIG. 3) plays the role of the auxiliary throttle mechanism 15 on the second refrigerant flow path B side.

  Further, in FIG. 5A, the first pressure reducer 4 is configured by a pressure control valve (variable throttle) that controls the high pressure based on the radiator outlet refrigerant temperature sensed by the temperature sensing unit 4a. You may comprise the 1st pressure reduction device 4 with the fixed aperture_diaphragm | restriction which does not have the temperature sensing part 4a like FIG.5 (b). In this way, the decompression means for the first and second evaporators 5 and 11 can be configured with only one fixed throttle, and the cost can be further reduced.

  Similarly, also in the above-described third embodiment (FIG. 4), the first pressure reducer 4 may be configured with a fixed throttle.

(Fifth embodiment)
FIG. 6 shows a fifth embodiment. In the first embodiment, the first pressure reducer 4 is also composed of a fixed throttle such as an orifice. Thereby, both the pressure reducers 4 and 10 on the front seat side and the rear seat side are constituted by fixed throttles, and the cost reduction effect can be further improved.

  However, according to the fifth embodiment, since the function of high pressure control is lost, the high pressure is controlled in accordance with fluctuations in cycle operation conditions, and efficient operation by optimal COP cannot be performed.

  However, also in the fifth embodiment, since the outlet refrigerant of the second evaporator 11 is merged with the refrigerant of the first refrigerant flow path A on the outlet side of the accumulator 9, the phase of the outlet refrigerant of the first evaporator 5 The state can always be stably controlled in the vicinity of the saturated gas without being influenced by the phase state of the outlet refrigerant of the second evaporator 11. Moreover, the point which can use effectively the liquid-phase refrigerant | coolant contained in the exit refrigerant | coolant of the 2nd evaporator 11 for the cooling of a high pressure refrigerant | coolant in the internal heat exchanger 3 is the same as 1st Embodiment etc.

(Sixth embodiment)
FIG. 7 shows a sixth embodiment. In the first embodiment, the second refrigerant flow path B is provided with an electromagnetic valve 16 serving as an on-off valve that interrupts the refrigerant flow in the second refrigerant flow path B.

  Specifically, when the solenoid valve 16 is provided on the inlet side of the second decompressor 10 in the second refrigerant flow path B and the cooling action of the second evaporator 11 is not necessary, that is, the rear seat side air conditioning. When the operation of the unit 12 is stopped, the electromagnetic valve 16 is closed to block the refrigerant flow in the second refrigerant flow path B. The electromagnetic valve 16 may be provided on the outlet side of the second pressure reducer 10.

  In FIG. 7, the second pressure reducer 10 and the electromagnetic valve 16 are illustrated as separate parts. However, since the second pressure reducer 10 is a fixed throttle, the housing (not shown) of the electromagnetic valve 16. An orifice that forms a throttle passage can be integrally formed with the refrigerant flow path. As a result, the second pressure reducer 10 can be integrally formed with the electromagnetic valve 16.

  Further, instead of the electromagnetic valve 16, an electrical three-way switching valve is provided at a branch point C of the first refrigerant flow path A and the second refrigerant flow path B, and the first switching is performed by switching the flow of the three-way switching valve. A state in which the refrigerant flows only in the refrigerant flow path A, a state in which the refrigerant flows only in the second refrigerant flow path B, and a state in which the refrigerant flows in both flow paths A and B may be selected. For the three-way switching valve, refer to the three-way switching valve 17 shown in FIGS.

(Seventh embodiment)
FIG. 8 shows a seventh embodiment. In the refrigeration cycle without the internal heat exchanger 3, the outlet side of the second evaporator 11 (the downstream end of the second refrigerant flow path B) is connected to the first decompressor 4. A connection is made between the outlet side and the inlet side of the first evaporator 5. That is, the series passage of the second decompressor 10 and the second evaporator 11 is connected in parallel to the first decompressor 4. Thereby, the refrigerant that has passed through the second evaporator 11 flows to the inlet side of the first evaporator 5.

