JP2008256240A - Ejector type refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To exercise high cycle efficiency COP by making a flow rate Vdi of an outflow refrigerant of a diffuser portion of an ejector type refrigerating cycle approach an optimum flow rate Vdif-max. <P>SOLUTION: An air-conditioning control device 20 adjusts a refrigerant discharge capacity of a compressor 11 to keep the flow rate Vdif of the outflow refrigerant of the diffuser portion 15d calculated on the basis of a total refrigerating capacity Qre determined from enthalpy H1out of outlet refrigerant of a first evaporator 17, enthalpy H2out of output refrigerant of a second evaporator 18 and the like, and a dryness Xnoz of the mixed refrigerant determined from dryness Znoz of injected refrigerant of a nozzle portion 15a of an ejector 15, enthalpy of outflow refrigerant of the diffuser portion 15d and the like, within a predetermined range. Thus, a pressure loss ΔP1drop in the first evaporator 17 is kept to be pressure rise ΔPej of the diffuser portion 15d or less, to make the flow rate Vdif of the outflow refrigerant of the diffuser portion approach the maximum flow rate Vdif-max. <P>COPYRIGHT: (C)2009,JPO&INPIT

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.
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, the refrigerant branched at the branching portion is directly flowed into the second evaporator without going through the decompression device or the like, so that the static pressure of the refrigerant at the outlet of the expansion device and 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 from the static pressure of the refrigerant but also by the dynamic pressure of the refrigerant that has flowed out of the ejector. As a result, the refrigerating capacity can be surely exhibited in the second evaporator.

  Furthermore, since the refrigerant whose pressure has been increased by the pressure increasing action of the diffuser portion of the ejector is sucked into the compressor through the first evaporator, the driving power of the compressor can be reduced. That is, in the cycle of the prior application example, it is possible to achieve both improvement of cycle efficiency (COP) and reliable display of the refrigerating capacity in the second evaporator by reducing the compressor driving power.

  By the way, in the cycle of the prior application example, the refrigerant flows into the second evaporator by the action of the dynamic pressure of the refrigerant flowing out from the ejector, so that the refrigerant flowing out from the ejector has a predetermined flow velocity. Note that the flow rate of the refrigerant flowing out of the ejector is specifically the flow rate Vdif of the refrigerant flowing out of the diffuser portion of the ejector.

  Therefore, the present inventors investigated the relationship between the flow rate Vdif of the refrigerant flowing out of the diffuser section and the cycle efficiency COP in order to further improve the COP. FIG. 9 is a graph showing the investigation results. According to FIG. 9, the value of the cycle efficiency COP changes to have a peak (maximum value) as the flow rate Vdif of the refrigerant flowing out of the diffuser section changes.

  The reason is that although the flow rate of the refrigerant flowing into the first and second evaporators can be increased as the flow velocity Vdif of the refrigerant flowing out of the diffuser section, the pressure loss of the first and second evaporators is the refrigerant flow rate. This is because when Vdif increases, the driving power of the compressor is increased and the COP is lowered.

  That is, when the pressure loss of the first evaporator increases and becomes larger than the pressure increase amount in the diffuser portion of the ejector, the increase amount of the suction refrigerant pressure of the compressor with respect to the refrigerant evaporation pressure in the second evaporator decreases. For this reason, the COP improvement effect by the drive power reduction of the above-mentioned compressor cannot fully be acquired.

  Therefore, in the cycle of the prior application example, in order to operate the cycle while exhibiting high cycle efficiency COP, the flow rate Vdif of the refrigerant flowing out of the diffuser section is operated so as to approach the optimum flow rate Vdif-max at which the cycle efficiency COP reaches a peak. There is a need to.

  In view of the above points, the present invention brings the flow rate Vdif of the refrigerant flowing out of the diffuser part closer to the optimum flow rate Vdif-max in the ejector-type refrigeration cycle in which the refrigerant flows into the plurality of evaporators from the branch part arranged on the downstream side of the ejector. Therefore, the purpose is 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 high-temperature and high-pressure refrigerant discharged from the compressor (11), and a radiator (12) The refrigerant is sucked from the refrigerant suction port (15b, 30b) by the flow of the high-speed jet refrigerant jetted from the nozzle portion (15a, 30a) for decompressing and expanding the downstream refrigerant, and the jet refrigerant and the refrigerant suction port ( Ejectors (15, 30) for increasing the pressure of the mixed refrigerant with the suction refrigerant sucked from 15b, 30b) at the diffuser portions (15d, 30d), and the branch portions for branching the flow of the refrigerant flowing out of the diffuser portions (15d, 30d) (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 the other branched at the branch portion (100) Refrigerant It evaporated refrigerant suction port (15b, 30b) second evaporator flow out to the upstream side (18) and provided with,
The branch part (100) branches the refrigerant so that the dynamic pressure of the refrigerant flowing out of the diffuser part (15d, 30d) is maintained, and the second evaporator (18) is connected to a range where the dynamic pressure acts on the inside. The first feature is an ejector refrigeration cycle in which the pressure loss (ΔP1drop) in the first evaporator (17) is equal to or lower than the pressure increase amount (ΔPej) in the diffuser section (15d, 30d).

  According to this, since the pressure loss (ΔP1drop) is equal to or lower than the pressure increase amount (ΔPej), the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) can be brought close to the optimum flow rate (Vdif−max). High cycle efficiency can be exhibited as a whole cycle.

  That is, when the pressure loss (ΔP1drop) is larger than the pressure increase amount (ΔPej), the recovered energy recovered by the nozzle portions (15a, 30a) of the ejectors (15, 30) passes through the first evaporator (17). It will be consumed by the pressure loss.

  For this reason, the refrigerant | coolant suction pressure of a compressor (11) cannot fully be raised with respect to the refrigerant | coolant evaporation pressure in a 2nd evaporator (18), but the cycle efficiency improves by the drive power reduction of a compressor (11). It becomes difficult to obtain the effect.

  On the other hand, if the pressure loss (ΔP1drop) is equal to or lower than the pressure increase amount (ΔPej), the driving power of the compressor (11) can be reduced to improve the cycle efficiency. That is, the diffuser part outflow refrigerant flow velocity (Vdif) can be brought close to the optimum flow velocity (Vdif-max).

Further, in the ejector refrigeration cycle having the first feature, the flow rate (Vdif), the pressure increase amount (ΔPej), and the discharge refrigerant flow rate (Gn) discharged from the compressor (11) in the diffuser section (15d, 30d). The relationship among the passage length (Le) of the refrigerant passage from the refrigerant inlet to the refrigerant outlet of the first evaporator (17), the passage diameter (De) of the refrigerant passage, and the pipe friction coefficient (λ) of the refrigerant passage is
Vdif ≦ 2π × ΔPej / Gn × (De / 2) ² × (De / Le) / λ
It may be.

  In the present invention, the flow path length (Le), the passage diameter (De), and the pipe friction coefficient (λ) do not mean only values derived from the exact refrigerant passage dimensions. As will be described later in the embodiment, for example, when the refrigerant passage of the first evaporator (17) has a complicated passage shape, the flow path length (Le) when the refrigerant is replaced with an equivalent pipe, It means to include the values of the path diameter (De) and the pipe friction coefficient (λ).

