JP4915250B2 - Ejector refrigeration cycle - Google Patents

Description

  The present invention relates to an ejector-type refrigeration cycle having an ejector.

  Conventionally, in Patent Documents 1 and 2, a branching part for branching the refrigerant flow is arranged downstream of the ejector, and one of the branched refrigerants flows into the first evaporator, and the outlet side of the first evaporator is The other refrigerant that is connected to the suction side of the compressor and further branched is allowed to flow into the second evaporator via a decompression device such as an expansion valve, and the outlet side of the second evaporator is connected to the refrigerant suction port of the ejector. An ejector-type refrigeration cycle connected to is disclosed.

  In the cycle of Patent Document 1, a branch portion is configured inside the gas-liquid separator disposed on the downstream side of the ejector, and the liquid phase refrigerant is caused to flow into the second evaporator. In the cycle of Patent Document 2, A bifurcated portion is formed by a U-shaped distributor having a three-way joint structure, and by allowing gas-liquid two-phase refrigerant or liquid-phase refrigerant to flow into the second evaporator, both evaporators can exhibit refrigeration capacity. Yes.

Furthermore, in the cycle of Patent Documents 1 and 2, the refrigerant boosted by the pressure increasing action of the diffuser portion of the ejector is sucked into the compressor through the first evaporator, so that the driving power of the compressor is reduced and the cycle is performed. The efficiency (COP) is improved.
Japanese Patent No. 2818965 JP 2006-292351 A

  However, in the cycles of Patent Documents 1 and 2, since the decompression device is disposed between the branching portion and the second evaporator, a loss occurs in the kinetic energy of the refrigerant when passing through the decompression device. Therefore, when all of this kinetic energy is lost, the refrigerant on the outlet side of the decompression device is second evaporated only by the action of the pressure difference between the static pressure of the refrigerant on the outlet side of the decompression device and the static pressure of the refrigerant suction port of the ejector. Must flow into the vessel.

  Therefore, if the pressure loss between the inlet and outlet of the second evaporator becomes higher than the pressure difference between the static pressure of the refrigerant on the outlet side of the decompression device and the static pressure of the refrigerant at the refrigerant suction port of the ejector, the second evaporator It becomes impossible to flow in the refrigerant. As a result, in the cycles of Patent Documents 1 and 2, the second evaporator may not be able to exhibit the refrigerating capacity.

  In view of this, the inventors of the present invention previously branched the flow of the refrigerant at a branch portion disposed downstream of the ejector in Japanese Patent Application No. 2006-292437 (hereinafter referred to as the prior application example). One refrigerant flows into the first evaporator, the outlet side of the first evaporator is connected to the suction side of the compressor, and the other refrigerant flows directly into the second evaporator without going through the decompression device. A cycle is proposed in which the outlet side of the second evaporator is connected to the refrigerant suction port of the ejector.

  In the cycle of this prior application, as in the cycles of Patent Documents 1 and 2, the refrigeration capacity can be exhibited by both evaporators, and the driving power of the compressor can be reduced to improve the cycle efficiency.

  Further, since the refrigerant branched at the branching portion is directly flowed into the second evaporator without passing through the decompression device or the like, the static pressure of the refrigerant on the outlet side of the expansion device and the static pressure of the refrigerant at the refrigerant suction port of the ejector The refrigerant on the downstream side of the ejector can be caused to flow into the second evaporator not only by the action of the pressure difference but also by the action of the dynamic pressure of the refrigerant flowing out of the ejector. As a result, the refrigerating capacity can be surely exhibited in the second evaporator.

  In the cycle of the prior application example, since the refrigerant flowing out from the ejector is directly flowed into the second evaporator, the dryness of the refrigerant flowing into the second evaporator is higher than the cycle of Patent Documents 1 and 2. Prone. For this reason, the enthalpy of the refrigerant on the inlet side of the second evaporator is also increased, and the refrigerating capacity that can be exhibited by the second evaporator is likely to be reduced.

  Therefore, in the cycle of the prior application example, for example, by using an internal heat exchanger, the enthalpy of the refrigerant on the upstream side of the nozzle part of the ejector is decreased, and the decrease in the refrigerating capacity of the second evaporator is suppressed.

  However, if the refrigerant flowing into the nozzle portion becomes supercooled due to the decrease in enthalpy, the refrigerant is less likely to boil at the nozzle portion, and the nozzle efficiency is lowered. The nozzle efficiency is energy conversion efficiency when converting the pressure energy of the refrigerant into kinetic energy. Therefore, in order to improve the nozzle efficiency, it is desirable to allow a refrigerant in a saturated liquid phase state or a gas-liquid two-phase state to flow into the nozzle portion.

  And if nozzle efficiency falls, the flow velocity of the refrigerant | coolant flow injected from a nozzle part cannot fully be accelerated, but the collection | recovery energy amount of an ejector will fall. For this reason, the amount of pressure increase in the diffuser portion of the ejector decreases, and it becomes difficult to obtain the effect of improving the cycle efficiency by reducing the driving power of the compressor.

This will be described in detail with reference to the Mollier diagram of FIG. FIG. 11 schematically shows the state of the refrigerant during the operation of the ejector refrigeration cycle of the prior application example. First, the solid line in FIG. 11 shows a case where the refrigerant flowing into the nozzle portion is operated so as to be in a gas-liquid two-phase state. Discharged from the compressor refrigerant (A _G point of Fig. 11) is being radiated so that the gas-liquid two-phase state by the radiator or the like (A _G point of Fig. 11 → B _G point), to the nozzle unit Inflow.

The refrigerant flowing into the nozzle portion is isentropically to expand (B _G point of Fig. 11 → C _G point), second mixing with the evaporator downstream refrigerant (C _G point of Fig. 11 → D _G point), boosted by the diffuser part of the ejector (D _G point of Fig. 11 → E _G point), it is branched by the branch portion. Thus, in FIG. 11, amount of energy recovered ejector can be represented by [Delta] I _G, boost amount in the diffuser section can be expressed by [Delta] P _G.

One of the refrigerants branched at the branching part exhibits an endothermic effect in the first evaporator (E _G point → F _G point in FIG. 11), and is pressurized again by the compressor (F in FIG. 11). _G point → A _G point). The other refrigerant exhibits an endothermic effect while reducing the pressure due to the pressure loss of the second evaporator (E _G point → G _G point in FIG. 11), and is sucked into the refrigerant suction port of the ejector (FIG. 11). _GG point → _DG point). Therefore, in FIG. 11, the total value of the refrigerating capacity in the first and second evaporators can be expressed as ΔH _G 1 + ΔH _G 2.

Next, the broken line in FIG. 11 shows a case where the refrigerant flowing into the nozzle portion is operated so as to be in a supercooled state. That is, the refrigerant discharged from the compressor (A _L point in FIG. 11) is cooled until it reaches a supercooled state (A _L point → B _L point in FIG. 11), and flows into the nozzle portion. In addition, in the broken line of FIG. 11, the code | symbol which shows the state of the refrigerant | coolant corresponding to the above-mentioned solid line has each changed the subscript from G to L, and has shown it.

When the refrigerant flowing into the nozzle portion is in a supercooled state (point B _{L in} FIG. 11), the refrigerant flowing into the nozzle portion is in a gas-liquid two-phase state (point _{BG in} FIG. 11) Since the enthalpy of the refrigerant on the upstream side of the nozzle portion becomes smaller than that, the total value of the refrigerating capacity in the first and second evaporators can be increased as shown in FIG. 11 (in FIG. 11, ΔH _G 1 + ΔH _G 2 → ΔH _L 1 + ΔH _L 2).

However, if the refrigerant flowing into the nozzle portion is in a supercooled state, the nozzle efficiency of the nozzle portion is lowered, and the refrigerant expands under reduced pressure as B _L → C _L in FIG. For this reason, the amount of energy recovered by the ejector decreases (ΔI _G → ΔI _{L in} FIG. 11), and the pressure increase amount ΔP in the diffuser section decreases (ΔP _G → ΔP _{L in} FIG. 11). As a result, it becomes difficult to obtain the effect of improving the cycle efficiency by reducing the driving power of the compressor.

  In other words, in the cycle of the prior application example, even if the enthalpy of the refrigerant upstream of the nozzle portion is reduced and the total value of the refrigeration capacity in the first and second evaporators is increased, the cycle efficiency is improved by reducing the driving power of the compressor. Since it becomes difficult to obtain the effect, the cycle efficiency of the entire cycle may be reduced instead.

  Therefore, in order to operate the cycle of the prior application example while exhibiting high cycle efficiency, the total value of the refrigeration capacities in the first and second evaporators is increased so as not to lower the nozzle efficiency of the nozzle portion. There is a need.

  In view of the above points, the present invention provides an ejector-type refrigeration cycle in which a refrigerant flows into a plurality of evaporators from a branch portion disposed on the downstream side of the ejector. The purpose is to increase the total value of the refrigerating capacity and to exhibit high cycle efficiency as a whole cycle.

In order to achieve the above object, in the present invention, a compressor (11) that compresses and discharges a refrigerant, a radiator (12) that radiates heat from a high-pressure refrigerant compressed by the compressor (11), and a radiator (12) A decompression means (14, 30, 32) for decompressing and expanding the downstream side refrigerant, and a high pressure jetted from the nozzle (15a) for further decompressing and expanding the refrigerant decompressed by the decompression means (14, 30, 32). The ejector (15) that sucks the refrigerant from the refrigerant suction port (15b) by the speed refrigerant flow, the branch (100) that branches the flow of the refrigerant downstream of the ejector (15), and the branch (100) are branched. The first evaporator (17) that evaporates the other refrigerant and flows out to the compressor (11) suction side, and the other refrigerant branched by the branch section (100) evaporates and the refrigerant suction port (15b) The second evaporator (1 ) And equipped with a,
The branch section (100) includes an introduction pipe section (16a) for allowing the refrigerant on the downstream side of the ejector (15) to flow in, and a first outlet pipe section for allowing the refrigerant in the introduction pipe section (16a) to flow to the first evaporator (17) side. (16b) and a second lead-out pipe part (16c) that causes the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch section (100), the introduction pipe section (16a) and the second outlet pipe section (16c) are in the same direction as the refrigerant inflow direction of the introduction pipe section (16a) and the refrigerant outlet direction of the second outlet pipe section (16c). Arranged to face
The branching unit (100), the ejector (15) second evaporator inlet tube portion as the dynamic pressure of the outflow refrigerant is maintained from (16a) passes through the second outlet pipe portion (16c) from ( 18) Branch the refrigerant flow to the side ,
The decompression means (14, 30, 32) is such that the state of the refrigerant flowing into the nozzle portion (15a) is less than a predetermined reference supercooling degree (SSc) or lower than a predetermined reference dryness degree (SGr). An ejector-type refrigeration cycle in which the refrigerant is expanded under reduced pressure so as to be in a gas-liquid two-phase state is a first feature.

