US20090229306A1

US20090229306A1 - Vapor compression refrigerating cycle apparatus

Info

Publication number: US20090229306A1
Application number: US12/380,792
Authority: US
Inventors: Etsuhisa Yamada; Haruyuki Nishijima; Gouta Ogata; Mika Gocho; Hideya Matsui
Original assignee: Denso Corp
Current assignee: Denso Corp
Priority date: 2008-03-13
Filing date: 2009-03-04
Publication date: 2009-09-17
Also published as: CN101532740A; CN101532740B; DE102009012360A1; JP2009222255A

Abstract

A vapor compression refrigerating cycle apparatus includes a compressor, a radiator, a first decompressing device, a second decompressing device, a flow distributor, an ejector, and a suction-side evaporator. The vapor compression refrigerating cycle apparatus is configured such that refrigerant pressure (P0) at an inlet of the first decompressing device, refrigerant pressure (P) at an inlet of a nozzle portion of the ejector, refrigerant pressure (P2) at an outlet of the nozzle portion satisfy a pressure relationship of 0.1×(P0−P2)≦(P0−P)≦0.6×(P0−P2). Alternative to or in addition to the pressure relationship, the vapor compression refrigerating cycle apparatus is configured such that a dryness of refrigerant at the inlet of the nozzle portion is in a range between 0.003 and 0.14.

Description

CROSS REFERENCE TO RELATED APPLICATION

This application is based on Japanese Patent Application No. 2008-64665 filed on Mar. 13, 2008, the disclosure of which is incorporated herein by reference.

FIELD OF THE INVENTION

The present invention relates to a vapor compression refrigerating cycle apparatus having an ejector as a refrigerant decompressing and circulating device.

BACKGROUND OF THE INVENTION

A vapor compression refrigerating cycle apparatus is, for example, described in JP-A-2007-23966 (US2006/0266072 A1). The described refrigerating cycle apparatus has an ejector as a decompressing device for decompressing condensed refrigerant and two evaporators. The ejector generally has a nozzle portion, a suction portion, a mixing portion and a pressure-increase portion.
The nozzle portion draws a part of the refrigerant downstream of a radiator, and decompresses and expands the drawn refrigerant in an isenthalpic manner. The suction portion draws a remaining part of the refrigerant from one of the evaporators. The part of the refrigerant is jetted from the nozzle portion at high velocity, and is mixed with the remaining part of the refrigerant drawn from the suction portion. Further, the mixed refrigerant is increased in pressure through the pressure-increase portion, and is then discharged from the ejector. The refrigerant is further conducted to the other evaporator to be evaporated, and is then drawn into the compressor.

SUMMARY OF THE INVENTION

In such a vapor compression refrigerating cycle apparatus, in a case where refrigerant drawn into a nozzle portion of an ejector is in a gas and liquid two-phase condition, it is difficult to improve ejector efficiency while appropriately controlling the flow rate of the refrigerant at the nozzle portion. As such, it is difficult to stably maintain a coefficient of performance (COP) of the refrigerating cycle apparatus at a sufficient level.
The present invention is made in view of the foregoing matter, and it is an object of the present invention to provide a vapor compression refrigerating cycle apparatus capable of controlling a condition of refrigerant at the nozzle portion of the ejector to a predetermined condition, thereby to maintain the COP at the sufficient level.
According to a first aspect of the present invention, a vapor compression refrigerating cycle apparatus includes a compressor, a radiator, first and second decompressing devices, a flow distributor, an ejector and a suction-side evaporator. The compressor draws and compresses refrigerant. The radiator radiates heat of refrigerant discharged from the compressor. The first decompressing device decompresses refrigerant discharged from the radiator. The flow distributor separates refrigerant decompressed by the first decompressing device into at least a first flow and a second flow. The ejector includes a nozzle portion and a suction portion. The nozzle portion draws refrigerant of the first flow, and decompresses and expands the refrigerant of the first flow to generate a refrigerant jet flow. The suction portion draws refrigerant of the second flow by the refrigerant jet flow from the nozzle portion. The second decompressing device decompresses the refrigerant of the second flow. The suction-side evaporator evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector. Further, the vapor compression refrigerating cycle apparatus is configured such that refrigerant pressure (P0) at an inlet of the first decompressing device, refrigerant pressure (P) at an inlet of the nozzle portion, refrigerant pressure (P2) at an outlet of the nozzle portion satisfy a pressure relationship of 0.1×(P0−P2)>(P0−P)≦0.6×(P0−P2):
Accordingly, because the refrigerant pressure at the inlet of the nozzle portion becomes an optimum condition, a distribution ratio of the refrigerant to the suction-side evaporator and the nozzle portion can be set to an optimum ratio. Therefore, capacity of the suction-side evaporator and nozzle efficiency are both improved. As such, the COP of the vapor compression refrigerating cycle apparatus improves. For example, the pressure relationship is achieved by adjusting a throttle degree of at least one of the first decompressing device, the second decompressing device and the nozzle portion.
According to a second aspect of the present invention, a vapor compression refrigerating cycle apparatus includes a compressor, a radiator, first and second decompressing devices, a flow distributor, an ejector and a suction-side evaporator. The compressor draws and compresses refrigerant. The radiator radiates heat of refrigerant discharged from the compressor. The first decompressing device decompresses refrigerant discharged from the radiator. The flow distributor separates refrigerant decompressed by the first decompressing device into at least a first flow and a second flow. The ejector includes a nozzle portion and a suction portion. The nozzle portion draws refrigerant of the first flow, and decompresses and expands the refrigerant of the first flow to generate a refrigerant jet flow. The suction portion draws refrigerant of the second flow by the refrigerant jet flow from the nozzle portion. The second decompressing device decompresses the refrigerant of the second flow. The suction-side evaporator evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector. Further, the vapor compression refrigerating cycle apparatus is configured such that the refrigerant at an inlet of the nozzle portion has a dryness in a range between 0.003 and 0.14.
Accordingly, because the dryness of the refrigerant at the inlet of the nozzle portion is controlled to an optimum condition, a distribution ratio of the refrigerant to the suction-side evaporator and the nozzle portion can be set to an optimum ratio. Therefore, capacity of the suction-side evaporator and nozzle efficiency are both improved. As such, the COP of the vapor compression refrigerating cycle apparatus improves.