  Also in the seventh embodiment, the second pressure reducer 10 is a fixed throttle and does not have a function of controlling the phase state of the outlet refrigerant of the second evaporator 11. Therefore, when the cooling load of the second evaporator 11 is reduced, the outlet refrigerant of the second evaporator 11 may be in the state of wet steam containing liquid phase refrigerant. Also in that case, since the outlet refrigerant of the second evaporator 11 flows into the inlet side of the first evaporator 5, this liquid-phase refrigerant can be effectively used for exhibiting the cooling performance of the first evaporator 5.

  Then, since the saturated gas after passing through the accumulator 9 is sucked into the compressor 1, the return of the liquid refrigerant to the compressor 1 can be reliably prevented.

  In addition, since the second pressure reducer 10 is configured with a fixed throttle, the opening degree of the second pressure reducer 10 does not change even when the cycle operation condition fluctuates. Can be controlled stably.

(Eighth embodiment)
FIG. 9 shows an eighth embodiment, in which the first pressure reducer 4 of the seventh embodiment is arranged upstream of the branch point C of the first refrigerant flow path A and the second refrigerant flow path B. The arrangement of the first pressure reducer 4 is the same as in the second embodiment (FIG. 3).

  However, since the outlet side of the second evaporator 11 is connected to the inlet side of the first evaporator 5 in the eighth embodiment, an auxiliary throttle is formed in the first refrigerant flow path A immediately after the branch point C. A mechanism 15 is provided so as to generate a predetermined pressure loss between both ends of the series passage of the second decompressor 10 and the second evaporator 11. As in the fourth embodiment (FIG. 5), the auxiliary throttle mechanism 15 can be constituted by a fixed throttle such as an orifice or a capillary tube.

  Also in the eighth embodiment, since the refrigerant that has passed through the second evaporator 11 flows to the inlet side of the first evaporator 5, the same operational effects as those of the seventh embodiment can be exhibited. In addition, since the second decompressor 10 is disposed in the flow path of the low-pressure refrigerant after passing through the first decompressor 4, the throttle opening degree of the second decompressor 10 is the same as that of the second decompressor 10 of the seventh embodiment. The second pressure reducer 10 can be easily manufactured as long as it is sufficiently large.

(Ninth embodiment)
FIG. 10 shows a ninth embodiment, in which the second pressure reducer 10 provided in the second refrigerant flow path B of the eighth embodiment is eliminated.

  Therefore, according to the ninth embodiment, since the second refrigerant flow path B having only the second evaporator 11 is connected in parallel to the auxiliary throttle mechanism 15, the flow resistance of the auxiliary throttle mechanism 15 and the second refrigerant are The refrigerant flow distribution to the first evaporator 5 and the second evaporator 11 is determined by the ratio of the flow path B to the flow path resistance. For this reason, the refrigerant flow distribution can be adjusted by selecting the flow path resistance (throttle opening) of the auxiliary throttle mechanism 15.

  In the present embodiment, the first pressure reducer 4 is configured by a pressure control valve (variable throttle) that controls high pressure, but the first pressure reducer 4 may be configured by a fixed throttle. In this way, the decompression means for the first and second evaporators 5 and 11 can be configured with only one fixed throttle, and the cost can be further reduced.

(10th Embodiment)
FIG. 11 shows a tenth embodiment. In the seventh embodiment (FIG. 8), an internal heat exchanger 3 is added to improve COP.

(Eleventh embodiment)
FIG. 12 shows an eleventh embodiment. In the tenth embodiment (FIG. 11), the electromagnetic valve 16 serving as an on-off valve that interrupts the refrigerant flow in the second refrigerant flow path B is replaced with the second refrigerant flow path B. Provided. This solenoid valve 16 plays the role which interrupts | blocks the refrigerant | coolant flow of the 2nd refrigerant | coolant flow path B at the time of operation stop of the backseat side air conditioning unit 12 similarly to the solenoid valve 16 of 6th Embodiment (FIG. 7).

  As described above in the sixth embodiment, the electromagnetic valve 16 is not limited to the inlet side of the second pressure reducer 10, and may be provided on the outlet side of the second pressure reducer 10, and the electromagnetic valve 16 includes a fixed throttle. The second decompressor 10 may be integrally formed.