  According to this, in order to increase the total value of the refrigerating capacity of the first and second evaporators (17, 18), until the enthalpy of the refrigerant on the upstream side of the decompression means (14, 30, 32) is in a supercooled state. Even if it is decreased, the refrigerant expanded under reduced pressure so as to be in a supercooled state below a predetermined reference supercooling degree (SSc) or a gas-liquid two-phase state below a predetermined reference dryness degree (SGr) ( 15a).

  That is, since the refrigerant in a state close to the saturated liquid phase state can flow into the nozzle portion (15a), boiling of the refrigerant in the nozzle portion is easily promoted, and the nozzle efficiency is not reduced. As a result, the total value of the refrigeration capacities of the first and second evaporators (17, 18) can be increased without causing a decrease in nozzle efficiency. As a result, high cycle efficiency can be exhibited as a whole cycle.

  Note that the degree of supercooling indicates the absolute value of the amount of temperature decrease from the saturation temperature corresponding to the condensing pressure of the refrigerant. Therefore, when the degree of supercooling is close to 0, the absolute temperature of the refrigerant increases. The saturated liquid phase is approached. On the other hand, the dryness indicates the ratio of the vapor (gas phase refrigerant) in the wet steam (gas-liquid two-phase refrigerant), so that the dryness close to 0 indicates a saturated liquid phase state. Get closer.

  In the ejector refrigeration cycle having the first feature, the pressure reducing means may be constituted by an electric variable throttle mechanism (14). Further, refrigerant state detection means (21 to 24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle portion (15a), and throttle mechanism control means (20c) for controlling the operation of the variable throttle mechanism (14). The throttle mechanism control means (20c) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state below the reference supercooling degree (SSc) or below the reference dryness degree (SGr) The operation of the variable throttle mechanism (14) may be controlled so as to be in a state.

  According to this, the refrigerant in a state close to the saturated liquid phase state can be more reliably flowed into the nozzle portion (15a).

  In the ejector refrigeration cycle of the first feature, the decompression means may be constituted by a variable throttle mechanism (30) whose throttle opening is changed by a mechanical mechanism. More specifically, the variable throttle mechanism (30) may be a temperature type expansion valve.

  In the ejector-type refrigeration cycle having the first feature, the pressure reducing means may be constituted by a fixed throttle mechanism (32).

  Furthermore, refrigerant state detection means (21-24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle section (15a), and discharge capacity adjusting means (11a) for adjusting the refrigerant discharge capacity of the compressor (11). ) And a discharge capacity control means (20a) for controlling the operation of the discharge capacity adjusting means (11a). The discharge capacity control means (20a) has a reference supercooling state of the refrigerant flowing into the nozzle portion (15a). The operation of the discharge capacity adjusting means (11a) may be controlled so that the supercooled state is less than the degree (SSc) or the gas-liquid two-phase state is less than the reference dryness degree (SGr).

  Furthermore, refrigerant state detection means (21 to 24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle section (15a), and heat radiation capacity adjusting means (12a) for adjusting the heat radiation capacity of the radiator (12). And a heat radiation capacity control means (20b) for controlling the operation of the heat radiation capacity adjustment means (12a), and the heat radiation capacity control means (20b) is configured such that the state of the refrigerant flowing into the nozzle portion (15a) is a reference supercooling degree. (SSc) The operation of the heat radiation capacity adjusting means (12a) may be controlled so as to be in the supercooled state below or the gas-liquid two-phase state below the reference dryness (SGr).

  According to this, the refrigerant in a state close to the saturated liquid phase state can be more reliably flowed into the nozzle portion (15a).

Further, in the present invention, the compressor (11) that compresses and discharges the refrigerant, the radiator (12) that radiates the high-pressure refrigerant compressed by the compressor (11), and the downstream refrigerant of the radiator (12) Decompression means (33a) for decompressing and expanding, refrigerant heat radiation means (33) for dissipating the refrigerant in the decompression and expansion process in the decompression means (33a), and nozzle for further decompressing and expanding the refrigerant decompressed by the decompression means (33a) An ejector (15) that sucks the refrigerant from the refrigerant suction port (15b) by a high-speed refrigerant flow injected from the section (15a), a branch unit (100) that branches the flow of the refrigerant on the downstream side of the ejector (15), and a branch The first evaporator (17) that evaporates one refrigerant branched in the section (100) and flows out to the compressor (11) suction side, and the other refrigerant branched in the branch section (100) evaporates Let the refrigerant suction port Comprises 15b) second evaporator flow out to the upstream side and (18),
The branch section (100) includes an introduction pipe section (16a) for allowing the refrigerant on the downstream side of the ejector (15) to flow in, and a first outlet pipe section for allowing the refrigerant in the introduction pipe section (16a) to flow to the first evaporator (17) side. (16b) and a second lead-out pipe part (16c) that causes the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch section (100), the introduction pipe section (16a) and the second outlet pipe section (16c) are in the same direction as the refrigerant inflow direction of the introduction pipe section (16a) and the refrigerant outlet direction of the second outlet pipe section (16c). Arranged to face
The branching unit (100), the ejector (15) second evaporator inlet tube portion as the dynamic pressure of the outflow refrigerant is maintained from (16a) passes through the second outlet pipe portion (16c) from ( 18) Branch the refrigerant flow to the side ,
The decompression means (33a) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state equal to or lower than a predetermined reference supercooling degree (SSc) or a predetermined reference dryness (SGr). An ejector refrigeration cycle that decompresses and expands the refrigerant so as to be in a state is a second feature.

  According to this, like the ejector refrigeration cycle of the first feature, the total value of the refrigeration capacities of the first and second evaporators (17, 18) can be increased without causing a decrease in nozzle efficiency. it can. As a result, high cycle efficiency can be exhibited as a whole cycle.

Additionally, the refrigerant heat dissipating means (33) is, therefore dissipates refrigerant decompression expansion process in the decompression means (33a), for example, B ₅ points Mollier diagram of FIG. 6 to be described in embodiments described later → B _'5 points As shown in FIG. 5, the refrigerant enthalpy can be reduced at the same time as the pressure of the refrigerant is reduced.

  That is, without reducing the enthalpy of the refrigerant on the upstream side of the decompression means (33a), by radiating the refrigerant in the decompression and expansion process by the decompression means (33a), the enthalpy of the downstream side refrigerant of the decompression means (33a) is reduced, The total value of the refrigerating capacity of the first and second evaporators (17, 18) can be increased.

  In the ejector refrigeration cycle of the second feature, the refrigerant heat dissipating means includes an internal heat exchanger (33) for exchanging heat between the radiator (12) downstream refrigerant and the compressor (11) suction refrigerant. In addition, the decompression means (33a) may form a high-pressure side refrigerant flow path for allowing the refrigerant on the downstream side of the radiator (12) to pass through the internal heat exchanger (33).

  According to this, heat can be exchanged between the refrigerant passing through the decompression means (33a) and the refrigerant sucked by the compressor (11), so that the refrigerant heat dissipation means can be easily configured. Furthermore, since the decompression means (33a) can be integrated into the internal heat exchanger (33) as a high-pressure side refrigerant flow path of the internal heat exchanger (33), the cycle can be reduced in size.

  Furthermore, specifically, the pressure reducing means may be a capillary tube (33a). Since the capillary tube depressurizes the refrigerant by the effect of reducing the area of the refrigerant passage and the frictional force in the refrigerant passage, the capillary tube has an elongated shape having a predetermined refrigerant passage length.

  Therefore, by adopting a capillary tube as the decompression means (14, 30, 32), it becomes easy to secure a heat exchange area when heat exchange is performed between the capillary tube passing refrigerant and the compressor (11) suction side refrigerant, The refrigerant passing through the capillary tube can be efficiently radiated.

  In the ejector refrigeration cycle having the second feature, the refrigerant state detecting means (21 to 24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle portion (15a), and the compressor (11 ), And the discharge capacity control means (20a) is a supercooling state in which the state of the refrigerant flowing into the nozzle portion (15a) is below the reference supercooling degree (SSc). Or you may come to control the action | operation of a compressor (11) so that it may become a gas-liquid two-phase state below reference | standard dryness (SGr).

  According to this, the refrigerant in a state close to the saturated liquid phase refrigerant can be more surely flowed into the nozzle portion (15a), and the nozzle efficiency is not unnecessarily lowered.

  By the way, the present inventors investigated and examined the details of the relationship between the state of the refrigerant flowing into the nozzle portion (15a) and the cycle efficiency in a cycle having the same configuration as that of the first embodiment described later. The survey results shown are obtained.

  The horizontal axis in FIG. 12 employs the degree of supercooling and the degree of dryness of the refrigerant flowing into the nozzle portion (15a) as an index indicating the state of the refrigerant. The degree of supercooling 0 (K) and the degree of dryness 0 mean a saturated liquid phase state. The vertical axis in FIG. 12 represents the cycle efficiency (COP) improvement rate with respect to the cycle efficiency of the prior application example. Therefore, a COP improvement rate of 1 or more means an improvement in cycle efficiency.

  As is clear from FIG. 12, if the refrigerant flowing into the nozzle portion (15a) is in a supercooled state with a supercooling degree of 10K or less or a gas-liquid two-phase state with a dryness degree of 0.2 or less, the prior application example Cycle efficiency can be improved with respect to this cycle. Therefore, the refrigerant in the above state is a refrigerant in a state where the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency.

  Furthermore, if the supercooling state is 5K or less or the gas-liquid two-phase state is 0.1 or less, the cycle efficiency can be greatly improved by 30% or more compared to the cycle of the prior application example. .

  Therefore, in the ejector-type refrigeration cycle having the first and second features described above, the reference supercooling degree (SSc) may be 10K, and the reference dryness degree (SGr) may be 0.2. Further, preferably, the reference supercooling degree (SSc) is 5K, and the reference dryness degree (SGr) may be 0.1.

Further, in the present invention, the compressor (11) that compresses and discharges the refrigerant, the radiator (12) that radiates the high-pressure refrigerant compressed by the compressor (11), and the downstream refrigerant of the radiator (12) An ejector (35) that sucks the refrigerant from the refrigerant suction port (35b) by a high-speed refrigerant flow injected from the nozzle portion (35a) that decompresses and expands, and a branching portion that branches the flow of the refrigerant on the downstream side of the ejector (35) ( 100), the first evaporator (17) that evaporates one refrigerant branched at the branch portion (100) and flows out to the compressor (11) suction side, and is branched at the branch portion (100) A second evaporator (18) that evaporates the other refrigerant and flows out to the upstream side of the refrigerant suction port (35b);
The branch section (100) includes an introduction pipe section (16a) for allowing the refrigerant on the downstream side of the ejector (15) to flow in, and a first outlet pipe section for allowing the refrigerant in the introduction pipe section (16a) to flow to the first evaporator (17) side. (16b) and a second lead-out pipe part (16c) that causes the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch section (100), the introduction pipe section (16a) and the second outlet pipe section (16c) are in the same direction as the refrigerant inflow direction of the introduction pipe section (16a) and the refrigerant outlet direction of the second outlet pipe section (16c). Arranged to face
The branching unit (100), the ejector (15) second evaporator inlet tube portion as the dynamic pressure of the outflow refrigerant is maintained from (16a) passes through the second outlet pipe portion (16c) from ( 18) Branch the refrigerant flow to the side ,
The refrigerant passage (35f) in the nozzle portion (35a) is formed with first and second throttle portions (35g, 35h) configured to restrict the cross-sectional area of the passage, and the first throttle portion (35g) The ejector-type refrigeration cycle, which is disposed upstream of the two throttle portions (35h) in the refrigerant flow direction and depressurizes the refrigerant so as to be in a gas-liquid two-phase state, is a third feature.