BRIEF DESCRIPTION OF THE DRAWINGS

Other objects, features and advantages of the present invention will become more apparent from the following detailed description made with reference to the accompanying drawings, in which like parts are designated by like reference numbers and in which:

FIG. 1 is a schematic block diagram of a vapor compression refrigerating cycle apparatus according to an embodiment of the present invention;

FIG. 2 is a graph showing a relationship between enthalpy and pressure in the vapor compression refrigerating cycle apparatus according to the embodiment;

FIG. 3 is a graph showing an operation of the vapor compression refrigerating cycle apparatus according to the embodiment;

FIG. 4 is a graph showing a relationship between refrigerant pressure and a COP improvement effect of the vapor compression refrigerating cycle apparatus according to the embodiment;

FIG. 5 is a graph showing a relationship between dryness of refrigerant at an inlet of a nozzle portion of an ejector and the COP improvement effect of the vapor compression refrigerating cycle apparatus according to the embodiment;

FIG. 6 is a schematic block diagram of a vapor compression refrigerating cycle apparatus according to another embodiment of the present invention; and

FIG. 7 is a graph showing a relationship between enthalpy and pressure in the vapor compression refrigerating cycle apparatus shown in FIG. 6.

DETAILED DESCRIPTION OF EXEMPLARY EMBODIMENTS

An exemplary embodiment of the present invention will now be described with reference to FIGS. 1 to 5. FIG. 1 shows an example of a vapor compression refrigerating cycle apparatus 10 of the present embodiment. The refrigerating cycle apparatus 10 is an ejector-type refrigerating cycle apparatus including an ejector 5, which serves as a decompressing device for decompressing refrigerant and a pump for transporting the refrigerant. The refrigerating cycle apparatus 10 is, for example, employed in a vehicle refrigerating unit, a vehicle air conditioner and the like.
The refrigerating cycle apparatus 10 generally includes a compressor 1, a radiator 2, a first decompressing device 3, an ejector 5, a flow distributor 6, a second decompressing device 4 and a suction-side evaporator 8. In the example shown in FIG. 1, the refrigerating cycle apparatus 10 further includes a discharge-side evaporator 7. The compressor 1, the radiator 2, the first decompressing device 3, the ejector 5, the flow distributor 6 and the discharge-side evaporator 7 (hereinafter, referred to as the first evaporator 7) are connected in a form of loop through pipes.
The flow distributor 6 distributes the refrigerant, which has been decompressed through the first decompressing device 3 into a first flow that is in communication with a nozzle portion 5 a of the ejector 5 and a second flow that is in communication with a suction portion 5 b of the ejector 5 through a branch passage 9. That is, the branch passage 9 diverges from the flow distributor 6 and connects to the suction portion 5 b of the ejector 5. The second decompressing device 4 and the suction-side evaporator 8 (hereinafter, referred to as the second evaporator 8) are disposed on the branch passage 9.
The compressor 1 draws and compresses refrigerant. The compressor 1 discharges high pressure refrigerant toward the radiator 2. The compressor 1 is driven by a vehicle engine through an electromagnetic clutch, a pulley, and a belt. The compressor 1 is any-types of compressor, such as, a variable capacity-type compressor that is capable of adjusting a discharge rate in accordance with a change in discharge capacity, a fixed capacity-type compressor that is capable of adjusting a discharge rate in accordance with a change in a rate of operation thereof by on and off operations of the electromagnetic clutch, an electric compressor that is capable of adjusting a discharge rate by controlling a rotation speed of an electric motor, or the like.
The radiator 2 is disposed downstream of the compressor 1 with respect to a flow of refrigerant. The radiator 2 performs heat exchange between the high pressure refrigerant discharged from the compressor 1 and air, thereby to condense the refrigerant. The air is, for example, outside air drawn from an outside of a passenger compartment of a vehicle and forcibly applied to the radiator 2 such as by a blower (not shown).
Here, the refrigerant is not limited to a specific refrigerant. In the present embodiment, for example, the refrigerant is R404A. In a case where a chlorofluorocarbon-base refrigerant, such as R404A, is used, the refrigerating cycle apparatus is operated under a subcritical condition where pressure on a high-pressure side does not exceed the critical pressure.
In this case, therefore, the radiator 2 serves as a condenser for condensing the refrigerant therein. In a case where carbon dioxide is used as the refrigerant, the refrigerating cycle apparatus is operated under a supercritical condition where pressure on the high-pressure side exceeds the critical pressure. In this case, the refrigerant radiates heat while maintaining in a supercritical condition, and thus is not condensed.
The first decompressing device 3 serves to decompress the high pressure refrigerant having passed through the radiator 2. The first decompressing device 3 is, for example, an expansion valve.
The expansion valve 3 is, for example, a temperature operation-type in which a valve opening degree is controlled to adjust a superheat degree to a predetermined condition based on a temperature of refrigerant at an outlet of the first evaporator 7.