Here, when the refrigerant passage of the first evaporator (17) is approximated as a pipeline, generally, the pressure loss due to the pipeline friction of the first evaporator (17) can be expressed by the following formula F1.
ΔP1drop = λ × (Le / De ) × (ρdif × Vdif 2/2) ... (F1)
Note that ρdif is the density of the refrigerant at the outlet of the diffuser section (15d, 30d).

Further, the flow rate of the refrigerant passing through the first evaporator (17) can be expressed by the following formula F2.
Gn = ρdif × Vdif × π (De / 2) ² (F2)
In addition, the refrigerant | coolant flow rate which passes a 1st evaporator (17) is equal to the discharge refrigerant | coolant flow rate (Gn) of a compressor (11).

In the ejector refrigeration cycle having the first feature, the pressure loss (ΔP1drop) is equal to or lower than the pressure increase amount (ΔPej).
ΔP1drop ≦ ΔPej (F3)
It means that the relation of the above formula F3 is established.

Therefore, when Formulas F1 and F2 are applied to the relationship of Formula F3, Formula F4 can be derived below.
Vdif ≦ 2π × ΔPej / Gn × (De / 2) ² × (De / Le) / λ (F4)
Therefore, if the expression F4 is satisfied, the expression F3 can be satisfied more specifically.

  As a result, the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) can be brought close to the optimum flow rate (Vdif-max), and high cycle efficiency can be exhibited as a whole cycle.

  In Formula F4, the lower limit value of the flow velocity (Vdif) of the refrigerant flowing out of the diffuser sections (15d, 30d) is not defined. However, in the ejector refrigeration cycle having the first characteristic, the second evaporator (18) is provided inside. The dynamic pressure of the refrigerant flowing out of the diffuser portion (15d, 30d) acts. Accordingly, the flow rate (Vdif) of the refrigerant flowing out of the diffuser section (15d, 30d) is a flow rate in a range where dynamic pressure can be applied to the inside of the second evaporator (18), and the lower limit is determined by this range. it can.

The present invention also includes a compressor (11) that compresses and discharges the refrigerant, a radiator (12) that radiates heat from the high-temperature and high-pressure refrigerant discharged from the compressor (11), and a refrigerant that is downstream of the radiator (12). The refrigerant is sucked from the refrigerant suction port (15b, 30b) by the flow of the high-speed jet refrigerant jetted from the nozzle portion (15a, 30a) for decompressing and expanding, and sucked from the jet refrigerant and the refrigerant suction port (15b, 30b). Ejector (15, 30) for increasing the pressure of the mixed refrigerant with the sucked refrigerant at the diffuser section (15d, 30d), the branch section (100) for branching the flow of the refrigerant flowing out of the ejector (15, 30), and the branch section The first evaporator (17) that evaporates one refrigerant branched at (100) and flows out to the suction side of the compressor (11), and the other refrigerant branched at the branch portion (100) evaporates to become a refrigerant Suction port (15b Comprises 30b) second evaporator flow out to the upstream side and (18),
The total refrigeration capacity (Qre), which is the total value of the refrigeration capacity (Q1) exhibited by the first evaporator (17) and the refrigeration capacity (Q2) exhibited by the second evaporator (18), and the mixed refrigerant The ejector-type refrigeration cycle, which adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser sections (15d, 30d) based on the dryness (Xmix) of the second, is a second feature.

  According to this, since the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) is adjusted based on the total refrigeration capacity (Qre) and the dryness (Xmix) of the mixed refrigerant, the diffuser part (15d, 30d). ) The cycle can be operated so that the flow rate (Vdif) of the effluent refrigerant approaches the optimum flow rate (Vdif-max), and high cycle efficiency can be exhibited as a whole cycle.

  In the ejector refrigeration cycle having the second feature, the refrigerant flow rate adjusting means (11a, 12a, 17a, 30e, 30f, 31, 32) for adjusting the flow rate (Vdif) of the refrigerant flowing out of the diffuser portion (15d, 30d). And a refrigerant flow rate control means (20) for controlling the operation of the refrigerant flow rate adjustment means (11a, 12a, 17a, 30e, 30f, 31, 32), and the refrigerant flow rate control means (20) Qre) and the dryness (Xmix) of the mixed refrigerant, the refrigerant flow rate adjusting means (11a, 12a, 17a, so that the flow rate (Vdif) of the refrigerant flowing out of the diffuser section (15d, 30d) is within a predetermined range. 30e, 30f, 31, 32) may be controlled.

  According to this, the diffuser part outflow refrigerant | coolant flow velocity (Vdif) can be made to approach the optimal flow velocity (Vdif-max) still more reliably.

  More specifically, the refrigerant flow rate control means (20) includes an increase in refrigerant enthalpy (ΔHcomp) increased in the compressor (11), a decrease in refrigerant enthalpy (ΔHcond) in the radiator (12), The total refrigeration capacity (Qre) determined by the enthalpy (H1out) of the first evaporator (17) outlet refrigerant and the enthalpy (H2out) of the second evaporator (18) outlet side refrigerant, and the dryness of the injected refrigerant (Xnoz) Diffuser section (15d, 30d) Diffuser section (15d, 30d) diffusing section (15d, 30d) outflow refrigerant calculated based on dryness (Xmix) of mixed refrigerant obtained by enthalpy (Hdif) of refrigerant flowing out and enthalpy (Hmix) of mixed refrigerant Refrigerant flow rate adjusting means (11a, 12a, so that the flow rate (Vdif) is in a predetermined range. 17a, 30e, 30f, 31, 32) need only be controlled.

  In the ejector-type refrigeration cycle of the second feature described above, specifically, the refrigerant flow rate adjusting means includes discharge capacity changing means (11a) that changes the refrigerant discharge capacity of the compressor (11), and the refrigerant flow speed. The control means (20) may adjust the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the discharge capacity changing means (11a).

  Further, the refrigerant flow rate adjusting means includes a heat dissipation capacity changing means (12a) that changes the refrigerant heat dissipation capacity of the radiator (12), and the refrigerant flow rate control means (20) operates the heat dissipation capacity changing means (12a). By controlling, the flow velocity (Vdif) of the refrigerant flowing out of the diffuser portion (15d, 30d) may be adjusted.

  Further, the refrigerant flow rate adjusting means is constituted by an evaporation capacity changing means (17a) for changing the refrigerant evaporation capacity of at least one of the first evaporator (17) and the second evaporator (18). 20) may adjust the flow velocity (Vdif) of the refrigerant flowing out of the diffuser sections (15d, 30d) by controlling the operation of the evaporation capacity changing means (17a).

  Further, the refrigerant flow rate adjusting means includes throttle opening degree changing means (30e, 30f) for changing the throttle passage area of the nozzle portion (30a), and the refrigerant flow rate control means (20) is the throttle opening degree changing means (30e). 30f), the flow rate (Vdif) of the refrigerant flowing out of the diffuser section (30d) may be adjusted.