  According to this, in order to increase the total value of the refrigerating capacity of the first and second evaporators (17, 18), the first throttle part ( The refrigerant expanded under reduced pressure until it becomes a gas-liquid two-phase state at 35 g) can be further expanded under reduced pressure at the second throttle part (35h). As a result, the boiling of the refrigerant in the second throttle part (35h) is promoted, and the nozzle efficiency of the entire nozzle part (35a) does not decrease.

  As a result, as in the ejector refrigeration cycle having the first and second characteristics, the total value of the refrigeration capacities of the first and second evaporators (17, 18) is increased without causing a decrease in nozzle efficiency. Can do. As a result, high cycle efficiency can be exhibited as a whole cycle.

  In the ejector-type refrigeration cycle of the third feature, specifically, the refrigerant passage (35f) gradually has a passage sectional area from the refrigerant inlet side of the nozzle portion (35a) to the first throttle portion (35g). It may be formed in a shape that decreases and gradually decreases again after the passage sectional area from the first throttle portion (35g) to the second throttle portion (35h) gradually increases.

  According to this, it is possible to easily form the shape of the refrigerant passage (35f) of the nozzle part (35a) that generates boiling nuclei at the first throttle part (35g) and boiles at the second throttle part (35h).

  In the ejector refrigeration cycle having the third feature described above, the ejector is a variable ejector (35) having throttle opening adjusting means (35k, 35l) for adjusting the throttle opening of the first throttle section (35g). There may be. Furthermore, the throttle opening degree control means (20d) for controlling the operation of the throttle opening degree adjusting means (35k, 35l) is provided, and the throttle opening degree control means (20d) has a refrigerant flowing into the second throttle portion (35h). The operation of the throttle opening adjustment means (35k, 35l) may be controlled so that the gas-liquid two-phase state is obtained.

  According to this, the refrigerant in the gas-liquid two-phase state can be more reliably flowed into the second throttle part (35h), and the nozzle efficiency of the entire nozzle part (15a) is not lowered.

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

(First embodiment)
A first embodiment of the present invention will be described with reference to FIGS. FIG. 1 is an overall configuration diagram of an example in which an ejector refrigeration cycle 10 of the present invention is applied to a vehicle air conditioner. First, in the ejector refrigeration cycle 10, the compressor 11 sucks, compresses and discharges refrigerant, and is driven to rotate by a driving force transmitted from a vehicle traveling engine (not shown) via a pulley and a belt. The

  In this embodiment, a swash plate type variable displacement compressor capable of continuously variably controlling the discharge capacity by a control signal from the outside is adopted as the compressor 11. Further, in the swash plate type variable displacement compressor, the refrigerant discharge capacity can be changed by changing the discharge capacity. The discharge capacity is the geometric volume of the working space where the refrigerant is sucked and compressed, that is, the cylinder volume between the top dead center and the bottom dead center of the piston stroke.

  Specifically, this swash plate type variable displacement compressor is formed in a swash plate chamber (not shown) that introduces suction refrigerant and discharge refrigerant, and the ratio of suction refrigerant and discharge refrigerant that is introduced into the swash plate chamber. The electromagnetic capacity control valve 11a for adjusting the swash plate and a swash plate (not shown) that displaces the tilt angle according to the pressure of the swash plate chamber, and the piston stroke (discharge capacity) according to the tilt angle of the swash plate ) Is changed.

  The electromagnetic capacity control valve 11a constitutes the discharge capacity adjusting means of the present embodiment, and the valve opening degree (ratio between the suction refrigerant and the discharge refrigerant) is adjusted by the control current of the air conditioning control device 20 described later. Further, in the swash plate type variable displacement compressor, the discharge capacity can be continuously changed in a range of approximately 0% to 100%. Therefore, by reducing the discharge capacity to approximately 0%, the compressor 11 can be reduced. It can be substantially deactivated.

  Therefore, in the present embodiment, the compressor 11 is always configured to be coupled to the vehicle running engine via the pulley and the belt V. Of course, even a variable displacement compressor may be configured such that power can be transmitted from the vehicle running engine via an electromagnetic clutch.

  When a fixed capacity compressor is employed as the compressor 11, an electromagnetic clutch can be employed as the discharge capacity adjusting means. That is, the refrigerant discharge capacity can be adjusted by operating the compressor intermittently with the electromagnetic clutch and performing the operation rate control for controlling the ratio of the on / off operation. Furthermore, if an electric compressor is employed as the compressor 11, the electric motor serves as a discharge capacity adjusting means, and the refrigerant discharge capacity can be adjusted by adjusting the rotation speed of the electric motor.

  A radiator 12 is connected to the discharge side of the compressor 11. The radiator 12 is a heat-dissipating heat exchanger that performs heat exchange between the high-temperature and high-pressure refrigerant discharged from the compressor 11 and the outside air (air outside the passenger compartment) blown by the cooling fan 12a to radiate the high-pressure refrigerant.

  The cooling fan 12a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from an air conditioning control device 20 described later. And the heat dissipation capability of the radiator 12 is adjusted by the amount of air blown by the cooling fan 12a. Therefore, the cooling fan 12a constitutes the heat radiation capacity adjusting means of this embodiment.

  In the ejector refrigeration cycle 10 of this embodiment, a chlorofluorocarbon refrigerant is employed as the refrigerant, and a subcritical cycle in which the high-pressure side refrigerant pressure does not exceed the critical pressure of the refrigerant is configured. Therefore, the radiator 12 acts as a condenser that condenses the refrigerant.

  A high-pressure side refrigerant flow path 13 a of the internal heat exchanger 13 is connected to the downstream side of the radiator 12. The internal heat exchanger 13 exchanges heat between the refrigerant on the outlet side of the radiator 12 that passes through the high-pressure side refrigerant flow path 13a and the refrigerant on the suction side of the compressor 11 that passes through the low-pressure side refrigerant flow path 13b, and then exits the radiator 12 The side refrigerant is cooled. Thereby, the enthalpy difference (refrigeration capacity) of the refrigerant | coolant between the refrigerant | coolant inlet_port | entrance and outlet in the 1st and 2nd evaporators 17 and 18 mentioned later can be increased.

  Various configurations can be adopted as a specific configuration of the internal heat exchanger 13. Specifically, the refrigerant pipes forming the high-pressure side refrigerant flow path 13a and the low-pressure side refrigerant flow path 13b are brazed, joined by a joining means such as welding, pressure welding, soldering, etc., and heat exchange is performed. A double-pipe heat exchanger configuration in which the low-pressure side refrigerant flow path 13b is arranged inside the outer pipe forming the side refrigerant flow path 13a can be adopted.

  A variable throttle mechanism 14 is connected to the outlet side of the high-pressure side refrigerant flow path 13 a of the internal heat exchanger 13. The variable throttle mechanism 14 is a decompression unit that decompresses and expands the high-pressure liquid-phase refrigerant so as to be in a predetermined state, and causes the high-pressure liquid phase refrigerant to flow out downstream (specifically, the nozzle portion 15a side of an ejector 15 described later). It is also a flow rate adjusting means for adjusting the flow rate of the refrigerant.

  The variable throttle mechanism 14 is an electric variable throttle mechanism in which the opening degree of the throttle passage is adjusted by a control signal output from an air conditioning control device 20 described later. Specifically, it has a drive means constituted by a stepping motor or the like, and the displacement amount of the valve body is adjusted by this drive means, and the throttle opening (refrigerant flow rate) is adjusted by this valve body.

  An ejector 15 is connected to the downstream side of the variable throttle mechanism 14. The ejector 15 is a decompression unit that decompresses the refrigerant, and is also a refrigerant circulation unit that circulates the refrigerant by a suction action of a refrigerant flow injected at a high speed.

  Specifically, the ejector 15 squeezes the passage area of the intermediate-pressure refrigerant flowing out from the variable throttle mechanism 14 to be small, and further communicates with the nozzle portion 15a that further depressurizes the refrigerant and the refrigerant injection port 15e of the nozzle portion 15a. And a refrigerant suction port 15b for sucking the refrigerant that has flowed out of the second evaporator 18 described later.

  Furthermore, a mixing portion 15c for mixing the high-speed refrigerant flow injected from the nozzle portion 15a and the suctioned refrigerant sucked from the refrigerant suction port 15b is provided at the downstream side of the refrigerant flow of the nozzle portion 15a and the refrigerant suction port 15b. And has a diffuser portion 15d forming a pressure increasing portion on the downstream side of the refrigerant flow of the mixing portion 15c.

  The diffuser portion 15d is formed in a shape that gradually increases the refrigerant passage area, and acts to decelerate the refrigerant flow to increase the refrigerant pressure, that is, to convert the velocity energy of the refrigerant into pressure energy.

  A refrigerant distributor 16 that branches the flow of the refrigerant and distributes it downstream is connected to the downstream side of the ejector 15 (specifically, the outlet side of the diffuser portion 15d). The refrigerant distributor 16 has a T-shaped three-way joint structure formed by joining substantially linear pipes having different pipe diameters.

  More specifically, the refrigerant distributor 16 allows the refrigerant to flow into the introduction pipe part 16a, the first outlet pipe part 16b that causes the refrigerant to flow out to the first evaporator 17 side, and the refrigerant to flow out to the second evaporator 18 side. A second outlet pipe portion 16c is provided. Therefore, the branch part 100 is formed inside the refrigerant distributor 16.

  Furthermore, in the refrigerant distributor 16, the refrigerant inflow direction of the introduction pipe part 16a and the refrigerant outflow direction of the second outlet pipe part 16c are coaxially oriented in the same direction, and the refrigerant outlet direction of the first outlet pipe part 16b is The direction is substantially perpendicular to the refrigerant inflow direction of the introduction pipe portion 16a and the refrigerant outflow direction of the second outlet pipe portion 16c.

  As described above, the flow direction of the refrigerant flowing into the introduction pipe section 16a and the flow direction of the refrigerant flowing out of the second outlet pipe section 16c are coaxially directed in the same direction, so that the refrigerant flowing into the introduction pipe section 16a is It flows out from the 2nd outlet pipe part 16c, without reducing a flow velocity as needed. Thereby, when the flow of the refrigerant is branched in the branch part 100 of the refrigerant distributor 16, the dynamic pressure of the refrigerant flowing out of the diffuser part 15d of the ejector 15 is maintained.