Alternatively, the first decompressing device 3 can be a fixed flow control valve, an electric controlled flow control valve in which a refrigerant flow rate is variably controlled, or the like.
The high pressure refrigerant is decompressed into a gas and liquid two-phase condition by controlling a decompressing rate through the first decompressing device 3, and is then conducted to the flow distributor 6. Here, the gas and liquid two-phase refrigerant forms stratified flow, linear flow, slag flow, and the like in accordance with dryness, velocity and the like. Further, the gas and liquid two-phase refrigerant forms an upper and lower separated flow in which gas-phase refrigerant is located above liquid refrigerant.
The flow distributor 6 is a block member having such as a cubic shape and a rectangular parallelepiped shape. The flow distributor 6 is formed with multiple passages therein, and serves to distribute the refrigerant decompressed through the first decompressing device 3 into at least two flows at predetermined rates.
The flow distributor 6 at least has a first passage that is in communication with the first decompressing device 3, a second passage diverging from the first passage and connecting to the branch passage 9 for conducting the refrigerant toward the second evaporator 8, and a third passage diverging from the first passage and is in communication with the nozzle portion 5 a of the ejector 5. The first to third passages constitute a distribution rate adjusting part.
Each of the first to third passages has a predetermined shape and a passage area (cross-sectional area) and is located at a predetermined position, such as a predetermined height. For example, the passage areas of the first to third passages satisfy a predetermined relationship. Therefore, the flow rate of refrigerant passing through each passage, the volume of liquid-phase refrigerant passing through each passage and the like are determined in accordance with pressure condition of the refrigerant. Further, the low distributor 6 can be provided with a valve device to vary the flow rates of the refrigerant passing through the respective passages.
The ejector 5 serves as a decompressing device for decompressing refrigerant and a circulating device for circulating the refrigerant by means of a drawing effect (dragging effect) generated by a jet flow of refrigerant. The ejector 5 generally has the nozzle portion 5 a, the suction portion 5 b, a mixing portion 5 c and a diffuser portion 5 d.
The nozzle portion 5 a is in communication with the third passage of the flow distributor 6. The nozzle portion 5 a draws the refrigerant of the first flow from the flow distributor 6, and decompresses and expands the refrigerant in an isenthalpic manner by reducing a passage area therein. The suction portion 5 b is disposed to be in communication with a jet port of the nozzle portion 5 a. The suction portion 5 b draws gas-phase refrigerant from the second evaporator 8.
The mixing portion 5 c mixes the refrigerant jetted from the jet port of the nozzle portion 5 a at high velocity with the refrigerant drawn from the suction portion 5 b. The diffuser portion 5 d is disposed downstream of the mixing portion 5 c. The diffuser portion 5 d is configured such that a passage area gradually reduces to reduce the velocity of the refrigerant and increase the refrigerant in pressure. That is, the diffuser portion 5 d has a function of converting velocity energy of the refrigerant into pressure energy. Therefore, the diffuser portion 5 d can be also referred to as a pressure-increase portion.
Accordingly, in the ejector 5, pressure is rapidly reduced in the nozzle portion 5 a, and is the lowest at the outlet of the nozzle portion 5 a. Since the refrigerant decompressed in the nozzle portion 5 a is mixed with the refrigerant drawn from the suction portion 5 b in the mixing portion 5 c, the pressure gradually increases. The pressure is then increased in the diffuser portion 5 d due to the decrease in velocity.
The first evaporator 7 is disposed downstream of the diffuser portion 5 d with respect to the flow of refrigerant. The first evaporator 7 is a heat absorber that performs heat exchange between the refrigerant discharged from the ejector 5 and air, which is forcibly applied to the first evaporator 7, thereby to achieve a heat absorbing effect due to evaporation of the refrigerant. A discharge side of the first evaporator 7 is in communication with a suction side of the compressor 1.
The second decompressing device 4 is, for example, constructed of a capillary tube, such as a spiral tubule. The second decompressing device 4 is disposed on the branch passage 9. The second decompressing device 4 serves to decompress refrigerant flowing in the second evaporator 8 and control a flow rate of the refrigerant. The second decompressing device 4 can be a variable decompressing device such as an electric control expansion valve, in place of the capillary tube.
The second evaporator 8 is disposed on the branch passage 9 downstream of the second decompressing device 4 with respect to the flow of refrigerant. The second evaporator 8 is a heat absorber, similar to the first evaporator 7. That is, the second evaporator 8 achieves a heat absorbing effect by evaporating the refrigerant.
For example, the second evaporator 8 is located downstream of the first evaporator 7 with respect to the flow of air. Thus, the air having passed through the first evaporator 7 is further cooled while passing through the second evaporator 8 by exchanging heat with the refrigerant flowing inside of the second evaporator 8. Then, the air is conducted to a predetermined space, such as for an air conditioning operation.