  Further, the refrigerant flow rate adjusting means includes a high-pressure side throttle mechanism (31) disposed downstream of the radiator (12) and upstream of the ejectors (15, 30), and the refrigerant flow rate control means. (20) may adjust the flow velocity (Vdif) of the refrigerant flowing out of the diffuser portions (15d, 30d) by controlling the operation of the high-pressure side throttle mechanism (31).

  Further, the second evaporator (18) is disposed between the plurality of evaporators (18a, 18b) connected in series and the plurality of evaporators (18a, 18b), and the low pressure side throttle for decompressing and expanding the refrigerant. The refrigerant flow rate control means (20) has a mechanism (32), and adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the low pressure side throttle mechanism (32). It may be.

  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 the present embodiment, a swash plate type variable displacement compressor capable of continuously variably controlling the discharge capacity by a control signal from an air conditioning controller 20 described later is employed 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.

  Therefore, the electromagnetic capacity control valve 11a constitutes the discharge capacity changing means of the present embodiment, and the valve opening degree (the ratio between the suction refrigerant and the discharge refrigerant) is adjusted by the control current In of the air conditioning control device 20 described later. Is done. 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 changing 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. Further, when an electric compressor is employed as the compressor 11, the electric motor serves as a discharge capacity changing 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 V1 output from the 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. Accordingly, the cooling fan 12a constitutes a heat radiation capacity changing unit of the present 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 exchanged, A double-pipe heat exchanger configuration in which the high-pressure side refrigerant flow path 13a is arranged inside the outer pipe forming the side refrigerant flow path 13b can be adopted.

  An ejector 15 is connected to the outlet side of the high-pressure side refrigerant flow path 13 a of the internal heat exchanger 13. 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 high-pressure refrigerant flowing out from the high-pressure side refrigerant flow path 13a of the internal heat exchanger 13 and further depressurizes the refrigerant, and refrigerant injection of the nozzle part 15a It has a refrigerant suction port 15b that is disposed so as to communicate with the port and sucks 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 the hole used as a some refrigerant path in a rectangular parallelepiped metal block or resin block.

  In addition, in FIG. 1, although the cycle structure of this embodiment is shown schematically, the diffuser part 15d of the ejector 15, the refrigerant distributor 16, and the second evaporator 18 are connected so as to be close to each other directly or by a short pipe. It is desirable to do. 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 blower fan 17a is an electric blower whose rotational speed (amount of blown air) is controlled by a control voltage V2 output from the air conditioning control device 20 described later. And the refrigerant | coolant evaporation capability of the 1st evaporator 17 is adjusted with the amount of blowing air of this ventilation fan 17a.

  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.

  Therefore, the blower fan 17a is an evaporation capacity changing means capable of changing the refrigerant evaporation capacity of both the first evaporator 17 and the second evaporator 18. 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, the outline of the electric control unit of the present embodiment will be described with reference to FIG. FIG. 2 is a block diagram of the electric control unit of the present embodiment. The air conditioning control device 20 is composed of a well-known microcomputer including a CPU, ROM, RAM and the like and its peripheral circuits, and performs various calculations and processes based on a control program stored in the ROM, and the various actuators described above. Control the operation of

  Specifically, an electromagnetic capacity control valve 11a as discharge capacity changing means, a cooling fan 12a as heat dissipation capacity changing means, and a blower fan 17a as evaporation capacity changing means are connected to the output side of the air conditioning control device 20. The Therefore, the air conditioning control device 20 of the present embodiment functions as a refrigerant flow rate control means.

  Further, detection signals of a detection means group for detecting the following physical quantities are input to the input side of the air conditioning control device 20.

  As a temperature sensor that is a temperature detection means in the detection means group, the first temperature sensor 21a that detects the temperature T1out of the refrigerant at the outlet of the first evaporator 17 and the second temperature that detects the temperature T2out of the refrigerant at the outlet of the second evaporator 18 are used. The sensor 22a, the third temperature sensor 23a for detecting the temperature Tnozi of the refrigerant at the inlet of the nozzle portion 15a, the fourth temperature sensor 24a for detecting the temperature Tdif of the refrigerant flowing out of the diffuser portion 15d, and the fifth temperature for detecting the intake refrigerant temperature Tcompi of the compressor 11 A sensor 25a, a sixth temperature sensor 26a for detecting the refrigerant discharge refrigerant temperature Tcompo, a seventh temperature sensor 27a for detecting the radiator 12 outlet refrigerant temperature Tcondo, and the like are connected.

  In addition, as a pressure sensor which is a pressure detection means in the detection means group, a first pressure sensor 21b for detecting the pressure P1out of the refrigerant at the outlet of the first evaporator 17 and a pressure P2out of the refrigerant at the outlet of the second evaporator 18 are detected. A second pressure sensor 22b, a third pressure sensor 23b for detecting the pressure Pnozi of the inlet refrigerant of the nozzle portion 15a, a fourth pressure sensor 24b for detecting the pressure Pdif of the refrigerant flowing out of the diffuser portion 15d, and a first pressure detecting the compressor 11 intake refrigerant pressure Pcompi. A fifth pressure sensor 25b, a sixth pressure sensor 26b for detecting the compressor 11 discharge refrigerant pressure Pcompo, a seventh pressure sensor 27b for detecting the radiator 12 outlet refrigerant temperature Pcondo, and the like are connected.

  Further, a rotation speed sensor 28 which is a rotation speed detection means for detecting the rotation speed Nc of the compressor 11 is connected. In addition to the detection signals of the sensors, various operation signals from the operation panel 29 are also 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 the present embodiment having the above-described configuration will be described with reference to the Mollier diagram of FIG. 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. Gas-phase refrigerant of high temperature and high pressure discharged from the compressor (A ₁ point in FIG. 3) flows into the radiator 12, radiates heat by heat exchange with the blown been blown air (outside air) from the cooling fan (Fig. 3 A ₁ point → B 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. the cooling state (B ₁ point in FIG. 3 → C ₁ point).

The refrigerant that has flowed out of the high-pressure side refrigerant flow path 13 a of the internal heat exchanger 13 flows into the nozzle portion 15 a of the ejector 15. Refrigerant flowing into the nozzle portion 15a of the ejector 15 is isentropically depressurized expand (C ₁ point in FIG. 3 → D ₁ point).

  And the pressure energy of a refrigerant | coolant is converted into a velocity energy at the time of this decompression | expansion expansion, and a refrigerant | coolant is injected at high speed from the refrigerant | coolant injection port 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 part 15a becomes the mixed refrigerant is mixed in the mixing portion 15c of the ejector 15 (D ₁ point in FIG. 3 → E ₁ point ) And flows into the diffuser portion 15d.

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 (E ₁ point in FIG. 3 → F ₁ point). That is, the pressure difference between point F _{1 and} point E ₁ in FIG. 3 is the pressure increase amount ΔPej in the diffuser portion 15d.

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 (F ₁ point in FIG. 3 → G ₁ point). That is, the pressure difference between the F ₁ point and the G ₁ point in FIG. 3 is the pressure loss ΔP1drop in the first evaporator 17.

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 ( (G ₁ point → H ₁ point in FIG. 3). The refrigerant heated in the low-pressure side refrigerant flow path 13b of the internal heat exchanger 13 is sucked into the compressor 11 and compressed again (point H ₁ → point A _{1 in} FIG. 3).