(Twelfth embodiment)
FIG. 13 shows a twelfth embodiment. Instead of the electromagnetic valve 16 of the eleventh embodiment (FIG. 12), an electric three-way switching valve 17 is branched between the first refrigerant flow path A and the second refrigerant flow path B. It is provided at the point.

  By switching the flow path of the three-way switching valve 17, the refrigerant flows only in the first refrigerant flow path A, the refrigerant flows only in the second refrigerant flow path B, and the refrigerant flows in both flow paths A and B simultaneously. The state can be selected.

(13th Embodiment)
FIG. 14 shows a thirteenth embodiment. In the cycle configuration of the ninth embodiment (FIG. 10), an electric three-way switching valve 17 is provided at the branch point of the first and second refrigerant flow paths A and B. It corresponds to a thing. The flow path switching action of the three-way switching valve 17 is the same as in the twelfth embodiment.

  In the cycle configuration of the eighth embodiment (FIG. 9), an electric three-way switching valve 17 may be provided at the branch point of the first and second refrigerant flow paths A and B.

(14th Embodiment)
In the seventh embodiment (FIG. 8) to the thirteenth embodiment (FIG. 14), the outlet side of the second evaporator 11 is the outlet side of the first decompressor 4 and the inlet side of the first evaporator 5. In the fourteenth embodiment, the refrigerant that has passed through the second evaporator 11 flows to the inlet side of the first evaporator 5, but in the fourteenth embodiment, as shown in FIG. The second evaporator 11 is connected to the outlet side of the first evaporator 5 so that the passed refrigerant flows to the inlet side of the second evaporator 11.

  An electric three-way switching valve 17 is provided at a branch point between the outlet side of the first evaporator 5 and the inlet side of the second evaporator 11, and an auxiliary throttle mechanism 15 is provided on the outlet side of the three-way switching valve 17. Thus, a predetermined pressure loss is generated between both ends of the second refrigerant flow path B having only the second evaporator 11.

  By switching the flow path of the three-way switching valve 17, (1) a state in which refrigerant flows from the first pressure reducing device 4 to the first evaporator 5 → the three-way switching valve 17 → the auxiliary throttle mechanism 15 → the accumulator 9 (first evaporator 5 And (2) a state in which the refrigerant flows through the first decompressor 4 → the first evaporator 5 → the three-way switching valve 17 → the second evaporator 11 → the accumulator 9 (with the first evaporator 5). (A state in which the refrigerant flows in series through the second evaporator 11) and (3) a state in which the refrigerant after passing through the first evaporator 5 flows through the auxiliary throttle mechanism 15 and the second evaporator 11 in parallel. Yes.

  In the thirteenth and fourteenth embodiments, the three-way switching valve 17 and the auxiliary throttle mechanism 15 are illustrated separately. However, since the auxiliary throttle mechanism 15 can be configured by a fixed throttle, The auxiliary throttle mechanism 15 may be integrally formed in the refrigerant flow path.

(Fifteenth embodiment)
In the first embodiment and the like, the temperature sensing part 4a of the first pressure reducer 4 is arranged between the outlet side of the radiator 2 and the inlet side of the high-pressure side refrigerant flow path 3a of the internal heat exchanger 3, In the fifteenth embodiment, as shown in FIG. 16, the temperature sensing part 4 a of the first pressure reducer 4 is arranged on the outlet side of the high-pressure side refrigerant flow path 3 a of the internal heat exchanger 3. Even in this case, the same high pressure control function can be exhibited by the first pressure reducer 4.

(Sixteenth embodiment)
In the first embodiment and the like, the first pressure reducer 4 is constituted by a mechanical valve mechanism. However, in the sixteenth embodiment, the throttle opening degree of the first pressure reducer 4 is controlled by an electric mechanism as shown in FIG. It consists of an electric valve mechanism that adjusts.

  Specifically, the first pressure reducer 4 is provided with a valve body 4b for adjusting the throttle opening degree and an electric actuator 4c such as a servo motor for driving the valve body 4b. Control is performed by a control output of the control device 20 constituted by a computer or the like.