Alternatively, the first evaporator 7 and the second evaporator 8 are provided differently. For example, airs can be applied separately to the first evaporator 7 and the second evaporator 8 by blowers and the like, and the airs can be conducted to different spaces to be air-conditioned.
The first evaporator 7 and the second evaporator 8 can be constructed separately from each other. Alternatively, the first evaporator 7 and the second evaporator 8 can be integrated with each other. In a case where the first evaporator 7 and the second evaporator 8 are integrated with each other, the first evaporator 7 and the second evaporator 8 can be joined with each other by brazing. In this case, components of the first evaporator 7 and the second evaporator 8 are made of aluminum, for example. Further, the flow distributor 6, the second decompressing device 4 and the ejector 5 can be integrated with each other into a unit, and further fixed to the first and second evaporators 7, 8.
The vapor compression refrigerating cycle apparatus 10 can be further provided with an internal heat exchanger to perform heat exchange between the high pressure refrigerant flowing between the radiator 2 and the first decompressing device 3 and low pressure refrigerant to be drawn to the compressor 1. In this case, the high pressure refrigerant flowing between the radiator 2 and the expansion valve 3 is cooled by the heat exchange with the low pressure refrigerant. As such, enthalpy differential between refrigerant inlets and refrigerant outlets of the first evaporator 7 and the second evaporator 8 increases, and thus cooling capacity improves.
For example, an operation of the compressor 1 is controlled by a control unit (not shown). The control unit is constructed of a microcomputer including a CPU, a ROM, a RAM and the like and peripheral circuits. The control unit executes various computations and processing in accordance with control programs stored in the ROM to control operations of various devices including the compressor 1.
The control unit receives detection signals from various sensors and various manipulation signals from an operation panel (not shown). For example, the operation panel is provided with a temperature setting switch for setting a cooling temperature of a space to be cooled and an air conditioner operation switch for generating an operation command signal of the compressor 1.
Next, an operation of the vapor compression refrigerating cycle apparatus 10 will be described with reference to FIG. 2. In FIG. 2, points a1 through i1 correspond to locations a1 through i1 in FIG. 1.
When the electromagnetic clutch of the compressor 1 is electrically conducted in accordance with the signal generated from the control unit, the electromagnetic clutch becomes in a connected state and a driving force is transmitted from an engine of a vehicle to the compressor 1. When the operation of the compressor 1 is started, the gas-phase refrigerant is drawn into the compressor 1 from the first evaporator 7 and compressed in the compressor 1. The high temperature, high pressure refrigerant at a flow rate G (=Gn+Ge) is discharged from the compressor 1 toward the radiator 2. (g1→a1)
In the radiator 2, the high temperature, high pressure refrigerant is condensed by being cooled by the air. (a1→b1)
High pressure liquid-phase refrigerant flowing out from the radiator 2 at the flow rate G is decompressed and expanded into predetermined pressure by the first decompressing device 3. Thus, the gas and liquid two-phase refrigerant is generated. Here, refrigerant pressure at an inlet of the first decompressing device 3 is defined as P0.
The gas and liquid two-phase refrigerant flowing out from the first decompressing device 3 flows in the flow distributor 6. In the flow distributor 6, the gas and liquid two-phase refrigerant is separated into the first flow passing through the third passage toward the nozzle portion 5 a of the ejector 5 (b1→c1) and the second flow passing through the second passage toward the second decompressing device 4 (b1→h1), at predetermined flow rates. Here, the flow rate of the refrigerant of the first flow is defined as Gn, and the flow rate of the refrigerant of the second flow is defined as Ge. Refrigerant pressure at an inlet of the nozzle portion 5 a is defined as P.
The refrigerant flows in the nozzle portion 5 a of the ejector 5 at the flow rate Gn from the first flow. In the ejector 5, the refrigerant is decompressed and expanded in the isenthalpic manner through the nozzle portion 5 a. (c1→d1). Thus, the refrigerant pressure P reduces to refrigerant pressure P2 at the outlet of the nozzle portion 5 a. That is, in the nozzle portion 5 a, pressure energy of the refrigerant is converted into velocity energy, and thus the refrigerant is jetted from the jet port of the nozzle portion 5 a at high velocity. At this time, the gas-phase refrigerant of the flow rate Ge is drawn from the second evaporator 8 into the suction portion 5 b by the drawing effect generated by the jet flow of the refrigerant.
The refrigerant jetted from the nozzle portion 5 a and the refrigerant drawn into the suction portion 5 b are mixed with each other in the mixing portion 5 c (d1→e1, i1→e1), and then introduced in the diffuser portion 5 d. In the diffuser portion 5 d, since the passage area is gradually increased, velocity (expansion) energy of the refrigerant is converted into pressure energy. Thus, the refrigerant is increased in pressure (e1→f1).