  Such a refrigerant distributor 16 can be easily formed by joining metal pipes by joining means such as brazing and welding. Of course, it may be formed by bonding resin piping. Furthermore, you may form by providing a some refrigerant path in a rectangular parallelepiped metal block or resin block.

  In FIG. 1, the cycle configuration of the present embodiment is schematically shown. However, the diffuser portion 15d, the refrigerant distributor 16, the first evaporator 17, and the second evaporator 18 of the ejector 15 are directly or shortly piped. It is desirable to connect them closer to each other. By connecting in this way, the dynamic pressure of the refrigerant flowing out of the ejector 15 is maintained when the refrigerant flow is further branched.

  In the present embodiment, since the refrigerant distributor 16 and the second evaporator 18 are connected as described above, the second outlet pipe portion 16c and the second evaporator 18 of the refrigerant distributor 16 are provided with a throttle means. Are connected so that the dynamic pressure of the refrigerant flowing out of the diffuser portion 15 d of the ejector 15 acts on the inside of the second evaporator 18.

  The first evaporator 17 connected to the first outlet pipe portion 16b heat-exchanges one of the refrigerant branched by the refrigerant distributor 16 and the air blown by the blower fan 17a to evaporate the low-pressure refrigerant. This is an endothermic heat exchanger that exerts an endothermic effect. The cooling fan 12a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from an air conditioning control device 20 described later.

  Furthermore, an accumulator 19 is connected to the outlet side of the first evaporator 17. The accumulator 19 is a gas-liquid separator that separates the gas-phase refrigerant and liquid-phase refrigerant that have flowed into the accumulator 19 and stores excess refrigerant. The gas phase refrigerant outlet of the accumulator 19 is connected to the inlet side of the low-pressure side refrigerant flow path 13b of the internal heat exchanger 13, and the refrigerant suction side of the compressor 11 is connected to the outlet side of the low-pressure side refrigerant flow path 13b. Yes.

  On the other hand, the second evaporator 18 connected to the second outlet pipe portion 16c exchanges heat between the other refrigerant branched by the refrigerant distributor 16 and the blown air blown from the blower fan 17a, thereby reducing the pressure of the low-pressure refrigerant. It is a heat exchanger for heat absorption which cools blowing air by making it absorb heat. Further, the outlet side of the second evaporator 18 is connected to the refrigerant suction port 15 b of the ejector 15.

  In addition, the 1st evaporator 17 and the 2nd evaporator 18 of this embodiment are assembled | attached to the integral structure. Therefore, the blown air blown by the blower fan 17a flows in the direction of the arrow 200, first cooled by the first evaporator 17, and then cooled by the second evaporator 18 and flows into the space to be cooled (vehicle interior). . Therefore, in this embodiment, the same cooling object space can be cooled by the first evaporator 17 and the second evaporator 18.

  Next, an outline of the electric control unit of the present embodiment will be described. The air conditioning control device 20 includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof. The air conditioning control device 20 performs various calculations and processes based on the control program stored in the ROM, and controls the operations of the various actuators 11a, 12a, 14, 17a and the like.

  The air-conditioning control device 20 is configured integrally with the above-described control means for controlling the actuators. In the present embodiment, in particular, the operation of the electromagnetic capacity control valve 11a in the air-conditioning control device 20 is performed. The configuration (hardware and software) for controlling the discharge is the discharge capacity control means 20a, the structure for controlling the operation of the cooling fan 12a is the heat dissipation capacity control means 20b, and the structure for controlling the operation of the variable throttle mechanism 14 is the throttle mechanism control means 20c. To do. Of course, each control means 20a, 20b, 20c may be constituted by a separate control device.

  The air conditioning controller 20 also detects a first temperature sensor 21 that detects the temperature Th of the refrigerant on the downstream side of the high-pressure side refrigerant passage 13a of the internal heat exchanger 13, and detects the pressure Ph of the refrigerant on the downstream side of the high-pressure side refrigerant passage 13a. Detection signals such as a first pressure sensor 22 that detects the temperature Tn of the refrigerant on the inlet side of the nozzle 15a of the ejector 15 and a second pressure sensor 24 that detects the pressure Pn of the refrigerant on the inlet side of the nozzle 15a. Entered.

  In addition to the detection signals of the sensors, various operation signals from an operation panel (not shown) are input to the air conditioning control device 20. The operation panel is provided with an operation switch for operating the vehicle refrigeration apparatus, a temperature setting switch for setting the cooling temperature of the space to be cooled, and the like.

  Next, the operation of this embodiment having the above-described configuration will be described with reference to the Mollier diagram of FIG. Note that the solid line in FIG. 2 schematically shows the state of the refrigerant of the present embodiment, and the broken line shows a case where the refrigerant flowing into the nozzle portion is operated in a supercooled state in the cycle of the prior application example. ing. That is, the broken line in FIG. 2 is equivalent to the broken line in FIG. Further, in FIG. 2, the reference numerals indicating the state of the refrigerant are shown by changing the suffix of the reference numeral indicating the state of the corresponding refrigerant in FIG.

First, when the operation switch of the operation panel is turned on and the rotational driving force is transmitted to the compressor 11 from the vehicle running engine, the compressor 11 sucks the refrigerant, compresses it, and discharges it. The high-temperature and high-pressure gas-phase refrigerant discharged from the compressor (point A _{1 in} FIG. 2) flows into the radiator 12 and radiates heat by exchanging heat with the blown air (outside air) blown from the cooling fan (see FIG. 2). 2 A ₁ point → A ′ ₁ point).

The refrigerant that has flowed out of the radiator 12 flows into the high-pressure side refrigerant flow path 13a of the internal heat exchanger 13, exchanges heat with the refrigerant sucked by the compressor 11 that passes through the low-pressure side refrigerant flow path 13b, and is further cooled and excess. It will be in a cooling state (A ' ₁ point-> B ₁ point of FIG. 2).

In the present embodiment, for comparison with the cycle of the prior application example, it is assumed that the cooling is performed so that the degree of supercooling at the points B ₁ and _BL is equal. Refrigerant flowing out from the high-pressure side refrigerant flow path 13a is decompressed and expanded in intermediate pressure flows into the variable throttle mechanism 14 (B ₁ point of FIG. 2 → B _'1 point).

At this time, the throttle mechanism control means 20c of the air-conditioning control device 20 determines that the state of the refrigerant at B ′ ₁ is a supercooled state where the refrigerant is below a predetermined reference supercooling degree SSc (specifically, 10K) or a predetermined reference. A control signal is outputted so that the opening degree of the throttle passage of the variable throttle mechanism 14 is adjusted so that a gas-liquid two-phase state with a dryness of SGr (specifically 0.2) is obtained.

  More specifically, the throttle mechanism control means 20c is based on the detection value Th of the first temperature sensor 21 and the detection value Ph of the first pressure sensor 22, and is downstream of the high-pressure side refrigerant flow path 13a of the internal heat exchanger 13. The degree of supercooling of the refrigerant is calculated.

  Then, the degree of dryness of the refrigerant on the inlet side of the nozzle portion 15a is calculated from the degree of supercooling and the pressure difference between the detected value Ph of the first pressure sensor 22 and the detected value Pn of the second pressure sensor 24. Further, when the calculated dryness is greater than the reference dryness SGr (0.2), a control signal is output to the variable aperture mechanism 14 so that the calculated dryness is equal to or less than the reference dryness SGr.

  On the other hand, when the calculated dryness is equal to or less than the reference dryness SGr, the degree of supercooling of the refrigerant on the inlet side of the nozzle portion 15a is calculated based on the detection value Tn of the second temperature sensor 23 and the detection value Pn of the second pressure sensor. When the calculated value is larger than the reference supercooling degree SSc (10K), a control signal is output to the variable throttle mechanism 14 so as to be equal to or lower than the reference supercooling degree SSc.

  Therefore, in the present embodiment, the refrigerant state detection in which the first and second temperature sensors 21 and 23 and the first and second pressure sensors 22 and 24 detect the degree of supercooling or dryness of the refrigerant flowing into the nozzle portion 15a. Means are configured.

The refrigerant that has been decompressed by the variable throttle mechanism 14 until it reaches a supercooled state below the reference supercooling degree SSc or a gas-liquid two-phase state below the reference dryness SGr flows into the nozzle portion 15a of the ejector 15. The refrigerant flowing into the nozzle portion 15a of the ejector 15 expands under reduced pressure in an isentropic manner (point B ′ ₁ → point C _{1 in} FIG. 2).

  And the pressure energy of a refrigerant | coolant is converted into speed energy at the time of this decompression | expansion expansion, and a refrigerant | coolant is injected at high speed from the refrigerant | coolant injection port 15e of the nozzle part 15a. Due to the refrigerant suction action of the injected refrigerant, the refrigerant that has passed through the second evaporator 18 is sucked from the refrigerant suction port 15b.

Furthermore, the suction refrigerant sucked from injected injected refrigerant and the refrigerant suction port 15b of the nozzle section 15a are mixed in the mixing section 15c of the ejector 15 (C ₁ point in FIG. 2 → D ₁ point), the diffuser portion 15d Inflow. The expansion of the passage area in the diffuser portion 15d, since the velocity energy of refrigerant is converted into pressure energy, the pressure of the refrigerant is increased (D ₁ point of FIG. 2 → E ₁ point).

The flow of the refrigerant flowing out from the diffuser part 15d is branched at the branch part 100 inside the refrigerant distributor 16, and one of the branched refrigerants flows into the first evaporator 17 from the first outlet pipe part 16b, while it evaporated by absorbing heat from the blown air of the blower fan 17a, gradually reducing the pressure by the pressure loss in the first evaporator 17 (E ₁ point of FIG. 2 → F _'1 point).

Furthermore, the refrigerant that has flowed out of the first evaporator 17 flows into the low-pressure side refrigerant flow path 13b of the internal heat exchanger 13, and is heated by exchanging heat with the high-pressure refrigerant that passes through the high-pressure side refrigerant flow path 13a ( F ′ ₁ point → F ₁ point in FIG. 2). Refrigerant heated in the low-pressure side refrigerant flow path 13b of the internal heat exchanger 13 is again compressed is sucked into the compressor 11 (F ₁ point of FIG. 2 → A ₁ point).

On the other hand, the other refrigerant branched by the branch part 100 flows into the second evaporator 118 from the second outlet pipe part 16c, absorbs heat from the blown air of the blower fan 17a and evaporates, and then enters the second evaporator 18. The pressure is gradually reduced by the pressure loss and the suction action of the ejector 15 (E ₁ point → G ₁ point in FIG. 2). The refrigerant that has flowed out of the second evaporator 18 is sucked into the ejector 15 from the refrigerant suction port 15b (point G ₁ → point D _{1 in} FIG. 2).