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 (F ₁ point → I ₁ point in FIG. 3). The refrigerant flowing out of the second evaporator 18 is sucked into the ejector 15 from the refrigerant suction port 15b (I ₁ point → E ₁ point in FIG. 3).

  At this time, the air-conditioning control device 20 controls the refrigerant discharge capacity of the compressor 11 so that the pressure loss ΔP1drop in the first evaporator 17 is equal to or lower than the pressure increase amount ΔPej in the diffuser portion 15d.

  Specifically, the air-conditioning control device 20 has a flow path length Le, a representative passage diameter De, a pipe friction, and a pressure loss characteristic equivalent to those of the refrigerant passage from the refrigerant inlet to the refrigerant outlet of the first evaporator 17. The coefficient λ is stored, and the refrigerant discharge capacity of the compressor 11 is controlled from these values so as to satisfy the above-described formula F4.

As described above, when the expression F4 is satisfied, the pressure loss ΔP1drop in the first evaporator 17 becomes equal to or less than the pressure increase amount ΔPej in the diffuser portion 15d. Below, Formula F4 is shown again.
Vdif ≦ 2π × ΔPej / Gn × (De / 2) ² × (De / Le) / λ (F4)
Of the formula F4, boosting amount ΔPej the diffuser portion 15d, the second evaporator 18 (the pressure of the E ₁ point of FIG. 3) pressure P2out outlet refrigerant diffuser portion 15d pressure of the refrigerant flowing from Pdif (F ₁ point 3 The pressure can be calculated from the difference between the

  The discharge refrigerant flow rate Gn of the compressor 11 is determined by the rotation speed Nc of the compressor 11, the discharge capacity of the compressor 11 determined from the control signal In output to the electromagnetic capacity control valve 11a, and the air conditioning control device 20 in advance. It can be calculated by accumulating the compressor 11 intake refrigerant density calculated from the stored volume efficiency ηv of the compressor 11, the compressor 11 intake refrigerant temperature Tcompi, and the compressor 11 intake refrigerant pressure Pcompi.

  Next, calculation of the flow velocity Vdif of the refrigerant flowing out of the diffuser portion 15d will be described. In the air conditioning control device 20, the total refrigerating capacity Qre, which is the total value of the refrigerating capacity Q1 exhibited by the first evaporator 17 and the refrigerating capacity Q2 exhibited by the second evaporator 18, and the nozzle portion 15a are injected. Vdif is calculated based on the dryness Xmix of the mixed refrigerant of the injected refrigerant and the suction refrigerant sucked from the refrigerant suction port 15b.

First, the total refrigeration capacity Qre can be calculated by the following formula F5.
Qre = Gn × (ΔHcond−ΔHcomp) (F5)
Here, ΔHcond is an enthalpy reduction amount in the radiator 12 and ΔHcomp is an enthalpy increase amount in the compressor 11.

  ΔHcond is the enthalpy Hcompo of the compressor 11 discharge refrigerant calculated from the compressor 11 discharge refrigerant temperature Tcompo and the compressor 11 discharge refrigerant pressure Pcompo, and the radiator 12 outlet refrigerant temperature Tcondo and the radiator 12 outlet refrigerant temperature Pcondo. From the enthalpy Hcondo of the refrigerant at the outlet of the radiator 12 calculated from

  On the other hand, ΔHcomp can be calculated from the enthalpy Hcompi of the refrigerant discharged from the compressor 11 and the enthalpy Hcompi of the refrigerant sucked by the compressor 11 calculated from the compressor 11 intake refrigerant temperature Tcompi and the compressor 11 intake refrigerant pressure Pcompi.

Further, Qre can be expressed by the following formula F6.
Qre = Gn × (H1out−Hein) + Ge × (H2out−Hein) (F6)
Here, Hein is the enthalpy of the refrigerant that actually flows into the first and second evaporators 17 and 18 and is the enthalpy of the refrigerant branched at the branching section 100. Accordingly, the value is very close to the enthalpy Hdif of the refrigerant flowing out of the diffuser portion 15d.

  Further, Ge is the flow rate of refrigerant passing through the second evaporator 18, H1out is the enthalpy of the first evaporator 17 outlet refrigerant, and H2out is the enthalpy of the second evaporator 18 outlet refrigerant. H1out can be calculated from the temperature T1out of the first evaporator 17 outlet refrigerant and the pressure P1out of the first evaporator 17 outlet refrigerant, and H2out is the temperature T2out of the second evaporator 18 outlet refrigerant and the pressure of the second evaporator 18 outlet refrigerant. It can be calculated from P2out.

  Therefore, by calculating the total refrigeration capacity Qre by Formula F5 and Formula F6, the passing refrigerant flow rate Ge of the second evaporator 18 and the enthalpy Hein of the refrigerant flowing into the first and second evaporators 17 and 18 are used as variables. An expression is derived.

Next, the dryness Xmix of the mixed refrigerant of the injection refrigerant and the suction refrigerant can be expressed by the following formula F7.
Xmix = (Gn · Xnoz + Ge) / (Gn + Ge) (F7)
Here, Xnoz is the dryness of the refrigerant injected from the nozzle portion 15a of the ejector 15.

  Xnoz is determined in advance from the enthalpy Hnozi of the nozzle portion 15a inlet refrigerant determined from the nozzle portion 15a inlet refrigerant temperature Tnozi and the nozzle portion 15a inlet refrigerant pressure Pnozi, the design specifications of the nozzle portion 15a, and the like. Is calculated from the nozzle efficiency ηnoz stored in.

Further, the dryness Xmix can also be expressed by the following formula F8.
Xmix = ((Hein−Hldif) × (Sgdif−Sldif) / (Hgdif−Hldif) − (Slmix−Sldif)) / (Sgmix−Slmix) (F8)
Here, Hldif is the enthalpy of the saturated liquid phase refrigerant among the refrigerant flowing out of the diffuser portion 15d, and Hgdif is the enthalpy of the saturated gas phase refrigerant among the refrigerant flowing out of the diffuser portion 15d.

  Sldif is the entropy of the saturated liquid phase refrigerant in the refrigerant flowing out of the diffuser unit 15d, and Sgdif is the entropy of the saturated gas phase refrigerant in the refrigerant out of the diffuser unit 15d.

  These Hldif, Hgdif, Sldif, and Sgdif are all calculated from the calculated enthalpy Hdif by calculating the enthalpy Hdif of the diffuser part 15d outflow refrigerant from the temperature Tdif of the diffuser part 15d outflow refrigerant and the pressure Pdif of the diffuser part 15d outflow refrigerant. It can be determined with reference to the characteristics of the refrigerant (corresponding to the Mollier diagram) stored in advance in the air conditioning controller 20.

  Slmix is the entropy of the saturated liquid phase refrigerant in the refrigerant at the mixing section 15c of the ejector 15, and Sgmix is the entropy of the saturated gas phase refrigerant in the refrigerant at the mixing section 15c. The Slmix and Sgmix can be determined by calculating the enthalpy Hmix of the refrigerant at the inlet of the mixing unit 15c and referring to the characteristics of the refrigerant stored in the air conditioning controller 20 in advance from the calculated Hmix.