  On the other hand, a temperature sensor 21 that detects the refrigerant temperature and a pressure sensor 22 that detects the refrigerant pressure are provided on the outlet side of the radiator 2, and detection signals from both the sensors 21 and 22 are input to the control device 20. The control device 20 calculates a target pressure for COP maximization based on the temperature detected by the temperature sensor 21 (heat radiator outlet refrigerant temperature), and the actual high pressure detected by the pressure sensor 22 is calculated as the target. The opening of the valve body 4b of the first pressure reducer 4 (throttle opening) is adjusted by controlling the operation of the electric actuator 4c so that the pressure is reached.

  FIG. 17 shows an example in which the first pressure reducer 4 is configured by an electric valve mechanism in the cycle configuration of the first embodiment, but the first pressure reducer that performs a high-pressure control function also in other embodiments. Of course, 4 may be constituted by an electric valve mechanism as well.

  Further, the installation position of the pressure sensor 22 is not limited to the outlet side of the radiator 2 and may be set anywhere in the high-pressure side refrigerant passage from the discharge side of the compressor 1 to the inlet side of the decompressors 4 and 10. Good.

(17th Embodiment)
In each of the embodiments described above, an example in which the second pressure reducer 10 is configured with a fixed throttle has been described, but the second pressure reducer 10 may be configured with a variable throttle.

  The seventeenth embodiment relates to the second pressure reducer 10 constituted by this variable throttle. 18 (a) and 18 (b) show a second pressure reducer 10 composed of a variable throttle according to the seventeenth embodiment. The second pressure reducer 10 opens the valve according to the differential pressure between the high pressure and the low pressure. It consists of a differential pressure valve that adjusts the degree.

  Specifically, the second decompressor 10 has a cylindrical case 10b that is fitted and fixed in the piping member 10a of the second refrigerant flow path B, and the valve body 10c is placed in the cylindrical case 10b with a case shaft. It is stored so as to be movable in the direction (left-right direction in FIG. 18A).

  As shown in FIGS. 18B and 18C, the valve body 10c has a polygonal body (in the illustrated example, a quadrangle) body, and the corner 10d of the polygon body is the inner peripheral surface of the cylindrical case 10b. It faces through a minute gap. The valve body 10c moves in the case axial direction by a guide action by the corner 10d of the polygon main body and the inner peripheral surface of the cylindrical case 10b.

  And the refrigerant path 10f is formed of the space | gap part between the flat surface 10e between the corner | angular parts of a polygon main-body part, and the internal peripheral surface of the cylindrical case 10b.

  Of the polygonal main body of the valve body 10c, a projecting portion 19g is formed in a truncated cone shape on the end surface on the upstream side (high pressure side) of the refrigerant flow. As shown in FIG. 18 (a), a valve seat portion 10h is integrally formed at an upstream portion of the valve body 10c on the inner peripheral surface of the cylindrical case 10b. The valve seat portion 10h is a small-diameter portion that projects in a stepped manner inward from the inner peripheral surface of the cylindrical case 10b.

  The inner diameter dimension of the valve seat portion 10h is designed to be smaller than the outer dimension of the polygonal main body portion of the valve body 10c so that the conical surface of the protruding portion 19g contacts the valve seat portion 10h. FIG. 18A shows a state in which the protruding portion 19g of the valve body 10c is in contact with the valve seat portion 10h by the spring force of the coiled spring 10i.

  In this state, the refrigerant passage 10f between the flat surface 10e of the polygonal main body of the valve body 10c and the inner peripheral surface of the cylindrical case 10b is fully closed. For this reason, the upstream high-pressure passage 10k and the downstream low-pressure passage 10m communicate with each other only by the throttle passage hole 10j that passes through the central portion of the valve body 10c.

  On the other hand, the differential pressure (high pressure PH-low pressure PL) between the high pressure PH of the high pressure passage 10k and the low pressure PL of the low pressure passage 10m increases, and the force due to this differential pressure causes the coiled spring 10i. When the spring force is overcome, the valve body 10c moves to the left in FIG. 18A, and the protruding portion 19g of the valve body 10c is separated from the valve seat portion 10h.

  As a result, the refrigerant passage 10f on the outer peripheral side of the valve body 10c opens, so that the high-pressure passage 10k and the low-pressure passage 10m communicate with each other through both the throttle passage hole 10j and the refrigerant passage 10f. Therefore, the second pressure reducer 10 functions as a differential pressure valve (variable throttle) in which the throttle opening increases due to an increase in the differential pressure between the high pressure PH and the low pressure PL.