  As described above, in the present embodiment, the refrigerant flow is branched by the refrigerant distributor 16 disposed on the outlet side of the diffuser portion 15d, and the refrigerant flows into the first evaporator 17 and the second evaporator 18. Therefore, the first evaporator 17 and the second evaporator 18 can exhibit a cooling action at the same time.

  At this time, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the second evaporator 18 can be reduced with respect to the refrigerant evaporation pressure (refrigerant evaporation temperature) of the first evaporator 17 by the suction action of the ejector 15. A temperature difference between the refrigerant evaporation temperature of the first evaporator 17 and the second evaporator 18 and the indoor blowing air of the blower fan 17a can be secured, and the indoor blowing air can be efficiently cooled.

  In the present embodiment, the flow of the refrigerant is branched in the branch portion 100 so that the dynamic pressure of the refrigerant flowing out from the diffuser portion 15d of the ejector 15 is maintained, so that the diffuser is placed inside the second evaporator 18. The dynamic pressure of the refrigerant flowing out from the portion 15d can be applied.

  Thus, when the refrigerant flows into the second evaporator 18, not only the pressure difference between the static pressure of the refrigerant on the downstream side of the diffuser portion 15d and the static pressure of the refrigerant at the refrigerant suction port 15b, but also the refrigerant that has flowed out of the diffuser portion 15d. Therefore, the refrigerant can surely flow into the second evaporator 18.

  Further, since the compressor 11 suction side is connected to the downstream side of the first evaporator 17, the refrigerant can surely flow into the first evaporator 17 by the suction action of the compressor 11. Therefore, the refrigerating capacity can be surely exhibited in both the evaporators 17 and 18.

  Further, in the present embodiment, the throttle mechanism control means 20c controls the operation of the variable throttle mechanism 14, and the state of the refrigerant flowing into the nozzle portion 15a is the supercooled state or the standard with the reference supercooling degree SSc (10 ° C.) or less. The refrigerant is expanded under reduced pressure until a gas-liquid two-phase state with a dryness of SGr (0.2) or less is obtained.

Accordingly, as described with reference to FIG. 12 described above, the refrigerant in a state where the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency, and can flow into the nozzle portion 15a. . Therefore, as shown in FIG. 2, the recovered energy amount ΔI ₁ of the ejector 15 of the present embodiment is larger than the recovered energy amount ΔI _L of the prior application example. As a result, more than the boost amount [Delta] P _L of the boost amount [Delta] P ₁ even prior application examples cycles of the diffuser portion 15d.

  That is, in order to increase the total value of the refrigeration capacities of the first and second evaporators 17 and 18, the internal heat exchanger 13 cools the refrigerant on the upstream side of the variable throttle mechanism 14 and reduces the enthalpy until the supercooling state is reached. Even if it decreases, the nozzle efficiency does not decrease.

Furthermore, as shown in FIG. 2, the total value ΔH ₁ 1 + ΔH ₁ 2 of the refrigerating capacity in the first and second evaporators 17 and 18 in the present embodiment is the refrigeration of the prior application example due to the increase in the amount of recovered energy described above. It is equal to or greater than the total capacity value ΔH _L 1 + ΔH _L 2.

  As a result, in the present embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased, resulting in a high cycle as a whole cycle. Efficiency can be demonstrated.

  In this embodiment, the reference supercooling degree SSc is set to 10K and the reference dryness degree SGr is set to 0.2. However, each of the sensors 21 to 24 constituting the refrigerant state detecting means has a sufficient detection capability. In this case, the reference supercooling degree SSc may be set to 5K, and the reference dryness degree SGr may be set to 0.1. Thereby, cycle efficiency can be further improved.

(Second Embodiment)
In the first embodiment, an example in which the electric variable aperture mechanism 14 is employed has been described. In the present embodiment, an example in which a mechanical variable aperture mechanism 30 is employed as shown in the overall configuration diagram of FIG. explain. In FIG. 3, the same or equivalent parts as those in the first embodiment are denoted by the same reference numerals. The same applies to the following embodiments.

  The variable throttle mechanism 30 is a pressure reducing means for reducing the high-pressure liquid-phase refrigerant to an intermediate pressure, and is also a flow rate adjusting means for adjusting the flow rate of the refrigerant flowing out to the nozzle portion 15a side of the ejector 15. Furthermore, in the present embodiment, specifically, a temperature type expansion valve whose throttle opening is changed by a mechanical mechanism is employed as the variable throttle mechanism 30. Therefore, the aperture mechanism control means 20c is also abolished.

  The variable throttle mechanism 30 has a temperature sensing unit 30a disposed on the downstream side of the first evaporator 17, and the downstream side of the first evaporator 17 based on the temperature and pressure of the refrigerant on the downstream side of the first evaporator 17. The degree of superheat of the refrigerant is detected, and the valve opening degree (refrigerant flow rate) is adjusted so that the degree of superheat of the refrigerant on the downstream side of the first evaporator 17 becomes a predetermined value. Therefore, for example, when the degree of superheat of the refrigerant on the downstream side of the first evaporator 17 increases, the valve opening degree is increased, and when the degree of superheat decreases, the valve opening degree is reduced.

  Furthermore, the variable throttle mechanism 30 of the present embodiment is configured such that when the valve opening reaches the maximum opening, the degree of supercooling of the refrigerant flowing into the nozzle portion 15a is equal to or less than the reference supercooling degree SSc (10 ° C.). The maximum opening is adjusted. Further, when the valve opening becomes the minimum opening, the minimum opening is performed so that the dryness of the refrigerant flowing into the nozzle portion 15a is equal to or less than the reference dryness SGr (0.2) without being completely closed. The degree has been adjusted.

  That is, in the variable throttle mechanism 30 of the present embodiment, the state of the refrigerant flowing into the nozzle portion 15a is at least a supercooled state having a reference supercooling degree SSc (10 ° C.) or less or a reference dryness degree SGr (0.2) or less. The refrigerant is expanded under reduced pressure in a range where a gas-liquid two-phase state is achieved.

  Furthermore, in this embodiment, the variable throttling mechanism 30 adjusts the degree of superheat of the refrigerant on the downstream side of the first evaporator 17 to a predetermined value. Therefore, the accumulator 19 is abolished and the receiver 12 is disposed downstream of the radiator 12. 31 is arranged.

  The receiver 31 is a gas-liquid separator that separates the gas-liquid of the high-pressure side refrigerant and stores excess refrigerant. The high-pressure side refrigerant flow path 13 a of the internal heat exchanger 13 is connected to the liquid-phase refrigerant outlet of the receiver 31. Other configurations are the same as those of the first embodiment.

  Therefore, even if this embodiment is operated, it basically operates as shown in FIG. 2 similarly to the first embodiment, and the first evaporator 17 and the second evaporator 18 can simultaneously and reliably exert a cooling action. . Further, since the variable throttle mechanism 30 decompresses and expands the refrigerant as described above, the refrigerant in a state in which the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency to the nozzle portion 15a. Can flow in.

  As a result, as in the first embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

(Third embodiment)
In the first and second embodiments, the example in which the variable aperture mechanisms 14 and 30 are employed has been described. However, in this embodiment, the fixed aperture mechanism 32 is employed as shown in the overall configuration diagram of FIG. As the fixed throttle mechanism 32, a capillary tube, an orifice, or the like can be specifically used. Therefore, the aperture mechanism control means 20c is also abolished. Other configurations are the same as those of the first embodiment.

  Next, the operation of this embodiment will be described. The ejector refrigeration cycle 10 of the present embodiment basically includes the first embodiment by adjusting the refrigerant discharge capacity of the compressor 11 by the discharge capacity control means 20a controlling the operation of the electromagnetic capacity control valve 11a. Operates in the same way.

  Specifically, the discharge capacity control means 20a detects the first and second temperature sensors 21 and 23 and the first and second pressure sensors 22 and 24 constituting the refrigerant state detection means, as in the first embodiment. Based on the value, the degree of supercooling and the degree of dryness of the refrigerant on the inlet side of the nozzle portion 15a are calculated, and the state of the refrigerant on the side of the nozzle portion 15a is the supercooled state or the standard dryness of the reference supercooling degree SSc (10 ° C.) The refrigerant discharge capacity is adjusted so that a gas-liquid two-phase state of SGr (0.2) or less is obtained.

  Therefore, even if the ejector refrigeration cycle of the present embodiment is operated, the first evaporator 17 and the second evaporator 18 can exhibit the cooling action simultaneously and reliably as in the first embodiment. Further, the discharge capacity control means 20a controls the operation of the electromagnetic capacity control valve 11a, and the refrigerant in a state in which the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency. It can be made to flow into part 15a.

  As a result, as in the first embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

(Fourth embodiment)
In the third embodiment, the example in which the discharge capacity control unit 20a controls the operation of the electromagnetic capacity control valve 11a has been described. However, in the present embodiment, the heat dissipation capacity control unit 20b controls the operation of the cooling fan 12a. explain. Therefore, the overall configuration of the present embodiment is the same as that of the third embodiment (FIG. 4).

  The ejector refrigeration cycle 10 of this embodiment operates basically in the same manner as in the first embodiment by adjusting the heat dissipation capacity of the radiator 12 by controlling the operation of the cooling fan 12a by the heat dissipation capacity control means 20b. To do.

  Specifically, the heat radiation capacity control means 20b is detected by the first and second temperature sensors 21 and 23 and the first and second pressure sensors 22 and 24 constituting the refrigerant state detection means, as in the first embodiment. Based on the value, the degree of supercooling and the degree of dryness of the refrigerant on the inlet side of the nozzle portion 15a are calculated, and the state of the refrigerant on the side of the nozzle portion 15a is the supercooled state or the standard dryness of the reference supercooling degree SSc (10 ° C.) The heat dissipation capability is adjusted so that a gas-liquid two-phase state of SGr (0.2) or less is obtained.

  Therefore, even if the ejector refrigeration cycle of the present embodiment is operated, the first evaporator 17 and the second evaporator 18 can exhibit the cooling action simultaneously and reliably as in the first embodiment. Further, the heat radiation capacity control means 20b controls the operation of the cooling fan 12a, and the refrigerant in a state in which the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency to the nozzle portion 15a. Can flow in.

  As a result, as in the first embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

  In the present embodiment, the state of the refrigerant flowing into the nozzle portion 15a is set to the above-described state only by adjusting the heat dissipation capability of the radiator 12, but the compressor 11 described in the third embodiment The refrigerant discharge capacity may be adjusted simultaneously.

(Fifth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 5, a throttle mechanism integrated type throttle internal heat exchanger 33 is adopted for the internal heat exchanger 13 of the first embodiment. Other configurations are the same as those of the first embodiment.

  As in the first embodiment, the internal heat exchanger 33 exchanges heat between the refrigerant on the downstream side of the radiator 12 that passes through the high-pressure side refrigerant flow path and the refrigerant on the suction side of the compressor 11 that passes through the low-pressure side refrigerant flow path. Thus, the radiator 12 downstream side refrigerant is radiated. Therefore, the enthalpy difference (refrigeration capacity) of the refrigerant between the refrigerant inlet and outlet in the first and second evaporators 17 and 18 can be increased.