  Specifically, the enthalpy Hnozo of the nozzle part 15a inlet refrigerant, the enthalpy Hnozo of the nozzle part 15a outlet refrigerant determined from the decompression characteristics of the nozzle part 15a stored in the air conditioning control device 20 in advance, the refrigerant of the outlet of the second evaporator 18 The enthalpy Hmix of the inlet of the mixing unit 15c is determined from the flow rate ratio of the enthalpy H2out and the discharge refrigerant flow rate Gn to the passing refrigerant flow rate Ge.

  Therefore, the passage refrigerant flow rate Ge of the second evaporator 18 and the enthalpy Hein of the refrigerant flowing into the first and second evaporators 17 and 18 are changed by obtaining the dryness Xmix of the mixed refrigerant by the equations F7 and F8. The following formula is derived.

  Next, the air-conditioning control device 20 calculates the passing refrigerant flow rate Ge and the enthalpy Hein by solving the above two equations using the passing refrigerant flow rate Ge and the enthalpy Hein as variables.

Further, the flow rate Gdif of the refrigerant flowing out of the diffuser portion 15d is calculated by the following formula F9.
Gdif = Gn + Ge (F9)
Further, the air-conditioning control device 20 calculates the density ρdif of the refrigerant flowing out from the enthalpy Hein by the diffuser unit 15, and calculates the flow velocity Vdif of the refrigerant discharged from the diffuser unit 15d by the following formula F10.
Vdif = Gdif / (ρdif × Adif) (F10)
Here, Adif is the refrigerant passage area at the outlet of the diffuser portion 15d, and is a value stored in the air conditioning controller 20 in advance from the design specifications of the ejector 15.

  The air conditioning control device 20 controls the operation of the electromagnetic capacity control valve 11a, which is a discharge capacity changing means of the compressor 11, so that Vdif calculated as described above satisfies the formula F4.

  In this embodiment, since it operates as described above, the first evaporator 17 and the second evaporator 18 can simultaneously exert a cooling action. At this time, the suction action of the ejector 15 can reduce the refrigerant evaporation pressure (refrigerant evaporation temperature) of the second evaporator 18 with respect to the refrigerant evaporation pressure (refrigerant evaporation temperature) of the first evaporator 17. 1. The temperature difference between the refrigerant evaporation temperature of the first and second evaporators 17 and 18 and the indoor blown air of the blower fan 17a can be secured, and the indoor blown air can be efficiently cooled.

  Furthermore, 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 refrigerant can flow into the second evaporator 18 reliably by applying the dynamic pressure of the refrigerant flowing out from the portion 15d.

  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 air conditioning control device 20 controls the operation of the electromagnetic capacity control valve 11a, which is a discharge capacity changing means of the compressor 11, so that the flow velocity Vdif of the refrigerant at the outlet of the diffuser 15d satisfies Formula F4. Therefore, the flow velocity Vdif of the refrigerant at the outlet of the diffuser portion 15d can be made close to the optimum flow velocity Vdif-max shown in FIG. As a result, the driving power of the compressor 11 can be reliably reduced and high cycle efficiency can be exhibited as a whole cycle.

  In other words, the ejector refrigeration cycle of the present embodiment includes the compressor 11 intake refrigerant temperature Tcompi, the compressor 11 intake refrigerant pressure Pcompi, the compressor 11 discharge refrigerant temperature Tcompo, the compressor 11 discharge refrigerant pressure Pcompo, and the radiator 12 outlet refrigerant. Temperature Tcondo, radiator 12 outlet refrigerant temperature Pcondo, first evaporator 17 outlet refrigerant temperature T1out, first evaporator 17 outlet refrigerant pressure P1out, second evaporator 18 outlet refrigerant temperature T2out, second evaporator 18 outlet Refrigerant pressure P2out, nozzle section 15a inlet refrigerant temperature Tnozi, nozzle section 15a inlet refrigerant pressure Pnozi, diffuser section 15d refrigerant refrigerant temperature Tdif, diffuser section 15d refrigerant refrigerant pressure Pdif, compressor 11 rotation speed Nc, compression Volume efficiency ηv of machine 11 and nose of nozzle portion 15a Based on the physical quantity related to the efficiency Itanoz, by adjusting the flow rate Vdif of the diffuser portion 15d refrigerant flowing, so close to the optimal flow velocity Vdif-max, it can be operated while exhibiting high cycle efficiency of the entire cycle.

(Second and third embodiments)
In the second embodiment, in the cycle configuration of FIG. 1 similar to the first embodiment, the air-conditioning control device 20 is a heat dissipation capability changing unit so that Vdif calculated in the same manner as the first embodiment satisfies the formula F4. The operation of a certain cooling fan 12a is controlled. Therefore, in the second embodiment, the cycle can be operated while exhibiting high cycle efficiency as in the first embodiment.

  Further, in the third embodiment, the air-conditioning control device 20 changes the evaporation capacity so that Vdif calculated in the same manner as in the first embodiment satisfies the formula F4 in the cycle configuration of FIG. 1 similar to the first embodiment. The operation of the blower fan 17a as a means is controlled. Therefore, in the second embodiment, the cycle can be operated while exhibiting high cycle efficiency as in the first embodiment.

  Of course, the air conditioning control device 20 controls the operation of two or more of the discharge capacity changing means (electromagnetic capacity control valve 11a), the heat radiation capacity changing means (heat radiation fan 12a), and the evaporation capacity changing means (air blowing fan 17a). Thus, the flow velocity Vdif of the refrigerant flowing out of the diffuser portion 15d may be adjusted.

(Fourth embodiment)
In the first embodiment described above, the fixed ejector 15 in which the throttle passage area of the nozzle portion 15a is fixed is acting, but in this embodiment, as shown in the overall configuration diagram of FIG. 4, the nozzle portion 30a. The variable ejector 30 is configured so that the throttle passage area can be changed.

  The variable ejector 30 includes a nozzle portion 30a that decompresses and expands the high-pressure refrigerant that has flowed out of the internal heat exchanger 13, a needle valve 30e that is disposed inside the nozzle portion 30a and adjusts the opening degree of the throttle passage area of the nozzle portion 30. An electric actuator 30f that displaces the needle valve 30e in the axial direction of the nozzle portion 30a is provided.

  In the present embodiment, the throttle valve changing means is constituted by the needle valve 30e and the electric actuator 30f. As the electric actuator 30f, for example, a motor actuator such as a stepping motor or an electromagnetic solenoid mechanism can be employed. In addition, the electric actuator 30 f is driven and controlled by a control signal output from the air conditioning controller 20.

  Further, similarly to the ejector 15 of the first embodiment, the variable ejector 30 includes a refrigerant suction port 30b that sucks the refrigerant that has flowed out of the second evaporator 18, a high-speed refrigerant flow that is injected from the nozzle portion 30a, and a refrigerant suction port. The mixing unit 30c mixes with the suctioned refrigerant sucked from 30b, and the diffuser unit 30d forms a pressure increasing unit. Other configurations are the same as those of the first embodiment.