  FIG. 19 shows the refrigerant flow rate adjustment characteristics of the second pressure reducer 10 constituted by the differential pressure valve shown in FIG. 18. The horizontal axis P0 indicates that the protrusion 19g of the valve body 10c is separated from the valve seat 10h ( In a region (1) where the differential pressure between the high pressure PH and the low pressure PL is smaller than the predetermined differential pressure P0, a throttle passage hole is formed between the high pressure passage 10k and the low pressure passage 10m. Since only 10j communicates, the second pressure reducer 10 acts as a fixed throttle.

  On the other hand, in the region (2) where the differential pressure between the high pressure PH and the low pressure PL increases to a predetermined differential pressure P0 or more, the protruding portion 19g of the valve body 10c opens (opens) from the valve seat portion 10h. The high-pressure passage 10k and the low-pressure passage 10m communicate with each other through both the throttle passage hole 10j and the refrigerant passage 10f. Here, the opening area of the refrigerant passage 10f increases as the differential pressure increases. Accordingly, in the region (2), the second pressure reducer 10 acts as a variable throttle.

  In the seventeenth embodiment, the second pressure reducer 10 is configured by a differential pressure valve (variable throttle), but this differential pressure valve is very simple compared to an expansion valve that controls the degree of superheat of the evaporator outlet refrigerant. It can be manufactured at low cost.

  And since there is a correlation that the differential pressure between the high pressure PH and the low pressure PL increases as the cooling load increases, the refrigerant flow rate to the second evaporator 11 can be increased as the cooling load increases.

  Furthermore, what is important is that the differential pressure valve constituting the second pressure reducer 10 is a variable throttle that changes in the direction of decreasing the valve opening when the high pressure PH decreases, so the second pressure reducer 10 is made a variable throttle. Is that the high pressure control function of the first pressure reducer 4 is not adversely affected.

  That is, if the second pressure reducer 10 is a temperature type expansion valve that performs superheat degree control, when the cooling load of the second evaporator 11 increases, the cycle behavior shown in FIG. The high pressure control by 4 becomes unstable and causes hunting. However, since the differential pressure valve that constitutes the second pressure reducer 10 changes in the direction of decreasing the valve opening when the high pressure PH decreases, when the high pressure PH decreases, the decrease in the high pressure PH is suppressed. Acts on direction. On the contrary, when the high pressure PH rises, the valve opening degree of the differential pressure valve constituting the second pressure reducer 10 increases, and therefore acts to suppress the rise of the high pressure PH.

  As a result, even if the second pressure reducer 10 is configured by a differential pressure valve (variable throttle), the high pressure control function by the first pressure reducer 4 is not adversely affected.

  In the seventeenth embodiment, a differential pressure valve is used as a variable throttle constituting the second pressure reducer 10, but the valve opening degree is changed according to the high pressure PH as the variable throttle constituting the second pressure reducer 10. Alternatively, a high-pressure responsive valve to be used or a high-pressure refrigerant temperature responsive valve that changes the valve opening according to the high-pressure refrigerant temperature may be used. Since these high-pressure responsive valves and high-pressure refrigerant temperature responsive valves are variable throttles that change in the direction of decreasing the valve opening when the high-pressure pressure PH decreases, the same effects as the differential pressure valve can be exhibited.

  When the first pressure reducer 4 is configured by a fixed throttle, for example, the first pressure reducer 4 of the fifth embodiment (FIG. 6) is not a fixed throttle but a variable throttle such as the differential pressure valve described in the seventeenth embodiment. You may comprise.

  Similarly, like the first pressure reducer 4 in the third embodiment (FIG. 4), the fourth embodiment (FIG. 5), and the ninth embodiment (FIG. 10), the refrigerant introduced into both the evaporators 5 and 11 is used. The pressure reducer for reducing the pressure may be configured with a variable throttle such as the differential pressure valve described in the seventeenth embodiment instead of the fixed throttle.

(Eighteenth embodiment)
In the seventh embodiment (FIG. 8), the first pressure reducer 4 is a high pressure control type pressure reducer. In the eighteenth embodiment shown in FIG. 20, the first pressure reducer 4 is the seventeenth embodiment. The variable throttle such as the differential pressure valve described with reference to FIG. 18, that is, a variable throttle that changes in the direction of decreasing the valve opening when the high pressure PH decreases. The second decompressor 10 is a fixed throttle.

  In the eighteenth embodiment, both the first pressure reducer 4 and the second pressure reducer 10 may be configured by variable throttles such as differential pressure valves, and both the first pressure reducer 4 and the second pressure reducer 10 are fixed. A diaphragm may be used.

(Nineteenth embodiment)
In the tenth embodiment (FIG. 11), the first pressure reducer 4 is constituted by a high pressure control type pressure reducer, and the temperature sensing part 4 a is provided on the outlet side of the radiator 2 and the high pressure side refrigerant flow of the internal heat exchanger 3. In the nineteenth embodiment shown in FIG. 21, the temperature-sensing part 4a of the first pressure reducer 4 is internally heat exchanged as in the fifteenth embodiment (FIG. 16). It arrange | positions at the exit side of the high voltage | pressure side refrigerant | coolant flow path 3a of the container 3. FIG.

  In the nineteenth embodiment, the second pressure reducer 10 is configured with a fixed throttle or a variable throttle such as the differential pressure valve described in FIG. 18 of the seventeenth embodiment.

(Other embodiments)
In each of the above embodiments, one of the two evaporators 5 and 11 connected in parallel is used for air conditioning on the front side of the vehicle interior, and the other evaporator 11 is used on the rear seat side of the vehicle interior. However, for example, one evaporator 5 may be used for vehicle interior air conditioning, and the other evaporator 11 may be used for cooling a refrigerator mounted in the vehicle interior.

  In addition to CO2, a refrigerant such as ethylene, ethane, or nitrogen oxide may be used as the supercritical cycle refrigerant.

  Further, the present invention may be applied to a normal refrigeration cycle (subcritical cycle) in which the high pressure does not exceed the critical pressure of the refrigerant.

DESCRIPTION OF SYMBOLS 1 Compressor 2 Radiator 3 Internal heat exchanger 4 1st pressure reduction device (pressure reduction means)
5 First evaporator 6 Front seat air conditioning unit 9 Accumulator 10 Second decompressor (pressure reducing means)
11 Second evaporator 12 Rear seat air conditioning unit.

Claims (22)