  Furthermore, the high-pressure side refrigerant flow path of the internal heat exchanger 33 is formed by a capillary tube 33a as decompression means, and the low-pressure side refrigerant flow path is defined by a low-pressure side refrigerant pipe 33b through which the compressor 11 suction-side refrigerant passes. Is formed. Accordingly, the internal heat exchanger 33 constitutes a refrigerant heat radiating means for radiating heat from the radiator 12 outlet side refrigerant in the decompression and expansion process passing through the capillary tube 33a.

  As a specific configuration of the internal heat exchanger 33, a configuration in which the capillary tube 33a and the low-pressure side refrigerant pipe 33b are joined by a joining means such as welding, pressure welding, or soldering to exchange heat, or a low-pressure side refrigerant is used. A double-pipe heat exchanger configuration in which the capillary tube 33a is disposed inside the pipe 33b can be employed.

  Here, since the capillary tube 33a depressurizes the refrigerant by the effect of reducing the area of the refrigerant passage and the frictional force in the refrigerant passage, the capillary tube 33a has an elongated shape having a predetermined refrigerant passage length. Therefore, by adopting the capillary tube 33a as the decompression means, it becomes easy to secure a heat exchange area between the capillary tube 33a and the low-pressure side refrigerant pipe 33b, and the refrigerant passing through the capillary tube 33a can be efficiently radiated.

  Next, the operation of the present embodiment having the above-described configuration will be described with reference to the Mollier diagram of FIG. The solid line in FIG. 6 schematically shows the state of the refrigerant of the present embodiment, and the broken line shows a case where the refrigerant flowing into the nozzle portion is operated in a supercooled state in the cycle of the prior application example. ing. That is, the broken line in FIG. 6 is equivalent to the broken line in FIG. Further, in FIG. 6, the reference numerals indicating the state of the refrigerant are shown by changing the suffix of the reference numeral indicating the state of the corresponding refrigerant in FIG. 11 to 5.

Similarly to the first embodiment, the driving force is transmitted to the compressor 11, and the high-temperature and high-pressure gas-phase refrigerant (point A _{5 in} FIG. 6) discharged from the compressor flows into the radiator 12 and dissipates heat. (a _5-point of Figure 6 → B ₅ points). At this time, the state of the radiator 12 outlet side refrigerant (B ₅ points in FIG. 6) is controlled so as to approach the saturated liquid phase state.

  Specifically, for example, the heat radiation capacity control means 20b adjusts the amount of blown air by changing the rotational speed of the cooling fan 12a in accordance with the change in the refrigerant load of the cycle. The rotational speed control of the cooling fan 12a is not a control for strictly setting the state of the refrigerant at the outlet of the radiator 12 to a saturated liquid phase state. Therefore, depending on the operation state of the cycle, the state of the refrigerant at the outlet of the radiator 12 may change from the saturated liquid phase state to the gas-liquid two-phase side or the supercooling side state.

The refrigerant in a state close to the saturated liquid phase on the outlet side of the radiator 12 flows into the capillary tube 33a of the internal heat exchanger 33 and is reduced to an intermediate pressure, and passes through the low-pressure side refrigerant pipe 33b in the decompression and expansion process. compressor 11 suction refrigerant and by heat exchange with, further cooled to lower the enthalpy (B ₅ points in FIG. 6 → B _'5 points).

Furthermore, in the present embodiment, when the capillary tube 33a the outlet side refrigerant B _'5 points state reference subcooling degree SSc (specifically the state of the capillary tube 33a upstream refrigerant B ₅ points is in a saturated liquid phase state Specifically, the heat exchange amount of the internal heat exchanger 33 is adjusted so as to be in a supercooled state of 10K) or less or a gas-liquid two-phase state of a predetermined reference dryness SGr (specifically 0.2) or less. Has been.

The refrigerant flowing out from the capillary tube 33a is isentropically decompressed and expanded at the nozzle portion 15a in the same manner as in the first embodiment, passes through the mixing portion 15c → the diffuser portion 15d in this order, and is branched at the branch portion 100. (B ′ ₅ point → C ₅ point → D ₅ point → E ₅ point in FIG. 6).

Further, one refrigerant branched in the branching section 100 passes through the low pressure side refrigerant pipe 33b of the first evaporator and the internal heat exchanger 33 and is sucked into the compressor 11 and compressed (E in FIG. 6). ₅ points → F ' ₅ points → F ₅ points → A ₅ points). On the other hand, the other refrigerant branched at the branch portion 100 is sucked into the refrigerant suction port 15b of the ejector 15 passes through the second evaporator 18 (E ₅ points in FIG. 6 → G ₅ points → D ₅ points) .

  In this embodiment, since it operates as described above, similarly to the first embodiment, the first evaporator 17 and the second evaporator 18 can simultaneously and reliably exhibit a cooling action. Furthermore, the refrigerant in a state where the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency, can be caused to flow into the nozzle portion 15a.

Accordingly, as shown in FIG. 6, the recovered energy amount ΔI ₅ of the ejector 15 of the present embodiment is larger than the recovered energy amount ΔI _L of the prior application example. As a result, the boosted amount of the diffuser portion 15d [Delta] P ₅ becomes larger than the boost amount [Delta] P _L of the prior application example of the cycle.

Furthermore, as shown in FIG. 6, the total value ΔH ₅ 1 + ΔH ₅ 2 of the refrigerating capacity in the first and second evaporators 17 and 18 in the present embodiment is the refrigeration of the prior application example due to the increase in the recovered energy amount. It is equal to or greater than the total capacity value ΔH _L 1 + ΔH _L 2.

  As a result, as in the first embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

  Further, in this embodiment, since the internal heat exchanger 33 integrated with the throttle mechanism is adopted, as shown in FIG. 6, the state of the refrigerant on the inlet side and the outlet side of the capillary tube 33a is changed to the saturated liquid phase state. Even in a close state, the state of the refrigerant passing through the capillary tube 33a can be changed to a gas-liquid two-phase state.

As a result, a significant density changes due to the phase change of the refrigerant in the decompression expansion process of the capillary tube 33a (B ₅ points in FIG. 6 → B _'5 points) can be suppressed, vacuum capability and flow control capabilities of the capillary tube 33a stability Can be made.

(Sixth embodiment)
In the fifth embodiment, the example in which the heat dissipation capacity control unit 20b controls the operation of the cooling fan 12a to adjust the state of the refrigerant on the outlet side of the radiator 12 so as to approach the saturated liquid phase state has been described. Now, an example in which the discharge capacity control means 20a controls the operation of the electromagnetic capacity control valve 11a will be described. Therefore, the overall configuration of this embodiment is the same as that of the fifth embodiment (FIG. 5).

  The ejector type refrigeration cycle 10 of the present embodiment basically has the same structure as that of the fifth embodiment, as the discharge capacity control means 20a controls the operation of the electromagnetic capacity control valve 11a to adjust the refrigerant discharge capacity of the compressor 11. Operates similarly.

  Specifically, the discharge capacity control means 20a detects the first and second temperature sensors 21 and 23 and the first and second pressure sensors 22 and 24 constituting the refrigerant state detection means, as in the first embodiment. Based on the value, the degree of supercooling and the degree of dryness of the refrigerant on the inlet side of the nozzle portion 15a are calculated, and the state of the refrigerant on the side of the nozzle portion 15a is the supercooled state or the standard dryness of the reference supercooling degree SSc (10 ° C.) or less. The heat dissipation capability is adjusted so that a gas-liquid two-phase state of SGr (0.2) or less is obtained.

  Therefore, even if the ejector refrigeration cycle of the present embodiment is operated, the first evaporator 17 and the second evaporator 18 can exhibit the cooling action simultaneously and reliably as in the fifth embodiment. Further, the discharge capacity control means 20a controls the operation of the electromagnetic capacity control valve 11a, and the refrigerant in a state in which the total value of the refrigerating capacity of each evaporator can be increased without causing a decrease in nozzle efficiency. It can be made to flow into part 15a.

  As a result, as in the fifth embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

  Furthermore, in this embodiment, in order to further expand the total value of the refrigeration capacities of the first and second evaporators 17 and 18, the enthalpy is reduced until the state of the refrigerant at the outlet of the radiator 12 becomes a supercooled state. However, since the state of the refrigerant flowing into the nozzle portion 15a is the above-described state, the nozzle efficiency is not reduced.

(Seventh embodiment)
In the first embodiment, the example in which the variable throttle mechanism 14 is arranged on the upstream side of the nozzle portion 15a of the ejector 15 has been described. However, in the present embodiment, as shown in the overall configuration diagram of FIG. Thus, the variable aperture mechanism 14 and the aperture mechanism control means 20c are eliminated. Details of the variable ejector 35 will be described with reference to an enlarged sectional view of FIG. FIG. 8 is a sectional view in the axial direction of the variable ejector 35 and is an enlarged view around the nozzle portion 35a.

  The variable ejector 35 includes a nozzle portion 35 a that decompresses and expands the high-pressure refrigerant that has flowed out of the internal heat exchanger 13. The nozzle portion 35a has a substantially cylindrical shape with a sharp coolant injection port 35e side, and a refrigerant passage 35f through which high-pressure refrigerant passes is formed coaxially with the nozzle portion 35a. Further, the refrigerant passage 35f is formed with first and second restricting portions 35g and 35h by restricting the passage cross-sectional area as viewed from the axial direction.

  More specifically, in the refrigerant passage 35f, the passage sectional area from the refrigerant inlet 35i side of the nozzle portion 15a to the first throttle portion 35g gradually decreases, and the passage from the first throttle portion 35g to the second throttle portion 35h. After the cross-sectional area gradually increases, it gradually decreases again, and the cross-sectional area from the second throttle portion 35h to the refrigerant injection port 35e gradually increases.

That is, the nozzle part 35a is a so-called two-stage expansion nozzle, and the first throttle part 35g is arranged on the upstream side in the refrigerant flow direction with respect to the second throttle part 35h. Furthermore, between the first throttle portion 35g and the second throttle portion 35 h, the passage sectional area as viewed from the axial direction first, second throttle portion 35g, pressure recovery space 35j is formed larger than 35h .

  Further, the variable ejector 35 is disposed inside the refrigerant passage 35f, and adjusts the throttle opening of the first and second throttle portions 35g and 35h and the opening of the refrigerant injection port 35e, and the needle valve 35k. Has an electric actuator 35l for displacing the nozzle in the axial direction of the nozzle portion 35a. Therefore, in the present embodiment, the throttle valve opening adjusting means is configured by the needle valve 35k and the electric actuator 35l.

  As the electric actuator 35l, for example, a motor actuator such as a stepping motor or an electromagnetic solenoid mechanism can be employed. Further, the electric actuator 35l is driven and controlled by a control signal output from the aperture opening degree control means 20d of the air conditioning control device 20.