  Next, the operation of this embodiment in the above configuration will be described. In the present embodiment, similarly to the first embodiment, the flow velocity Vdif of the refrigerant flowing out of the diffuser section 30d is calculated, and the air conditioning control device 20 changes the throttle opening degree of the variable ejector 30 so that Vdif satisfies the above-described formula F4. The operation of the electric actuator 30f constituting the means is controlled.

  Therefore, even if the ejector refrigeration cycle of this embodiment is operated, the cycle can be operated while exhibiting high cycle efficiency as in the first embodiment. Further, the air conditioning control device 20 not only controls the operation of the throttle opening changing means of the variable ejector 30, but also the discharge capacity changing means, the heat radiation capacity changing means, and the evaporation capacity changing means described in the first to third embodiments. May be simultaneously controlled to adjust the flow velocity Vdif of the refrigerant flowing out of the diffuser portion 15d.

(Fifth embodiment)
In the present embodiment, as compared with the first embodiment, as shown in the overall configuration diagram of FIG. 5, the electric power that constitutes the high-pressure side throttle mechanism on the downstream side of the radiator 12 and on the upstream side of the ejector 15. A variable aperture mechanism 31 of the type is arranged.

  The variable throttle mechanism 31 is a decompression unit that decompresses and expands the high-pressure refrigerant on the downstream side of the radiator 12, and adjusts the flow rate of the refrigerant that flows to the downstream side (specifically, the nozzle portion 15a side of the ejector 15 described later). It is also a flow rate adjusting means. Therefore, in this embodiment, the variable throttle mechanism 31 constitutes the refrigerant flow rate adjusting means.

  The variable throttle mechanism 31 is an electric variable throttle mechanism in which the opening degree of the throttle passage is adjusted by a control signal output from the 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. Other configurations are the same as those of the first embodiment.

  When the ejector refrigeration cycle of this embodiment is operated, the flow rate Vdif of the refrigerant flowing out of the diffuser section 30d is calculated as in the first embodiment, and the air conditioning control device 20 is variable so that Vdif satisfies the above-described formula F4. The operation of the diaphragm mechanism 31 is controlled.

  Therefore, even if the ejector refrigeration cycle of this embodiment is operated, the cycle can be operated while exhibiting high cycle efficiency as in the first embodiment. Further, the air conditioning control device 20 not only controls the operation of the throttle opening changing means of the variable ejector 30, but also the discharge capacity changing means, the heat radiation capacity changing means, and the evaporation capacity changing means described in the first to third embodiments. May be simultaneously controlled to adjust the flow velocity Vdif of the refrigerant flowing out of the diffuser portion 15d.

(Sixth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 6, as the evaporator connected to the refrigerant suction port 15b of the ejector 15, two first and second evaporators 18a, 18b connected in series, and The second evaporator 18 is employed which is arranged between the two evaporators 18a and 18b and has an electric variable throttle mechanism 32 which is a low-pressure side throttle mechanism for decompressing and expanding the refrigerant. .

  First, the configuration of the variable aperture mechanism 32 is the same as that of the variable aperture mechanism 31 of the fifth embodiment. Moreover, in the 2nd evaporator 18 of this embodiment, it arrange | positions in order of the 1st evaporator 18a-> variable throttle mechanism 32-> 2nd evaporator 18b with respect to the flow direction of a refrigerant | coolant. That is, the first evaporator 18 is connected to the second outlet tube 16 c of the refrigerant distributor 16.

  Therefore, the dynamic pressure of the refrigerant flowing out of the diffuser portion 15d acts on the first evaporator 18a, so that the refrigerant can surely flow into the second evaporator 18. Furthermore, in the second evaporator 18 b downstream of the variable throttle mechanism 32, the refrigerant evaporation pressure (refrigerant evaporation temperature) can be reduced by the pressure reducing action of the variable throttle mechanism 32. Other configurations are the same as those of the first embodiment.

  Next, the operation of the present embodiment in the above configuration will be described with reference to the Mollier diagram of FIG. In FIG. 7, 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. 3 to 6.

First, when the compressor 11 is driven by a vehicle engine, the refrigerant branched into the branching section 100 and flowing into the first evaporator 17 is the compressor 11 → the radiator 12 → the internal heat exchanger as in the first embodiment. 13 circulates in the order of the high-pressure side refrigerant flow path 13 a → the ejector 15 → the first evaporator 17 → the low-pressure side refrigerant flow path 13 a of the internal heat exchanger 13 → the compressor 11, and the first evaporator 17 performs the cooling action. It works (A ₆ point → B ₆ point → C ₆ point → D ₆ point → E ₆ point → F ₆ point → G ₆ point → H ₆ point → A ₆ point in FIG. ₇ ).

On the other hand, the other refrigerant branched by the branch part 100 flows out of the second outlet pipe part 16 c and flows into the first evaporator 18 a of the second evaporator 18. The refrigerant that has flowed into the first evaporator 18a gradually reduces the pressure due to the pressure loss in the first evaporator 18a and the suction action of the ejector 15 while absorbing heat from the blown air of the blower fan 17a and evaporating ( F ₆ points of Figure 7 → F _'6 points).

Refrigerant flowing from the first evaporator section 18a is decompressed by the variable throttle mechanism 32 (F in FIG. 7 _'6 points → F'_'6 points), further flows into the second evaporator section 18b, the blower of the blower fan It absorbs heat from the air and evaporates (F ″ ₆ point → I ₆ point in FIG. 7). The refrigerant that has flowed out of the second evaporator 18b is sucked into the ejector 15 through the refrigerant suction port 15b (point I ₆ → E _{6 in} FIG. 7).

  At this time, similarly to the first embodiment, the flow velocity Vdif of the refrigerant flowing out of the diffuser portion 30d is calculated, and the air conditioning control device 20 controls the operation of the variable throttle mechanism 32 so that Vdif satisfies the above-described formula F4. Therefore, even if the ejector refrigeration cycle of this embodiment is operated, the cycle can be operated while exhibiting high cycle efficiency as in the first embodiment.

  Furthermore, in this embodiment, the dynamic pressure of the refrigerant flowing out from the diffuser portion 15d is applied to the first evaporator 18a disposed on the upstream side of the variable throttle mechanism 32 in the second evaporator 18, so that the second evaporator The refrigerant can surely flow into 18. Further, in the second evaporator 18b disposed on the downstream side of the variable throttle mechanism 32, the refrigerant evaporation pressure (refrigerant evaporation temperature) can be reduced by the pressure reducing action of the variable throttle mechanism 32.

  Of course, also in this embodiment, the air conditioning control device 20 not only controls the operation of the variable throttle mechanism 32, but also the discharge capacity changing means, the heat radiation capacity changing means, and the evaporation capacity changing explained in the first to third embodiments. The operation of the means may be controlled simultaneously.

(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 speed Vdif of the refrigerant flowing out of the diffuser portions 15d and 30d is calculated, and the air-conditioning control device 20 controls the operation of the refrigerant flow rate adjusting means so that this Vdif satisfies the relationship of Formula F4. However, of course, Vdif may be directly detected by a flow meter or the like, and the operation of the refrigerant flow rate adjusting means may be controlled so that the detected value satisfies the relationship of Formula F4.