The refrigerant flowing out from the diffuser portion 5 d at the flow rate G flows in the first evaporator 7. In the first evaporator 7, the low temperature, low pressure refrigerant is evaporated in a heat exchanging core portion by absorbing heat from the air (f1→g1). Pressure of the low temperature, low pressure refrigerant is defined as P1. The gas-phase refrigerant evaporated in the first evaporator 1 is drawn by the compressor 1 and is compressed again.
On the other hand, the refrigerant of the second flow is conducted in the branch passage 9 at the flow rate Ge and decompressed into the low pressure refrigerant by the second decompressing device 4 (b1→h1). The low pressure refrigerant is then conducted to the second evaporator 8. In the second evaporator 8, the low pressure refrigerant is evaporated by absorbing heat from the air (h1→i1), and becomes the gas-phase refrigerant. The gas-phase refrigerant is drawn into the suction portion 5 b at the flow rate Ge.
Accordingly, the refrigerant of the flow rate Gn is supplied to the first evaporator 7 and the refrigerant of the flow rate Ge is supplied to the second evaporator 8 through the second decompressing device 4. Therefore, cooling effects are achieved simultaneously by the first and second evaporators 7, 8.
In the present embodiment, the first decompressing device 3, the second decompressing device 4 and the nozzle portion 5 a have predetermined throttle degrees such that the refrigerant pressure P0 at the inlet of the first decompressing device 3, the refrigerant pressure P at the inlet of the nozzle portion 5 a and the refrigerant pressure P2 at the outlet of the nozzle portion 5 a satisfy the following pressure relationship (R1):
0.1×(P0−P2)≦P≦0.6×(P0−P2) (R1)
That is, the vapor compression refrigerating cycle apparatus 10 is configured such that a decrease in pressure, that is, a differential pressure between the refrigerant pressure P0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to a value that is obtained by multiplying a differential pressure between the inlet of the first decompressing device 3 and the outlet of the nozzle portion 5 a by a value that is at least 0.1 and at most 0.6.
In FIG. 2, _ΔP represents an increase in pressure by the ejector 5, such as by the diffuser portion 5 d. That is, _ΔP is a differential pressure (P1−P2) between the refrigerant pressure P1 flowing in the first evaporator 7 and a refrigerant evaporation pressure P2 in the second evaporator 8. Because suction pressure of the compressor 1 is increased by an effect of increasing in pressure by the diffuser portion 5 d, which is represented by _ΔP, the driving force of the compressor 1 can be reduced. As a result, the COP of the vapor compression refrigerating cycle apparatus 10 improves.
As shown in FIG. 2, the refrigerant evaporation pressure P2 of the second evaporator 8 is lower than the refrigerant evaporation pressure P1 of the first evaporator 7. Therefore, a refrigerant evaporation temperature of the second evaporator 8 is lower than a refrigerant evaporation temperature of the first evaporator 7.
In the case where the first evaporator 7 is disposed upstream of the second evaporator 8 with respect to the flow of air, it is possible to ensure both a temperature differential between the refrigerant evaporation temperature of the first evaporator 7 and the air and a temperature differential between the refrigerant evaporation temperature of the second evaporator 8 and the air. Accordingly, cooling performances of both the first and second evaporators 7, 8 effectively improve.
FIG. 3 shows relationships between differential pressure at inlets and outlets of flow rate control devices, such as the first decompressing device 3, the second decompressing device 4 and the nozzle portion 5 a, and the flow rates at the respective portions.
As shown in FIG. 3, the flow rate G of the first decompressing device 3 increases as the refrigerant pressure P at the inlet of the nozzle portion 5 a reduces, that is, as the differential pressure (P0−P) between the refrigerant pressure P0 of the inlet of the first decompressing device 3 and the refrigerant pressure P of the inlet of the nozzle portion 5 a increases. In this case, the differential pressure between the inlet and the outlet of each of the nozzle portion 5 a and the second decompressing device 4 reduces. As such, each of the flow rates Gn, Ge reduces. Further, the refrigerant pressure P at the inlet of the nozzle portion 5 a is determined to pressure where the flow rate G of the first decompressing device 3 is equal to the sum of the flow rate Gn of the nozzle portion 5 a and the flow rate Ge of the second decompressing device 4.
Further, a ratio of the flow rates Gn, Ge is determined based on a flow rate property by the differential pressure between the inlet and the outlet of the nozzle portion 5 a and a flow rate property by the differential pressure between the inlet and the outlet of the second decompressing device 4. Further, expansion energy recovered at the nozzle portion 5 a reduces as the refrigerant pressure P at the inlet of the nozzle portion 5 a reduces. As such, the increase in pressure _ΔP by the ejector 5 reduces.