  Further, as shown in FIG. 7, the variable ejector 35, like the ejector 15 of the first embodiment, has a refrigerant suction port 35b that sucks the refrigerant that has flowed out of the second evaporator 18, and a high speed that is ejected from the nozzle portion 35a. The mixing portion 35c for mixing the refrigerant flow and the suctioned refrigerant sucked from the refrigerant suction port 35b, and the diffuser portion 35d forming the pressure increasing portion are configured.

  In the present embodiment, since the pressure reducing means is not arranged on the upstream side of the nozzle portion 35a, the two temperature sensor 23 and the second pressure sensor 24 are omitted. Other configurations are the same as those of the first embodiment.

  Next, the operation of this embodiment having the above-described configuration will be described with reference to the Mollier diagram of FIG. Note that the solid line in FIG. 9 schematically shows the state of the refrigerant of the present embodiment, and the broken line shows a case where the refrigerant flowing into the nozzle portion is operated in a supercooled state in the cycle of the prior application example. ing. That is, the broken line in FIG. 9 is equivalent to the broken line in FIG. In FIG. 9, the reference numerals indicating the state of the refrigerant are shown by changing the subscript of the reference numeral indicating the state of the corresponding refrigerant in FIG. 11 to 7.

Like the first embodiment, the driving force is transmitted to the compressor 11, gas refrigerant of high temperature and high pressure state discharged from the compressor (A _7-point in Fig. 9) radiates heat flows into the radiator 12 Te (a _7-point in Fig. 9 → B ₇ points), this time, as in the fifth embodiment, the state of the radiator 12 refrigerant on the outlet side (B ₇ points in FIG. 9) is to approach the saturated liquid phase state Be controlled.

The refrigerant in a state near the saturated liquid phase on the outlet side of the radiator 12 flows into the nozzle portion 35 a of the variable ejector 35. The refrigerant flowing into the nozzle portion 35a of the variable ejector 35 is first decompressed and expanded by the first throttle portion 35 g (B ₇ points in FIG. 9 → B _'7 points).

At this time, throttle opening control means 20d of the control unit 20, B 'refrigerant in the state of ₇ points, and outputs a control signal so that the gas-liquid two-phase state in which predetermined electric actuator by the control signal 35l displaces the needle valve 35k to adjust the throttle opening of the first throttle 35g.

  More specifically, the throttle opening degree control means 20d is downstream of the high-pressure side refrigerant flow path 13a of the internal heat exchanger 13 based on the detection value Th of the first temperature sensor 21 and the detection value Ph of the first pressure sensor 22. The degree of supercooling of the side refrigerant is calculated. Then, from this degree of supercooling, the target throttle opening degree is determined so that the refrigerant state in the pressure recovery space 35j on the downstream side of the first throttle part 35g becomes a predetermined gas-liquid two-phase state, and becomes the target throttle opening degree. A control signal is output to the electric actuator 35l.

Decompressed and expanded refrigerant in the first throttle portion 35g is to recover the pressure in the pressure recovery space 35j (B 'in FIG. 9 ₇ points → B'_'7 points), to the refrigerant injection port 35e from the second diaphragm portion 35h In the refrigerant passage 35f that reaches, it is again decompressed and expanded (B ″ ₇ point → C ₇ point in FIG. 9).

  Here, in the nozzle part 35a of the present embodiment, when the opening degree of the first throttle part 35g is adjusted, the throttle opening degree of the second throttle part 35h and the opening degree of the refrigerant injection port 35e are also adjusted. Accordingly, the size and shape of the refrigerant passage 35f from the second throttle portion 35h to the refrigerant injection port 35e are designed so that the refrigerant flowing into the second throttle portion 35h is decompressed and expanded in an isentropic manner.

The refrigerant injected from the nozzle portion 35a is branched mixing portion 35c → passes in order of the diffuser portion 35d at the branch portion 100 (C ₇ points in FIG. 9 → D ₇ points → E ₇ points).

Further, one of the refrigerants branched at the branching part 100 exhibits an endothermic effect in the first evaporator 17, passes through the low-pressure side refrigerant flow path 13 b of the internal heat exchanger 13, and is sucked into the compressor 11. (E ₇ point → F ′ ₇ point → F ₇ point → A ₇ point in FIG. 9). On the other hand, the other refrigerant branched at the branch portion 100, exerts an endothermic effect by the second evaporator 18 is sucked into the refrigerant suction port 35b of the variable ejector 35 (E ₇ points in FIG. 9 → G ₇ points → D ₇ points).

  In this embodiment, since it operates as described above, similarly to the first embodiment, the first evaporator 17 and the second evaporator 18 can simultaneously and reliably exhibit a cooling action.

  Furthermore, with the nozzle portion 35a of the variable ejector 35 as a two-stage expansion nozzle, the refrigerant expanded under reduced pressure until the gas-liquid two-phase state is obtained by the first throttle portion 35g, the refrigerant reaching the refrigerant injection port 35e from the second throttle portion 35h. Since the passage 35f is decompressed and expanded again, the nozzle efficiency on the downstream side of the second throttle portion 35h does not decrease. That is, the nozzle efficiency as the whole nozzle part 35a is not reduced.

Therefore, as shown in FIG. 9, the recovered energy amount ΔI ₇ of the ejector 15 of the present embodiment is larger than the recovered energy amount ΔI _L of the prior application example. As a result, the boosted amount of the diffuser portion 15d [Delta] P ₇ becomes larger than the boost amount [Delta] P _L of the prior application example of the cycle.

Furthermore, as shown in FIG. 9, the total value ΔH ₇ 1 + ΔH ₇ 2 of the refrigerating capacity in the first and second evaporators 17 and 18 in the present embodiment is the refrigeration of the prior application example due to the increase in the amount of recovered energy described above. It is equal to or greater than the total capacity value ΔH _L 1 + ΔH _L 2.

  As a result, as in the first embodiment, the driving power of the compressor 11 can be reduced, and the total value of the refrigeration capacities of the first and second evaporators 17 and 18 can be increased. Overall, high cycle efficiency can be exhibited.

(Other embodiments)
The present invention is not limited to the above-described embodiment, and can be variously modified as follows.

  (1) In the above-described embodiment, the ejector refrigeration cycle 10 of the present invention is applied to a vehicle air conditioner, but the application of the present invention is not limited to this. For example, the present invention may be applied to a household refrigerator, a vending machine cooling device, a showcase with a refrigeration function, and the like. Further, an HC refrigerant may be adopted as the refrigerant.

  Furthermore, when applied to a refrigeration / refrigeration apparatus or the like with little load fluctuation, adjustment of the throttle opening degree of the variable throttle mechanism 14 may be eliminated as in the above-described embodiment. In this case, for example, in the configuration of the third to sixth embodiments, the state of the refrigerant on the inlet side of the nozzle portion 15a during normal operation is a supercooled state with a reference supercooling degree SSc or less or a gas-liquid two-phase with a reference dryness SGr or less. The pressure reducing capacity of the fixed throttle mechanism 32 and the capillary tube 33a may be adjusted in advance so as to be in a state.

  (2) In the above-described embodiment, the heat radiation capacity adjusting means is configured by the cooling fan 12a, but the heat radiation capacity adjusting means is not limited to this. For example, as the heat dissipation capacity adjusting means, a heat dissipation capacity may be adjusted by providing a bypass passage that bypasses the refrigerant in the middle of the radiator 12 to the downstream side of the radiator 12. Furthermore, a heat dissipation capability may be adjusted by providing a blocking mechanism that blocks the flow of the blown air from the cooling fan 12a.

  (3) In the above-described embodiment, the first evaporator 17 and the second evaporator 18 are assembled in an integrated structure. As specific means, for example, the configuration of the first evaporator 17 and the second evaporator 18 is used. The parts may be made of aluminum and joined to the unitary structure by brazing.

  Furthermore, it may be configured to be integrally coupled with an interval of 10 mm or less by mechanical engagement means such as bolt tightening. Further, as the first evaporator 17 and the second evaporator 18, a fin-and-tube type heat exchanger is adopted, and the fins of the first evaporator 17 and the second evaporator 18 are made common and the tubes are in contact with the fins. You may integrate as a structure divided | segmented by.

  (4) In the above-described embodiment, the same cooling target space is cooled by the first evaporator 17 and the second evaporator 18, but different cooling target spaces are cooled by both the evaporators 17 and 18. Good. Alternatively, the first evaporator 17 may be eliminated and the second evaporator 18 alone may be used for cooling.

  (5) In the above-described embodiment, the refrigerant distributor 16 having a T-shaped three-way joint structure is adopted, but the refrigerant distributor is not limited to this. For example, a refrigerant distributor having a Y-shaped three-way joint structure may be employed.

  In this case, the introduction pipe part for introducing the refrigerant, the first and second outlet pipe parts for discharging the refrigerant, the refrigerant outlet direction in the first outlet pipe part and the refrigerant outlet direction in the second outlet pipe part, The dynamic pressure of the refrigerant flowing out of the ejector 15 can be maintained by directing the refrigerant in the introduction pipe portion toward the target direction and intersecting at an acute angle.

  (6) In the seventh embodiment described above, in the refrigerant passage 35f of the nozzle portion 35a, the passage sectional area from the second throttle portion 35h to the refrigerant injection port 35e is gradually increased. It is good also as a shape which does not change the passage cross-sectional area from the throttle part 35h to the refrigerant injection port 35e.

  Furthermore, in the nozzle part 35a of the variable ejector 35, when the opening degree of the first throttle part 35g is adjusted, the throttle opening degree of the second throttle part 35h and the opening degree of the refrigerant injection port 35e are also adjusted. As shown in FIG. 3, the axial length of the needle valve 35k ′ may be changed so that only the throttle opening of the second throttle part 35h is adjusted simultaneously with the throttle opening adjustment of the first throttle part 35g. .

  (7) In the above embodiment, the radiator 12 is an outdoor heat exchanger that exchanges heat between the refrigerant and the outside air, and the first evaporator 17 and the second evaporator 18 are indoor heat exchangers to cool the vehicle interior. Although the first evaporator 17 and the second evaporator 18 are configured as outdoor heat exchangers that absorb heat from a heat source such as outside air, the radiator 12 heats a fluid to be heated such as air or water. The present invention may be applied to a heat pump cycle configured as an indoor heat exchanger.

It is a whole block diagram of the ejector-type refrigerating cycle of 1st Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of 1st Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 2nd Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 3rd Embodiment and 4th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 5th Embodiment and 6th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of 5th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 7th Embodiment. It is an expanded sectional view around the nozzle part of the ejector of a 7th embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of 7th Embodiment. It is an expanded sectional view around the nozzle part of the ejector of other embodiments. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigeration cycle of the example of a prior application. It is a graph which shows the relationship between the state of a refrigerant | coolant, and cycle efficiency.