  Under predetermined operating conditions, the pressure loss ΔP1drop in the first evaporator 17 and the pressure increase ΔPej in the diffuser portion 15d are directly detected without calculating or detecting Vdif, and the pressure loss ΔP1drop is equal to or less than the pressure increase ΔPej. As such, the operation of the refrigerant flow rate adjusting means may be controlled.

  Specifically, the pressure loss ΔP1drop may be a value obtained by subtracting the pressure P1out of the refrigerant at the outlet of the first evaporator 17 from the pressure Pdif of the refrigerant flowing out of the diffuser portion 15d. Further, the pressure increase amount ΔPej may be a value obtained by subtracting the pressure P2out of the refrigerant at the outlet of the second evaporator 18 from the pressure Pdif of the refrigerant flowing out of the diffuser portion 15d.

  (2) In the above-described embodiment, the necessary physical quantity required for calculating Vdif is calculated from the detection signal of the detection means group, but the calculation method of the required physical quantity is not limited to this. For example, as shown in the chart of FIG. 8, the required physical quantity can be estimated or calculated from other physical quantities. Note that the left column of FIG. 8 shows the required physical quantity, and the right column shows an example of an available physical quantity that can be used to calculate or estimate the corresponding required physical quantity.

  For example, in the above-described embodiment, the rotational speed Nc of the compressor 11 is directly detected by the rotational speed sensor 28. However, in the compressor 11 to which the driving force is transmitted from the vehicle travel engine as in the present embodiment. The rotation speed Nc of the compressor 11 can be calculated from the engine rotation speed. Therefore, the engine speed may be detected instead of detecting the speed Nc of the compressor 11.

  The first evaporator 17 outlet refrigerant pressure P1out may be estimated from the first evaporator 17 outlet refrigerant temperature T1out. Further, the refrigerant temperature T1out at the outlet of the first evaporator 17 may be estimated from the blown air temperature immediately after being cooled by the first evaporator 17, assuming that the refrigerant temperature is equal to the refrigerant evaporation temperature of the first evaporator 17. The same applies to the second evaporator 18.

  Moreover, the refrigerant temperature in the accumulator 19 may be substituted for the first evaporator 17 outlet refrigerant temperature T1out. The temperature of the space to be cooled may be substituted for the second evaporator 18 outlet refrigerant temperature T2out. The radiator 12 outlet refrigerant pressure Pcondo may be estimated from the radiator 12 outlet refrigerant temperature Tcondo. Furthermore, the radiator 12 outlet refrigerant temperature Tcondo may be estimated from the cooling fan 12a blown air temperature immediately after passing through the radiator 12.

  The nozzle portion 15 a inlet refrigerant temperature Tnozi may be estimated from the radiator 12 outlet refrigerant temperature Tcondo and the heat exchange amount in the internal heat exchanger 13. Further, in a cycle in which a receiver for separating the gas-liquid refrigerant and storing excess refrigerant is provided on the outlet side of the radiator 12, the refrigerant temperature in the receiver may be used as the radiator 12 outlet refrigerant temperature Tcondo.

  Further, the enthalpy increase amount ΔHcomp of the refrigerant increased in the compressor 11 may be estimated from the compressor 11 driving power (for example, engine output amount, input power amount, etc.) and the discharge refrigerant flow rate Gn of the compressor 11. The enthalpy reduction amount ΔHcond of the refrigerant reduced in the radiator 12 may be estimated from the air volume of the cooling fan 12a and the temperature difference between the radiator 12 intake air temperature (outside temperature) and the radiator 12 blown air temperature.

  Further, the enthalpy H1out of the first evaporator 17 outlet refrigerant may be estimated from the degree of superheat of the first evaporator 17 outlet refrigerant. The enthalpy H2out of the second evaporator 18 outlet refrigerant may be estimated from the degree of superheat of the second evaporator 18 outlet refrigerant. The enthalpy Hnozi and the entropy Snozi of the refrigerant at the inlet of the nozzle 15a may be estimated from the degree of superheat of the refrigerant at the inlet of the nozzle 15a.

  (3) In the overall configuration diagrams of FIGS. 1 and 4 to 6 in the above-described embodiment, each detection unit is schematically arranged, but the arrangement of each detection unit is not limited to this. As long as it is a part which can detect a desired physical quantity, it may be arranged anywhere. For example, the fourth temperature sensor 24a for detecting the temperature Tdif of the refrigerant flowing out of the diffuser unit 15d and the fourth pressure sensor 24b for detecting the pressure Pdif of the refrigerant flowing out of the diffuser unit 15d may be attached to the refrigerant distributor 16.

  (4) In the above-described embodiment, the heat dissipation capability changing means is configured by the cooling fan 12a, but the heat dissipation capability changing means is not limited to this. For example, as a heat dissipation capability changing means, a heat dissipation capability 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, you may adjust the heat dissipation capability by providing the interruption | blocking mechanism (damper mechanism) which interrupts | blocks the flow of the ventilation air of the cooling fan 12a.

  Further, the evaporation capacity changing means may be adjusted by adopting the same bypass mechanism, blocking mechanism, and the like. Further, in the present embodiment, the blower fan 17a constitutes the evaporation ability changing means for simultaneously changing the evaporation ability of both the first and second evaporators 18, but the evaporation ability of both the evaporators 17, 18 is constituted. You may employ | adopt the evaporation capability change means to change each independently.

  For example, a cooling target space may be cooled by the first evaporator 17 and the second evaporator 18, and a dedicated blower fan may be provided for both the evaporators 17 and 18. Furthermore, when the first evaporator 17 is abolished and the space to be cooled is cooled only by the second evaporator 18, a blower fan that blows air only to the second evaporator 18 may be provided.

  (5) In the above-described embodiment, the first evaporator 17 and the second evaporator 18 are assembled into an integral 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.

  (6) Although the refrigerant distributor 16 having a T-shaped three-way joint structure is employed in the above-described embodiment, 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.

  (7) In the above-described embodiment, an example in which the ejector refrigeration cycle 10 of the present invention is applied to a vehicle air conditioner and a vehicle refrigeration / refrigeration apparatus has been described, but the application of the present invention is not limited to the above. . For example, the present invention may be applied to commercial refrigerators, household refrigerators, vending machine cooling devices, showcases with a refrigeration function, and the like. Further, carbon dioxide or HC refrigerant may be employed as the refrigerant in addition to the fluorocarbon refrigerant.

  (8) In the above-described 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. In contrast, the first and second evaporators 17 and 18 are configured as outdoor heat exchangers that absorb heat from a heat source such as outside air, and the radiator 12 is heated fluid such as air or water. You may apply this invention to the heat pump cycle comprised as an indoor side heat exchanger which heats.

It is a whole block diagram of the ejector-type refrigerating cycle of 1st-3rd embodiment. It is a block diagram of the electrical machinery control part of the ejector type refrigerating cycle of a 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 4th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 5th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 6th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of 6th Embodiment. It is a graph which shows the physical quantity which can be used for calculating or estimating a required physical quantity. It is a graph which shows the relationship between the flow rate of a diffuser part exit refrigerant | coolant, and cycle efficiency.