Accordingly, in view of ensuring the performance of the evaporators 7, 8 and nozzle efficiency, it is preferable to set the ratio of the flow rates Gn, Ge to an optimum ratio as discussed hereinabove, and it is recognized that there is an optimum condition of the refrigerant pressure at the inlet of the nozzle portion 5 a. Further, it is realized that the nozzle efficiency is sufficient when the pressure relationship (R1) is satisfied because the pressure condition at the inlet of the nozzle portion 5 a is under the optimum condition. Moreover, it is realized that the refrigerating capacity (COP) is sufficiently achieved in a range of the refrigerant flow rate ratio, which is obtained when the pressure relationship (R1) is satisfied. The range of the refrigerant flow rate ratio corresponds to a nondimensional flow rate ratio (Ge/(Ge+Gn)).
FIG. 4 shows a relationship between a pressure ratio (P0−P)/(P0−P2) and a COP improvement effect. The pressure ratio (P0−P)/(P0−P2) is a ratio of the decrease in the refrigerant pressure P at the inlet of the nozzle portion 5 a with respect to the refrigerant pressure P0 at the inlet of the first decompressing device 3 to the decrease in refrigerant pressure P2 at the outlet of the nozzle portion 5 a with respect to the refrigerant pressure P0 at the inlet of the first decompressing device 3.
Here, the COP improvement effect is the improvement of the COP of the vapor compression refrigerating cycle apparatus 10 with respect to the COP of an expansion valve cycle apparatus. That is, the higher the value indicative of the COP improvement effect is, the more the COP of the vapor compression refrigerating cycle apparatus 10 is improved, as compared with the COP of the expansion valve cycle apparatus. The expansion valve cycle apparatus is a refrigerating cycle apparatus constructed by sequentially connecting the compressor, the radiator, the expansion valve and the evaporator in a form of closed circuit.
According to the graph of FIG. 4, the COP improvement effect is low in regions where the pressure ratio (P0−P)/(P0−P2) is small and large. Further, the COP improvement effect is high in a middle region between the regions. Particularly, when the pressure ratio (P0−P)/(P0−P2) is in a range between 0.1 and 0.6, the COP improvement effect is stable and at the highest level. That is, the pressure ratio (P0−P)/(P0−P2) is the optimum in the range between 0. and 0.6.
This is based on the following reasons. Since the refrigerant evaporation temperature of the second evaporator 8 is lower than the refrigerant evaporation temperature of the first evaporator 7, refrigerating capacity Qer of the entirety of the refrigerating cycle apparatus is increased by increasing the flow rate Ge of the refrigerant passing through the second evaporator 8. Thus, the COP improves. However, the flow rate Gn of the refrigerant passing through the nozzle portion 5 a reduces with an increase in the flow rate Ge. As a result, the increase in pressure _ΔP by the ejector 5 reduces.
Accordingly, when the flow rate Ge is excessively increased, the driving force L of the compressor 1 is excessively increased. As a result, the COP (Qer/L), which is obtained by a ratio of the refrigerating capacity Qer of the entirety of the refrigerating cycle apparatus to the driving force L of the compressor 1, reduces.
According to FIGS. 3 and 4, it is found that when the pressure relationship (R1) is satisfied, the COP of the refrigerating cycle apparatus 10 is sufficiently improved, as compared with the COP of the expansion valve cycle. Accordingly, the COP is ensured at a sufficient level.
For example, the pressure relationship (R1) is achieved by constructing the first decompressing device 3, the second decompressing device 4 and the ejector 5 to have the predetermined throttle degrees, respectively.
When the pressure relationship (R1) is satisfied, the refrigerant at the inlet of the nozzle portion 5 a is controlled under a predetermined pressure condition. Accordingly, the COP is sufficiently ensured.
FIG. 5 is a graph showing a relationship between dryness X of the refrigerant at the inlet of the nozzle portion 5 a and the COP improvement effect of the vapor compression refrigerating cycle apparatus 10.
The dryness X is a ratio of vapor in 1 kg wet vapor of the refrigerant at the inlet of the nozzle portion 5 a. That is, the dryness X means that refrigerant contains X kg of dry saturated vapor and (1−X) kg of saturated liquid. Here, the COP improvement effect means the improvement of the COP of the vapor compression refrigerating cycle apparatus 10 with respect to the COP of the expansion valve cycle apparatus, similar to FIG. 4. That is, the higher the value of the COP improvement effect is, the more the COP of the vapor compression refrigerating cycle apparatus 10 is improved, as compared with the COP of the expansion valve cycle apparatus.
According to FIG. 5, the COP improvement effect is low in regions where the dryness X is small and large. The COP improvement effect is high in the middle region. Particularly, in a region where the dryness X is at least 0.003 and at most 0.14, the COP improvement effect is stable and at the maximum level. That is, the dryness X is the optimum in the range between 0.003 and 0.14. Further, it is realized that the nozzle efficiency is sufficiently ensured when the dryness X is in the range between 0.003 and 0.14, similar to FIG. 3. In this case, however, the nozzle efficiency has a peak on a side adjacent to 0.003.
Therefore, in a case where the refrigerant at the inlet of the nozzle portion 5 a has the dryness X in the range between 0.003 and 0.14, the refrigerant pressure at the inlet of the nozzle portion 5 a can be maintained to an optimum condition, similar to FIG. 3, in accordance with the flow rate properties of the nozzle portion 5 a and the second decompressing device 4. Therefore, the refrigerating capacity of the evaporators 7, 8 and the increase in pressure _ΔP by the ejector 5 are ensured in a balanced condition. As such, the COP of the refrigerating cycle apparatus 10 is sufficiently increased, as compared with the expansion valve cycle apparatus.