A compressor (1) for sucking and compressing refrigerant;
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
Among the plurality of evaporators (5, 11), an accumulator (9) provided on the outlet side of one evaporator (5),
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Of the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Further, the one or more decompression means decompress the refrigerant supplied to at least the other evaporator (11) with a fixed throttle (4, 10).   The refrigeration cycle according to claim 1, wherein only one fixed throttle is provided, and the refrigerant supplied to the plurality of evaporators (5, 11) is decompressed by the single fixed throttle.   The one or more decompression means includes a decompression means (4) for decompressing the refrigerant supplied to the one evaporator (5), and the fixed for decompressing the refrigerant supplied to the other evaporator (11). The refrigeration cycle according to claim 1, further comprising an aperture (10) independently. A compressor (1) for sucking and compressing refrigerant;
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
Among the plurality of evaporators (5, 11), an accumulator (9) provided on the outlet side of one evaporator (5),
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Of the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Further, the one or more decompression means is configured to decompress the refrigerant supplied to at least the other evaporator (11) with a variable throttle (10),
The refrigeration cycle characterized in that the variable throttle (10) changes in a direction in which the opening degree decreases when the high pressure decreases.   The variable throttle (10) includes a differential pressure valve that changes its opening according to a differential pressure between a high pressure and a low pressure, a high pressure responsive valve that changes its opening according to a high pressure, and a high pressure refrigerant temperature. The refrigeration cycle according to claim 4, wherein the refrigeration cycle is any one of high-pressure refrigerant temperature responsive valves that change the valve opening.   The one or more decompression means includes a decompression means (4) for decompressing the refrigerant supplied to the one evaporator (5), and the variable for decompressing the refrigerant supplied to the other evaporator (11). 6. A refrigeration cycle according to claim 4 or 5, characterized by comprising an aperture (10) independently.   The pressure reducing means (4) on the side of the one evaporator (5) controls the high pressure according to the high pressure refrigerant temperature on the outlet side of the radiator (2). The refrigeration cycle described in 1. A compressor (1) for sucking and compressing refrigerant;
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
One decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
An accumulator (9) provided on the outlet side of one of the plurality of evaporators (5, 11);
An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). )
Among the plurality of evaporators (5, 11), the outlet side of the other evaporator (11) joins between the outlet side of the accumulator (9) and the inlet side of the internal heat exchanger (3),
Furthermore, the said one pressure reduction means controls a high pressure according to the high pressure refrigerant | coolant temperature by the side of the said heat radiator (2), The refrigerating cycle characterized by the above-mentioned. A compressor (1) for sucking and compressing refrigerant;
A radiator (2) for cooling the refrigerant discharged from the compressor (1);
At least one decompression means (4, 10) for decompressing the refrigerant on the outlet side of the radiator (2);
A plurality of evaporators (5, 11) for evaporating the low-pressure refrigerant after passing through the decompression means (4, 10);
An accumulator (9) provided between the outlet side of the plurality of evaporators (5, 11) and the suction side of the compressor (1),
A refrigeration cycle, comprising: a refrigerant passage through which a refrigerant exiting at least one of the plurality of evaporators (5, 11) is introduced to an inlet side of the other evaporator.   10. The refrigeration cycle according to claim 9, wherein the decompression unit includes a single decompression unit (4) that decompresses the refrigerant flowing into the plurality of evaporators (5, 11).   The refrigeration cycle according to claim 9, wherein a plurality of decompression means (4, 10) are provided as the decompression means corresponding to the plurality of evaporators (5, 11), respectively.   12. The pressure reducing means (4) of the plurality of pressure reducing means controls the high pressure according to the high pressure refrigerant temperature on the outlet side of the radiator (2). Refrigeration cycle.   The refrigeration cycle according to claim 12, wherein the other decompression means (10) of the plurality of decompression means is a fixed throttle. The other decompression means of the plurality of decompression means is a variable throttle (10),
The refrigeration cycle according to claim 12, wherein the variable throttle (10) changes in a direction in which the opening degree decreases when the high pressure decreases.   The refrigeration cycle according to claim 11, wherein at least one of the plurality of decompression means (4, 10) is a fixed throttle (10). At least one of the plurality of decompression means (4, 10) is a variable throttle (10),
The refrigeration cycle according to claim 11, wherein the variable throttle (10) changes in a direction in which the opening degree decreases when the high pressure decreases.   The variable throttle (10) includes a differential pressure valve that changes its opening according to a differential pressure between a high pressure and a low pressure, a high pressure responsive valve that changes its opening according to a high pressure, and a high pressure refrigerant temperature. The refrigeration cycle according to claim 14 or 16, wherein the refrigeration cycle is any one of high-pressure refrigerant temperature responsive valves that change the valve opening.   An internal heat exchanger (3) that exchanges heat between the low-pressure refrigerant between the outlet side of the accumulator (9) and the suction side of the compressor (1) and the high-pressure refrigerant on the outlet side of the radiator (2). The refrigeration cycle according to any one of claims 9 to 17, further comprising:   The on-off valve (16) for stopping the refrigerant flow to at least one of the plurality of evaporators (5, 11) is provided. Refrigeration cycle.   The refrigerating cycle according to any one of claims 1 to 18, further comprising a switching valve (17) for switching a refrigerant flow to the plurality of evaporators (5, 11).   The refrigeration cycle according to any one of claims 1 to 20, wherein the refrigeration cycle is a supercritical refrigeration cycle using a refrigerant having a high pressure equal to or higher than a critical pressure. A vehicle air conditioner comprising the refrigeration cycle according to any one of claims 1 to 21,
Among the plurality of evaporators (5, 11), one of the evaporators (5) is disposed in the air conditioning unit (6) on the front seat side of the vehicle interior,
Of the plurality of evaporators (5, 11), the other evaporator (11) is disposed in the air conditioning unit (12) on the rear seat side of the vehicle interior.

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