Explanation of symbols

11 ... compressor, 11a ... electromagnetic capacity control valve,
12 ... radiator, 12a ... cooling fan, 14, 30 ... variable throttle mechanism,
15 ... Ejector, 15a ... Nozzle part, 15b ... Refrigerant suction port,
17 ... 1st evaporator, 18 ... 2nd evaporator,
20 ... Air-conditioning control device, 20a ... Discharge capability control means, 20b ... Heat dissipation capability control means,
20c: throttle mechanism control means, 20d: throttle opening control means,
21 ... 1st temperature sensor, 22 ... 1st pressure sensor,
23 ... second temperature sensor, 24 ... second pressure sensor,
32 ... Fixed throttle mechanism, 33 ... Internal heat exchanger, 33a ... Capillary tube,
35 ... Variable ejector, 35a ... Nozzle part, 35b ... Refrigerant suction port,
35f ... refrigerant passage, 35g ... first throttle part, 35h ... second throttle part,
35k ... Needle valve, 35l ... Electric actuator, 100 ... Branching part.

Claims (18)

A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating heat from the high-pressure refrigerant compressed by the compressor (11);
Decompression means (14, 30, 32) for decompressing and expanding the downstream side refrigerant of the radiator (12);
An ejector (15) for sucking the refrigerant from the refrigerant suction port (15b) by a high-speed refrigerant flow ejected from the nozzle (15a) for further decompressing and expanding the refrigerant decompressed by the decompression means (14, 30, 32). When,
A branch part (100) for branching the flow of the refrigerant on the downstream side of the ejector (15);
A first evaporator (17) that evaporates one refrigerant branched at the branch (100) and flows out to the suction side of the compressor (11);
A second evaporator (18) that evaporates the other refrigerant branched at the branch portion (100) and flows out to the upstream side of the refrigerant suction port (15b),
The branch part (100) includes an introduction pipe part (16a) through which the refrigerant on the downstream side of the ejector (15) flows and a refrigerant in the introduction pipe part (16a) that flows out to the first evaporator (17) side. 1 lead-out pipe part (16b) and a second lead-out pipe part (16c) for allowing the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch part (100), the introduction pipe part (16a) and the second lead-out pipe part (16c) include a refrigerant inflow direction of the lead-in pipe part (16a) and a refrigerant in the second lead-out pipe part (16c). Arranged so that the outflow direction is in the same direction,
In the branch unit (100), wherein through said ejector (15) the second outlet pipe part from the inlet pipe portion (16a) as the dynamic pressure of the outflow refrigerant is maintained from (16c) Branching the refrigerant flow to the second evaporator (18) side ,
The depressurization means (14, 30, 32) includes a supercooled state in which the state of the refrigerant flowing into the nozzle portion (15a) is equal to or lower than a predetermined reference supercooling degree (SSc) or a predetermined reference dryness (SGr). An ejector-type refrigeration cycle, wherein the refrigerant is decompressed and expanded so as to be in the following gas-liquid two-phase state. The ejector-type refrigeration cycle according to claim 1, wherein the decompression means comprises an electric variable throttle mechanism (14). Furthermore, refrigerant state detection means (21-24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle part (15a),
Diaphragm mechanism control means (20c) for controlling the operation of the variable diaphragm mechanism (14),
The throttle mechanism control means (20c) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state below the reference supercooling degree (SSc) or below the reference dryness degree (SGr). The ejector refrigeration cycle according to claim 2, wherein the operation of the variable throttle mechanism (14) is controlled so as to be in a state. The ejector-type refrigeration cycle according to claim 1, wherein the decompression means comprises a variable throttle mechanism (30) whose throttle opening is changed by a mechanical mechanism. The ejector refrigeration cycle according to claim 4, wherein the variable throttle mechanism (30) is a temperature type expansion valve. The ejector-type refrigeration cycle according to claim 1, wherein the decompression means comprises a fixed throttle mechanism (32). Furthermore, refrigerant state detection means (21-24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle part (15a),
Discharge capacity adjusting means (11a) for adjusting the refrigerant discharge capacity of the compressor (11);
A discharge capacity control means (20a) for controlling the operation of the discharge capacity adjusting means (11a),
The discharge capacity control means (20a) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state below the reference supercooling degree (SSc) or below the reference dryness degree (SGr). The ejector refrigeration cycle according to claim 6, wherein the operation of the discharge capacity adjusting means (11a) is controlled so as to be in a state. Furthermore, refrigerant state detection means (21-24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle part (15a),
A heat dissipation capacity adjusting means (12a) for adjusting the heat dissipation capacity of the radiator (12);
A heat radiation capacity control means (20b) for controlling the operation of the heat radiation capacity adjustment means (12a),
The heat dissipation capacity control means (20b) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state below the reference supercooling degree (SSc) or below the reference dryness degree (SGr). The ejector refrigeration cycle according to claim 6 or 7, wherein the operation of the heat radiation capacity adjusting means (12a) is controlled so as to be in a state. A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating heat from the high-pressure refrigerant compressed by the compressor (11);
Decompression means (33a) for decompressing and expanding the downstream side refrigerant of the radiator (12);
Refrigerant heat radiating means (33) for radiating the refrigerant in the decompression and expansion process in the pressure reducing means (33a);
An ejector (15) for sucking the refrigerant from the refrigerant suction port (15b) by a high-speed refrigerant flow ejected from the nozzle (15a) for further decompressing and expanding the refrigerant decompressed by the decompression means (33a);
A branch part (100) for branching the flow of the refrigerant on the downstream side of the ejector (15);
A first evaporator (17) that evaporates one refrigerant branched at the branch (100) and flows out to the suction side of the compressor (11);
A second evaporator (18) that evaporates the other refrigerant branched at the branch portion (100) and flows out to the upstream side of the refrigerant suction port (15b),
The branch part (100) includes an introduction pipe part (16a) through which the refrigerant on the downstream side of the ejector (15) flows and a refrigerant in the introduction pipe part (16a) that flows out to the first evaporator (17) side. 1 lead-out pipe part (16b) and a second lead-out pipe part (16c) for allowing the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch part (100), the introduction pipe part (16a) and the second lead-out pipe part (16c) include a refrigerant inflow direction of the lead-in pipe part (16a) and a refrigerant in the second lead-out pipe part (16c). Arranged so that the outflow direction is in the same direction,
In the branch unit (100), wherein through said ejector (15) the second outlet pipe part from the inlet pipe portion (16a) as the dynamic pressure of the outflow refrigerant is maintained from (16c) Branching the refrigerant flow to the second evaporator (18) side ,
The depressurization means (33a) is a gas-liquid in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state equal to or lower than a predetermined reference supercooling degree (SSc) or a predetermined reference dryness (SGr). An ejector-type refrigeration cycle, wherein the refrigerant is expanded under reduced pressure so as to be in a two-phase state. The refrigerant heat dissipating means includes an internal heat exchanger (33) for exchanging heat between the refrigerant (12) downstream refrigerant and the compressor (11) suction refrigerant,
The said decompression means (33a) forms the high pressure side refrigerant flow path which allows the said heat sink (12) downstream refrigerant | coolant to pass through among the said internal heat exchangers (33). Ejector type refrigeration cycle. The ejector refrigeration cycle according to claim 9 or 10, wherein the decompression means is a capillary tube (33a). Furthermore, refrigerant state detection means (21-24) for detecting the degree of supercooling or dryness of the refrigerant flowing into the nozzle part (15a),
A discharge capacity control means (20a) for controlling the operation of the compressor (11),
The discharge capacity control means (20a) is a gas-liquid two-phase in which the state of the refrigerant flowing into the nozzle portion (15a) is a supercooled state below the reference supercooling degree (SSc) or below the reference dryness degree (SGr). The ejector refrigeration cycle according to any one of claims 9 to 11, wherein the operation of the compressor (11) is controlled so as to be in a state. The reference supercooling degree (SSc) is 10K,
The ejector refrigeration cycle according to any one of claims 1 to 12, wherein the reference dryness (SGr) is 0.2. The reference supercooling degree (SSc) is 5K,
The ejector refrigeration cycle according to any one of claims 1 to 13, wherein the reference dryness (SGr) is 0.1. A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating heat from the high-pressure refrigerant compressed by the compressor (11);
An ejector (35) for sucking the refrigerant from the refrigerant suction port (35b) by a high-speed refrigerant flow ejected from the nozzle (35a) for decompressing and expanding the refrigerant on the downstream side of the radiator (12);
A branch part (100) for branching the flow of the refrigerant on the downstream side of the ejector (35);
A first evaporator (17) that evaporates one refrigerant branched at the branch (100) and flows out to the suction side of the compressor (11);
A second evaporator (18) that evaporates the other refrigerant branched at the branch portion (100) and flows out to the upstream side of the refrigerant suction port (35b);
The branch part (100) includes an introduction pipe part (16a) through which the refrigerant on the downstream side of the ejector (15) flows and a refrigerant in the introduction pipe part (16a) that flows out to the first evaporator (17) side. 1 lead-out pipe part (16b) and a second lead-out pipe part (16c) for allowing the refrigerant in the introduction pipe part (16a) to flow out to the second evaporator (18) side,
In the branch part (100), the introduction pipe part (16a) and the second lead-out pipe part (16c) include a refrigerant inflow direction of the lead-in pipe part (16a) and a refrigerant in the second lead-out pipe part (16c). Arranged so that the outflow direction is in the same direction,
In the branch unit (100), wherein through said ejector (15) the second outlet pipe part from the inlet pipe portion (16a) as the dynamic pressure of the outflow refrigerant is maintained from (16c) Branching the refrigerant flow to the second evaporator (18) side ,
The refrigerant passage (35f) in the nozzle portion (35a) is formed with first and second throttle portions (35g, 35h) configured by narrowing the passage cross-sectional area,
The first throttle part (35g) is arranged upstream of the second throttle part (35h) in the refrigerant flow direction, and depressurizes the refrigerant to be in a gas-liquid two-phase state. Ejector refrigeration cycle. In the refrigerant passage (35f), the cross-sectional area of the passage from the refrigerant inlet side of the nozzle portion (35a) to the first throttle portion (35g) gradually decreases, and the second throttle portion (35g) to the second portion. The ejector-type refrigeration cycle according to claim 15, wherein the ejector refrigeration cycle is formed in a shape that gradually decreases after the passage sectional area to the throttle portion (35h) gradually increases. The said ejector is a variable ejector (35) which has a throttle opening adjustment means (35k, 35l) which adjusts the throttle opening of the said 1st aperture | diaphragm | squeeze part (35g), It is characterized by the above-mentioned. Ejector type refrigeration cycle. Furthermore, it comprises a throttle opening control means (20d) for controlling the operation of the throttle opening adjustment means (35k, 35l),
The throttle opening control means (20d) controls the operation of the throttle opening adjustment means (35k, 35l) so that the refrigerant flowing into the second throttle portion (35h) is in a gas-liquid two-phase state. The ejector-type refrigeration cycle according to claim 17.

Images