For example, the dryness X of the refrigerant at the inlet of the nozzle portion 5 a can be controlled in the range between 0.003 and 0.14 by setting the throttle degrees of the second decompressing device 4 and the ejector 5 are to predetermined degrees. That is, by setting the throttle degrees of the second decompressing device 4 and the ejector 5 to the predetermined degrees, the refrigerant at the inlet of the nozzle portion 5 a can be controlled to a predetermined condition, such as equivalent to the condition shown in FIG. 3. Therefore, the COP of the refrigerating cycle apparatus 10 improves.
Accordingly, in an example, the vapor compression refrigerating cycle apparatus 10 is configured such that the differential (P0−P) between the refrigerant pressure P0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to the value that is obtained by multiplying the differential (P0−P2) between the refrigerant pressure P0 and the refrigerant pressure P2 at the outlet of the nozzle portion 5 a by the value that is at least 0.1 and at most 0.6. The above pressure relationship (R1) can be achieved by setting at least one of the throttle degrees of the first decompressing device 3, the second decompressing device 4 and the nozzle portion 5 a to the predetermined degrees, for example.
In this case, since the decrease in pressure at the inlet of the nozzle portion 5 a can be in the optimum condition, the distribution ratio of the refrigerant to the second evaporator 8 and the nozzle portion 5 a becomes an optimum condition. Therefore, the performance of the evaporators 7, 8 and the efficiency of the ejector 5, such as the nozzle efficiency and the ejector efficiency, can be both ensured. Accordingly, the COP of the refrigerating cycle apparatus 10 improves, as compared with the expansion valve cycle apparatus.
Further, the above example can be employed to the vapor compression refrigerating cycle apparatus including at least the compressor 1, the radiator 2, the first decompressing device 3, the flow distributor 6, the ejector 5, the second decompressing device 4, and the suction-side evaporator 8. That is, even in the vapor compression refrigerating cycle apparatus without having the first evaporator 7, it can be configured to have the pressure relationship (R1). Also in this vapor compression refrigerating cycle apparatus, the similar effects are achieved.
As another example, the vapor compression refrigerating cycle apparatus 10 is configured such that the dryness X of the refrigerant at the inlet of the nozzle portion 5 a is in the range between 0.003 and 0.14. The dryness X in the range between 0.003 and 0.14 is achieved by setting at least one of the throttle degrees of the first decompressing device 3, the second decompressing device 4 and the nozzle portion 5 a to the predetermined degree.
In such a case, since the dryness of the refrigerant at the inlet of the nozzle portion 5 a can be in the optimum condition, the distribution ratio of the refrigerant to the second evaporator 8 and the nozzle portion 5 a becomes the optimum ratio. Therefore, the performance of the evaporators 7, 8 and the efficiency of the ejector 5, such as the nozzle efficiency and the ejector efficiency, can be both ensured. Accordingly, the COP of the refrigerating cycle apparatus 10 improves, as compared with the expansion valve cycle apparatus.
Further, the above example can be employed to the vapor compression refrigerating cycle apparatus including at least the compressor 1, the radiator 2, the first decompressing device 3, the flow distributor 6, the ejector 5, the second decompressing device 4, and the suction-side evaporator 8. That is, even in the vapor compression refrigerating cycle apparatus without having the first evaporator 7, it can be configured such that the refrigerant has the dryness X in the above range at the inlet of the nozzle portion 5 a. Also in this vapor compression refrigerating cycle apparatus, the similar effects are achieved.
As further another example, the vapor compression refrigerating cycle apparatus 10 can be configured such that the differential (P0−P) between the refrigerant pressure P0 at the inlet of the first decompressing device 3 and the refrigerant pressure P at the inlet of the nozzle portion 5 a is equal to the value that is obtained by multiplying the differential (P0−P2) between the refrigerant pressure P0 and the refrigerant pressure P2 at the outlet of the nozzle portion 5 a by the value that is at least 0.1 and at most 0.6, and the dryness of the refrigerant at the inlet of the nozzle portion 5 a is in the range between 0.003 and 0.14.
In such a case, the decrease in refrigerant pressure and the dryness of the refrigerant can be in the optimum conditions. Therefore, the vapor compression refrigerating cycle apparatus 10 can be operated while appropriately maintaining the pressure and enthalpy. Accordingly, the performance of the evaporators 7, 8 and the efficiency of the ejector 5 further improves, and the COP further improves.
In the above examples, the dryness X of the refrigerant at the inlet of the nozzle portion 5 a can be adjusted by the distribution rate adjusting means of the flow distributor 6. In this case, the mixing ratio of the liquid-phase refrigerant and the gas-phase refrigerant flowing toward the nozzle portion 5 a is controlled. Therefore, the dryness X can be more precisely adjusted.