Explanation of symbols

DESCRIPTION OF SYMBOLS 11 ... Compressor, 11a ... Electromagnetic capacity control valve, 12 ... Radiator, 12a ... Cooling fan,
15 ... Ejector, 15a, 30a ... Nozzle part, 15b, 30b ... Refrigerant suction port,
15d, 30d ... Diffuser section, 17 ... First evaporator, 17a ... Blower fan,
18 ... 2nd evaporator, 18a ... 1st evaporation part, 18b ... 2nd evaporation part, 20 ... Air-conditioning control apparatus,
30 ... Variable ejector, 30e ... Needle valve, 30f ... Electric actuator,
31, 32... Variable aperture mechanism, 100.

Claims (11)

A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating the high-temperature and high-pressure refrigerant discharged from the compressor (11);
The radiator (12) sucks the refrigerant from the refrigerant suction port (15b, 30b) by the flow of the high-speed jet refrigerant jetted from the nozzle portions (15a, 30a) for decompressing and expanding the downstream refrigerant, and An ejector (15, 30) for increasing the pressure of the refrigerant mixed with the suction refrigerant sucked from the refrigerant suction port (15b, 30b) at the diffuser section (15d, 30d);
A branch part (100) for branching the flow of the refrigerant flowing out of the diffuser part (15d, 30d);
A first evaporator (17) that evaporates one refrigerant branched by the branch part (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, 30b);
The branch part (100) branches the refrigerant so that the dynamic pressure of the refrigerant flowing out of the diffuser part (15d, 30d) is maintained,
The second evaporator (18) is connected to a range in which the dynamic pressure acts,
The ejector refrigeration cycle, wherein a pressure loss (ΔP1drop) in the first evaporator (17) is equal to or less than a pressure increase amount (ΔPej) in the diffuser section (15d, 30d). The flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d), the pressure increase amount (ΔPej), the flow rate of discharged refrigerant (Gn) discharged from the compressor (11), the refrigerant of the first evaporator (17) The relationship among the passage length (Le) of the refrigerant passage from the inlet to the refrigerant outlet, the passage diameter (De) of the refrigerant passage, and the pipe friction coefficient (λ) of the refrigerant passage,
Vdif ≦ 2π × ΔPej / Gn × (De / 2) ² × (De / Le) / λ
The ejector refrigeration cycle according to claim 1, wherein: A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating the high-temperature and high-pressure refrigerant discharged from the compressor (11);
The radiator (12) sucks the refrigerant from the refrigerant suction port (15b, 30b) by the flow of the high-speed jet refrigerant jetted from the nozzle portions (15a, 30a) for decompressing and expanding the downstream refrigerant, and An ejector (15, 30) for increasing the pressure of the refrigerant mixed with the suction refrigerant sucked from the refrigerant suction port (15b, 30b) at the diffuser section (15d, 30d);
A branch section (100) for branching the flow of refrigerant flowing out of the ejector (15, 30);
A first evaporator (17) that evaporates one refrigerant branched by the branch part (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, 30b);
A total refrigeration capacity (Qre) which is a total value of the refrigeration capacity (Q1) exhibited by the first evaporator (17) and the refrigeration capacity (Q2) exhibited by the second evaporator (18); An ejector-type refrigeration cycle characterized in that the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) is adjusted based on the dryness (Xmix) of the mixed refrigerant. Refrigerant flow rate adjusting means (11a, 12a, 17a, 30e, 30f, 31, 32) for adjusting the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d);
Refrigerant flow rate control means (20) for controlling the operation of the refrigerant flow rate adjustment means (11a, 12a, 17a, 30e, 30f, 31, 32),
The refrigerant flow rate control means (20) has a predetermined flow rate (Vdif) of the refrigerant flowing out of the diffuser section (15d, 30d) based on the total refrigeration capacity (Qre) and the dryness of the mixed refrigerant (Xmix). The ejector refrigeration cycle according to claim 3, wherein the operation of the refrigerant flow rate adjusting means (11a, 12a, 17a, 30e, 30f, 31, 32) is controlled so as to be within a range. The refrigerant flow rate control means (20) includes a refrigerant enthalpy increase (ΔHcomp) increased in the compressor (11), a refrigerant enthalpy decrease (ΔHcond) decreased in the radiator (12), and the first evaporation. The total refrigerating capacity (Qre) obtained by the enthalpy (H1out) of the outlet refrigerant (17) and the enthalpy (H2out) of the second evaporator (18) outlet side refrigerant, and the dryness (Xnoz) of the injected refrigerant The diffuser portions (15d, 30d) are calculated based on the dryness (Xmix) of the mixed refrigerant obtained by the enthalpy (Hdif) of the refrigerant flowing out and the enthalpy (Hmix) of the mixed refrigerant. ) The refrigerant flow so that the flow rate (Vdif) of the outflow refrigerant falls within a predetermined range. The ejector refrigeration cycle according to claim 4, wherein the operation of the speed adjusting means (11a, 12a, 17a, 30e, 30f, 31, 32) is controlled. The refrigerant flow rate adjusting means is composed of discharge capacity changing means (11a) for changing the refrigerant discharge capacity of the compressor (11),
The refrigerant flow rate control means (20) adjusts the flow speed (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the discharge capacity changing means (11a). Item 6. The ejector refrigeration cycle according to any one of Items 3 to 5. The refrigerant flow rate adjusting means is composed of a heat dissipation capacity changing means (12a) for changing the refrigerant heat dissipation capacity of the radiator (12).
The refrigerant flow rate control means (20) adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the heat dissipation capacity changing means (12a). Item 6. The ejector refrigeration cycle according to any one of Items 3 to 5. The refrigerant flow rate adjusting means is composed of evaporation capacity changing means (17a) for changing the refrigerant evaporation capacity of at least one of the first evaporator (17) and the second evaporator (18),
The refrigerant flow rate control means (20) adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the evaporation capacity changing means (17a). Item 6. The ejector refrigeration cycle according to any one of Items 3 to 5. The refrigerant flow rate adjusting means includes throttle opening changing means (30e, 30f) for changing the throttle passage area of the nozzle portion (30a),
The refrigerant flow rate control means (20) adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (30d) by controlling the operation of the throttle opening change means (30e, 30f). The ejector type refrigeration cycle according to any one of claims 3 to 5. The refrigerant flow rate adjusting means is composed of a high-pressure side throttle mechanism (31) disposed downstream of the radiator (12) and upstream of the ejector (15, 30),
The refrigerant flow rate control means (20) adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the high-pressure side throttle mechanism (31). Item 6. The ejector refrigeration cycle according to any one of Items 3 to 5. The second evaporator (18) is disposed between a plurality of evaporators (18a, 18b) connected in series and the low-pressure side throttle for decompressing and expanding the refrigerant by the plurality of evaporators (18a, 18b). Having a mechanism (32);
The refrigerant flow rate control means (20) adjusts the flow rate (Vdif) of the refrigerant flowing out of the diffuser part (15d, 30d) by controlling the operation of the low pressure side throttle mechanism (32). Item 6. The ejector refrigeration cycle according to any one of Items 3 to 5.

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