Other Embodiments

The vapor compression refrigerating cycle apparatus 10 can be further modified as follows.
It is not always necessary to have the discharge-side evaporator 7. For example, as shown in FIG. 6, the discharge-side evaporator 7 can be eliminated and an internal heat exchanger 70 that performs heat exchange between the high pressure refrigerant discharged from the radiator 2 and the low pressure refrigerant discharged from the ejector 5 can be added.
In this case, as shown in FIG. 7, the enthalpy of the low pressure refrigerant discharged from the ejector 5 can be increased from the point f1 to the point g1, and the enthalpy of the refrigerant flowing into the suction-side evaporator 8 can be decreased from the point b1 to a point b′1. As a result, the capacity of the suction-side evaporator 8 increases. Also in such a case, the refrigerating cycle apparatus can be configured to satisfy one of or both of the pressure relationship (R1) and the above optimum range of the dryness X. Thus, the similar effects can be achieved.
Further, the discharge-side evaporator 7 can be eliminated, and an accumulator as a low pressure-side gas and liquid separator for separating the refrigerant discharged from the ejector 5 into the gas-phase refrigerant and the liquid-phase refrigerant can be added. Also in such a case, the refrigerating cycle apparatus can be configured to satisfy one of or both of the pressure relationship (R1) and the above optimum range of the dryness X. Thus, the similar effects can be achieved.
The vapor compression refrigerating cycle apparatus 10 as discussed hereinabove can be employed to a heat pump cycle such as for a hot water supply apparatus or an interior air conditioner, and mounted in a movable unit such as a vehicle or in a fixed unit fixed at a predetermined location.
The refrigerant is not limited to R404 refrigerant. The refrigerant can be any other types, such as chlorofluorocarbon-base refrigerant, HC-base refrigerant, carbon dioxide refrigerant or the like, which can be used in the supercritical cycle and the subcritical cycle. Even when the refrigerant other than R404 is used, the similar effects can be achieved.
The pressure relationship (R1) can be achieved by various ways, instead by setting the throttle degree of at least one of the first decompressing device 3, the second decompressing device 4, and the fixed nozzle portion 5 a to the predetermined degree. For example, the ejector 5 has a flow rate variable nozzle portion in which the throttle degree of the nozzle portion is variable in accordance with the movement of a valve rod, in place of the fixed nozzle portion 5 a. In such a case, the pressure relationship (R1) can be satisfied by adjusting the throttle degree of the nozzle portion. As another example, the second decompressing device 4 can be constructed of a flow rate control variable-type decompressing device such as an electric control expansion valve, in place of the capillary tube 4. In such a case, the pressure relationship (R1) can be achieved by adjusting the throttle degree of the second decompressing device 4. An operation of the flow rate control variable-type decompressing device is, for example, controlled by the control unit.
The flow distributor 6 is not limited to the block member having the passages therein, but can be constructed of any other types of distributors. For example, the flow distributor 6 can be constructed of a manifold pipe having branched passages.
Additional advantages and modifications will readily occur to those skilled in the art. The invention in its broader term is therefore not limited to the specific details, representative apparatus, and illustrative examples shown and described.

Claims

1. A vapor compression refrigerating cycle apparatus comprising:

a compressor that draws and compresses refrigerant;

a radiator that radiates heat of refrigerant discharged from the compressor;

a first decompressing device that decompresses refrigerant downstream of the radiator;

a flow distributor that separates refrigerant decompressed by the first decompressing device into at least a first flow and a second flow;

an ejector that includes a nozzle portion and a suction portion, the nozzle portion that draws refrigerant of the first flow and decompresses and expands the refrigerant of the first flow to generate a refrigerant jet flow, the suction portion that draws refrigerant of the second flow by the refrigerant jet flow from the nozzle portion;

a second decompressing device that decompresses the refrigerant of the second flow; and

a suction-side evaporator that evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector, wherein it is configured that refrigerant pressure (P0) at an inlet of the first decompressing device, refrigerant pressure (P) at an inlet of the nozzle portion, refrigerant pressure (P2) at an outlet of the nozzle portion satisfy a pressure relationship of 0.1×(P0−P2)≦(P0−P)≦0.6×(P0−P2).

2. The vapor compression refrigerating cycle apparatus according to claim 1, wherein it is configured that the refrigerant at an inlet of the nozzle portion has a dryness in a range between 0.003 and 0.14.

3. The vapor compression refrigerating cycle apparatus according to claim 2, wherein

the flow distributor has a distribution rate adjusting part that adjusts flow rates of the first and second flows, and

the dryness of the refrigerant is adjusted by the distribution rate adjusting part.

4. The vapor compression refrigerating cycle apparatus according to claim 1, wherein

the pressure relationship is achieved by adjusting a throttle degree of at least one of the first decompressing device, the second decompressing device and the nozzle portion of the ejector.

5. The vapor compression refrigerating cycle apparatus according to claim 1, further comprising:

a discharge-side evaporator that evaporates refrigerant discharged from the ejector.

6. The vapor compression refrigerating cycle apparatus according to claim 1, further comprising:

an internal heat exchanger that performs heat exchange between refrigerant discharged from the radiator and refrigerant discharged from the ejector.

7. A vapor compression refrigerating cycle apparatus comprising:

a compressor that draws and compresses refrigerant;

a radiator that radiates heat of refrigerant discharged from the compressor;

a suction-side evaporator that evaporates refrigerant decompressed by the second decompressing device and discharges evaporated refrigerant toward the suction portion of the ejector, wherein

it is configured that the refrigerant at an inlet of the nozzle portion has a dryness in a range between 0.003 and 0.14.

8. The vapor compression refrigerating cycle apparatus according to claim 7, wherein

9. The vapor compression refrigerating cycle apparatus according to claim 7, further comprising:

10. The vapor compression refrigerating cycle apparatus according to claim 7, further comprising: