WO2011099052A1

WO2011099052A1 - Refrigeration system

Info

Publication number: WO2011099052A1
Application number: PCT/JP2010/000795
Authority: WO
Inventors: 鈴木隆; 篠崎　隆; 杉山　直樹
Original assignee: 株式会社Ｅ・Ｔ・Ｌ
Priority date: 2010-02-09
Filing date: 2010-02-09
Publication date: 2011-08-18
Also published as: JP5537573B2; JPWO2011099052A1

Abstract

Provided is a highly-efficient refrigeration system. Specifically, provided is a refrigeration system equipped with the following components which are connected by means of an annular refrigerant pipe in the following order: a compressor; a radiator; a refrigerant liquefying means for liquefying a refrigerant by converting the pressure energy of the refrigerant into velocity energy; an energy conversion means for converting the velocity energy of the refrigerant from the refrigerant liquefying means into potential energy or work energy while maintaining the degree of dryness; and a heat absorber.

Description

Refrigeration system

The present invention relates to a refrigeration system including a refrigerant liquefying unit that converts pressure energy of refrigerant into velocity energy to liquefy the refrigerant.

2. Description of the Related Art Conventionally, a heat conversion apparatus for condensation using a high-temperature / high-pressure refrigerant gas discharged from a compressor of a refrigeration system as a low-temperature refrigerant liquid, an isobaric cooling unit that cools the high-temperature / high-pressure refrigerant gas by an isobaric change, The remaining gas refrigerant partially liquefied in the cooling section is decompressed by the refrigerant acceleration phenomenon and liquefied with a decrease in enthalpy, and the refrigerant that has passed through the depressurization liquefaction section is depressurized and the enthalpy decreased by a refrigerant acceleration phenomenon. There has been proposed a condensing heat conversion device that includes a helical tube that is cooled by cooling (Patent Document 1). Further, the compressor sucks and compresses the refrigerant, the radiator that radiates heat of the high-pressure refrigerant discharged from the compressor, and the pressure energy of the high-pressure refrigerant on the downstream side of the radiator is converted into velocity energy to decompress and expand the refrigerant. And an ejector that sucks the refrigerant, and a branch flow path that introduces and sucks the refrigerant flow branched from between the radiator and the ejector of the refrigerant circulation path that includes the compressor, the radiator, and the ejector and circulates the refrigerant to the ejector. A refrigeration cycle using an ejector that is disposed in a branch flow path to depressurize the refrigerant flow, and an ejector that is disposed downstream of the flow of the refrigerant in the branch flow path and evaporates the refrigerant is proposed. (Patent Document 2). Each of these technologies has a function of converting the pressure energy of the refrigerant into velocity energy to liquefy the refrigerant, and therefore, a high-efficiency refrigeration system can be realized without using a decompression device that generates frictional heat.

Japanese Patent No. 4,411,349 JP 2008-8572 A

However, when the above-described conventional spiral tube or ejector is used, the refrigerant has a large velocity energy, so that frictional heat is generated in the downstream refrigerant pipe, the dryness of the refrigerant is increased, and the amount of heat absorbed at the heat absorbing portion. There is a problem that the efficiency is lowered by that amount.
Accordingly, an object of the present invention is to solve the above-described problems of the prior art and to provide a highly efficient refrigeration system.

The present invention relates to a compressor, a radiator, a refrigerant liquefying means for converting the pressure energy of the refrigerant into speed energy to liquefy the refrigerant, and a position where the speed energy of the refrigerant from the refrigerant liquefaction means is maintained in a dry state. It is characterized in that an energy conversion means for converting energy or work energy and a heat absorber are sequentially connected in an annular manner by a refrigerant pipe.
In this invention, since the refrigerant liquefying means that liquefies the refrigerant by converting the pressure energy of the refrigerant into velocity energy is provided, a highly efficient refrigeration system can be realized without using a decompression device that generates frictional heat. Since the energy conversion means for converting the velocity energy of the refrigerant from the liquefaction means into potential energy or work energy while maintaining the dryness is provided, the velocity energy of the refrigerant is converted into potential energy or work energy, the flow velocity decreases, and the downstream Friction heat is less likely to be generated in the refrigerant pipe, the dryness of the refrigerant is maintained, the amount of heat absorbed in the heat absorbing portion is increased, and the efficiency can be improved accordingly.

In this case, the refrigerant liquefying means may be a spiral tube that liquefies the refrigerant with reduced pressure and enthalpy reduction by utilizing a choke phenomenon due to the acceleration of the refrigerant.
The energy conversion means may include a tube body that contains a resistor, and the velocity energy of the refrigerant may be converted into work energy when the refrigerant exceeds the resistor in the tube body.
The resistor in the tubular body may be spring-biased toward the inlet side of the tubular body.
The energy conversion means may include a tube body containing a rotating body, and the velocity energy of the refrigerant may be converted into work energy when the refrigerant rotates the rotating body in the tube body.
The refrigerant liquefying means may be provided vertically so that the refrigerant flows from the bottom to the top, and the energy conversion means may convert the velocity energy of the refrigerant into potential energy.
In this case, the refrigerant liquefying means may be an ejector for converting the pressure energy of the refrigerant into velocity energy.

In the present invention, since the refrigerant liquefying means that liquefies the refrigerant by converting the pressure energy of the refrigerant into velocity energy is provided, a highly efficient refrigeration system can be realized without using a decompression device that generates frictional heat. Since the energy conversion means for converting the velocity energy of the refrigerant from the liquefaction means into potential energy or work energy while maintaining the dryness is provided, the velocity energy of the refrigerant is converted into potential energy or work energy, the flow velocity decreases, and the downstream Thus, frictional heat is not generated in the refrigerant piping, the dryness of the refrigerant is maintained, the amount of heat absorbed in the heat absorbing portion is increased, and the efficiency can be improved accordingly.

It is a block diagram which shows one embodiment of this invention. 1 is a Ph diagram of a refrigeration system according to an embodiment of the present invention. FIG. a to e are plan views of main components constituting the heat conversion apparatus for condensation. It is a block diagram which shows a tubular body. It is a block diagram which shows another pipe body. A to C are Ph diagrams, respectively. It is a block diagram which shows another embodiment.

Hereinafter, an embodiment of the present invention will be described with reference to the accompanying drawings.
In FIG. 1, the refrigeration system includes a compressor 1, a mini heat exchange device (isostatic cooling unit) 3, a helical tube (decompressed liquefaction unit) 6, a helical thin tube (decompressed cooling unit) 8, a tube body 10, and an evaporator 11. Are connected to each other by

refrigerant pipes

2, 4, 13, suction pipe 12, large and short pipes (expansion part) 5, branch pipe (expansion part) 7, and collecting pipe (expansion part) 9. The refrigeration function is realized by circulating the refrigerant in the direction of the arrow 21. Note that “mini” of the mini heat exchanger 3 or the mini fan 3-1 described later means “small”, and is used to clarify the feature of the present invention in which the condenser can be made smaller than before. Yes. Moreover, in this specification, each

apparatus group

3, 6, 7, 8, 9, and 10 from the mini heat exchange apparatus 3 to the pipe body 10 is collectively called the heat conversion apparatus 30 for condensation below.

Since the compressor 1 and the evaporator 11 are basically the same in structure and function as those used in the current refrigeration system, detailed description thereof is omitted here, and heat of condensation that is a feature of the present embodiment is omitted. The conversion device 30 will be described in detail.

FIG. 2 is a Ph diagram of the refrigeration cycle of the refrigeration system using the condensing heat conversion device 30 according to the present embodiment. A broken line indicates a conventional cycle, and a solid line indicates a cycle according to the present embodiment. In the conventional cycle, adiabatic compression by the compressor (points a to b), condensation due to heat release from the isobaric change by the condenser (points b to c), isoenthalpy changes due to the expansion valve throttling phenomenon (points c to points) d) The cycle is completed by evaporation (point d to point a) due to endothermic heat of isothermal expansion and isothermal expansion by the evaporator.

In the present embodiment, a high-temperature (40 ° C. or higher) and high-pressure (0.6 MPa or higher) gaseous refrigerant is discharged from the compressor 1 (point h to point i), and the mini heat constituting the condensation heat conversion device 30 is discharged. Part of the refrigerant (5 to 50% by weight) is liquefied by the exchange device 3 (point i to point j).
In FIG. 1, the mini heat exchange device 3 is a normal air-cooling type in which a heat dissipation fan is provided in a pipe through which the refrigerant passes. However, the mini heat exchange device 3 is not limited to this type, and may be a water-cooling type or the like. In the condenser of the conventional refrigeration system, almost all of the high temperature / high pressure gas discharged from the compressor is liquefied, but the mini heat exchange device 3 is only required to liquefy a part of the high temperature / high pressure gas. Small size is possible. When compared with a refrigeration system having the same cooling capacity and having the same type of heat exchange device (condenser), the mini heat exchange device of the present embodiment can be about 1/10 that of a conventional condenser. The mini heat exchanging device 3 is provided with a mini fan 3-1, which can be operated when a predetermined operating state is reached, as will be described later, to increase the heat exchanging capacity.

The refrigerant partially liquefied by the mini heat exchange device 3 enters the spiral pipe 6 through the refrigerant pipe 4 and the large and short pipes 5. When viewed in terms of the cross-sectional area of the refrigerant flow path, the large and short tubes 5 once become larger with respect to the mini heat exchange device 3, and the spiral tube 6 becomes smaller than the cross sectional area of the mini heat exchange device 3.

FIG. 3 is a plan view showing the shapes of the large and short tubes 5, the helical tube 6, the branch tube 7, the helical thin tube 8, and the collecting tube 9.
As shown in FIG. 3 (a), the large and short tubes 5 have a cylindrical shape with a central thick portion having a length L1 of 10 to 50 mm and an inner diameter D1 of 8 to 20 mm. Since both ends thereof are connected to the refrigerant pipe 4 and the spiral pipe 6, the shapes thereof are cylindrical shapes that can be connected by inserting the refrigerant pipe 4 and the spiral pipe 6, respectively. The inner diameter D1 of the central thick portion is preferably set larger than the inner diameter of either the refrigerant pipe 4 or the helical pipe 6.
As shown in FIG. 3B, the helical tube 6 has a form in which a thin tube is spirally wound. The inner diameter and the number of windings are determined from various specifications such as the refrigeration capacity of the refrigeration system, but the inner diameter is allowed to be 2 to 150 mm, preferably 2 to 50 mm, and most preferably the inner diameter is 3 to 8 mm. is there. For example, assembling a refrigerator of about 2000 cal / h using Freon refrigerant R134a, the inner diameter of the thin tube is 5 mm, the number of turns is 23, the diameter of the spiral is 30 mm, and the length of the thin tube is 2.3 m. The

refrigerant pipes

2 and 4 have an inner diameter of 7.7 mm, and the refrigerant pipe 13 and the suction pipe 12 have an inner diameter of 10.7 mm.

When the partially liquefied refrigerant enters the spiral tube 6, the refrigerant is accelerated by the suction action of the compressor 1 (referred to as an acceleration phenomenon of the refrigerant), and the amount of liquefaction is increased with reduced pressure and enthalpy reduction. The liquid is almost liquefied and becomes a medium pressure (0.4 to 0.6 MPa) liquid refrigerant at the outlet of the spiral tube 6 (point j to point k in FIG. 2). The conventional general cupilary tube is quite a thin tube, and the kinetic energy of the refrigerant is lost due to the contracted flow and vortex generated by constricting the high-pressure liquid refrigerant, causing a pressure drop. The helical tube 6 has a very small constriction loss due to contraction / vortex and a small loss of kinetic energy of the refrigerant. In addition, the gas-liquid separation of the refrigerant is promoted by the centrifugal force in the spiral tube. Specifically, the liquid phase part is concentrated and attached to the inner wall of the pipe, and the gas phase part flows through the central part of the pipe. A choke phenomenon due to the acceleration of the refrigerant appears, and the refrigerant is liquefied with a reduced pressure and a decrease in enthalpy. The main cause of the temperature drop in the spiral tube 6 is that the enthalpy of the refrigerant, which is thermal energy, is converted into velocity energy in the spiral tube 6 and the enthalpy of the refrigerant is reduced, leading to the phenomenon of a decrease in static temperature. Judged to be. That is, the helical tube 6 constitutes an energy conversion device that converts enthalpy into velocity energy.
In the design of the refrigeration system, the flow rate of the refrigerant in the spiral tube 6 is preferably set to be twice or more the flow rate in the mini heat exchange device 3.

In this configuration, the decompression liquefaction part is a spiral tube 6 wound spirally, but as shown in FIG. 2, as long as the gas refrigerant can be substantially liquefied with decompression and enthalpy reduction, It is not limited to a spiral tube, but may be a meandering tube or a straight tube. In this case, it is desirable to provide appropriate throttle means at the inlet of the meandering pipe or straight pipe, or at a plurality of locations in the middle of the pipe. In either case, in the reduced pressure liquefaction unit, the gas refrigerant is almost liquefied by means other than heat dissipation, that is, by conversion to enthalpy velocity energy.

The refrigerant that has become medium pressure liquid refrigerant in the helical tube 6 enters the helical thin tube 8 through the branch tube 7. As shown in FIG. 3 (d), the spiral tubule 8 has a form in which the tubule is spirally wound in the same manner as the spiral tube 6. The inner diameter of the spiral tube 8 is set to be smaller than the inner diameter of the spiral tube 6. For example, when the inner diameter of the spiral tube 6 is set to 3 to 8 mm, the inner diameter of the spiral tube 8 is desirably 1.2 to 3 mm. In the present embodiment, two spirally wound pieces are connected in parallel, but three or more pieces may be connected in parallel, or even one. Also, two spiral tubules with different winding directions connected in series, or a configuration in which they are further connected in parallel may be used. It is preferable that the cross-sectional area of the portion through which the refrigerant passes through the helical thin tube 8 (the sum of the plurality of cross-sectional areas connected in parallel) is smaller than the cross-sectional area of the threaded tube 6. By reducing the cross-sectional area, as described later, the refrigerant spins through the spiral tubule 8 and is accelerated to reduce the pressure, so that the cooling effect is enhanced.
For example, in the case of a refrigerator of about 2000 cal / h, the inner diameter of the thin tube is 2.5 mm, the winding number is 19 turns, the spiral diameter is 15 mm, and the length of the thin tube is 0.72 m. Configured.

As shown in FIG. 3C, the branch pipe 7 branches the refrigerant exiting from one spiral pipe 6 into two spiral narrow pipes 8. The main portion (thick portion) of the branch pipe 7 has a substantially cylindrical shape with a length L2 of 10 to 50 mm and an inner diameter D2 of 10 to 20 mm. Both ends connected to the helical tube 6 and the helical thin tube 8 are in the shape of a cylinder that can be connected by inserting the helical tube 6 and the helical thin tube 8 respectively. In the present embodiment, since the spiral tubule 8 is formed of two tubules, the connection side of the branch tube 7 has two connection holes. However, the number of connection holes is as follows. The number of tubules constituting the spiral tubule 8 is matched.
For example, the inner diameter D2 is preferably set larger than the inner diameter of either the spiral tube 6 or the spiral capillary tube 8.

When the substantially liquefied refrigerant enters the spiral tubule 8, the refrigerant is accelerated by the suction action of the compressor 1 or the like (referred to as an acceleration phenomenon of the refrigerant), and the liquefied refrigerant is cooled with reduced pressure and enthalpy reduction. . At the outlet of the helical thin tube 8, the pressure is reduced and cooled to become a low-temperature liquid, and the pressure is lowered to become a low-pressure (0.4 MPa or less) liquid (point k to point l in FIG. 2). The refrigerant in the spiral capillary 8 changes in a state substantially along the saturated liquid line L, as shown in FIG. Compared with a conventional cupilary tube, the helical thin tube 8 also has a very small throttle loss due to contraction and vortex and a small loss of kinetic energy of the refrigerant. In addition, the gas-liquid separation of the refrigerant is promoted by the centrifugal force in the spiral tubule 8, specifically, the liquid phase part is concentrated and attached to the inner wall of the pipe, and the gas phase part flows through the central part of the pipe. Similarly, a choke phenomenon due to the acceleration of the refrigerant appears in the phase portion, and the refrigerant is liquefied with a reduced pressure and a decrease in enthalpy. Note that the position where the choke phenomenon appears varies depending on the configuration of each device or the change in the flow rate, flow rate, etc. of the refrigerant. For example, it does not appear in the spiral tube 6 but may appear only after reaching the spiral tube 8.
The main cause of the temperature drop in the spiral tube 8 is that, similarly to the temperature drop in the spiral tube 6, the enthalpy of the refrigerant, which is thermal energy, is converted into velocity energy, the enthalpy is reduced, and the static temperature is lowered. It is judged that the phenomenon has occurred.
That is, like the spiral tube 6, the spiral capillary 8 also constitutes an energy conversion device that converts the enthalpy of the refrigerant into velocity energy.
In the design of the present refrigeration system, the flow rate of the refrigerant in the spiral capillary 8 is preferably at least twice the flow rate in the mini heat exchanger 3 and higher than the flow rate in the spiral tube 6.

In this configuration, the spiral tubule 8 is used. However, the configuration is not limited to a spiral shape and may be a meandering tube or a straight tube as long as the liquid refrigerant can be cooled with reduced pressure and enthalpy reduction. In this case, it is desirable to insert appropriate throttle means at the inlet of the meandering pipe or straight pipe, or at a plurality of locations in the middle of the pipe. In any case, in this configuration, the liquid refrigerant is cooled by means other than heat dissipation, that is, by conversion to enthalpy velocity energy.

In the present embodiment, as shown in FIG. 4, the helical tubule 8, the collecting tube 9 and the tubular body 10 are arranged vertically along the vertical line, and the tubular body 10 is connected to the outlet of the collecting tube 9. .
The tube body 10 has a substantially cylindrical tube body 10A having an enlarged diameter. The tube body 10A is formed with an inlet 10B and an outlet 10C to which a refrigerant pipe is connected, respectively. A resistor 10D that resists the flow of the refrigerant is accommodated. The resistor 10D is supported by a spring 10E, and when the refrigerant flows, the resistor 10D is pressed toward the outlet 10C against the spring force of the spring 10E. In this tubular body 10, the spring force of the spring 10 E, the weight of the resistor 10 D, and the like are set so that the resistor 10 D becomes resistant to the flow of the refrigerant and the flow rate of the refrigerant becomes substantially zero at the outlet 10 C.
In this configuration, as described above, the refrigerant has a large velocity energy at the outlet of the spiral capillary 8, and if the refrigerant flows into the downstream refrigerant pipe 13 while maintaining this large velocity energy, Frictional heat is generated in the refrigerant pipe 13, and the dryness of the refrigerant is increased accordingly. That is, the refrigerant in the spiral capillary 8 reaches from the state of point k shown in FIG. 2 to the state of point l. However, when frictional heat is generated in the refrigerant pipe 13, the degree of dryness of the refrigerant is increased, After passing through l, it moves to the state of the point m, and the heat absorption amount in the evaporator (heat absorption part) 11 is lowered from Q1 to Q2, and the efficiency is lowered by the difference.

In the present embodiment, the refrigerant that has become a low-temperature liquid by the helical thin tube 8 flows from below into the tubular bodies (energy conversion means) 10 that are arranged vertically along the vertical line, and in this tubular body 10 Against the spring force of the spring 10E, the resistor 10D is pushed away and flows out from above. During this time, the velocity energy of the refrigerant is converted into work energy (or potential energy).
That is, the velocity energy is released as work energy (or potential energy), and the flow velocity becomes substantially zero at the outlet 10 C of the tube body 10.
According to this configuration, since the flow rate of the refrigerant becomes substantially zero at the outlet 10C of the tube body 10, frictional heat is not generated in the refrigerant pipe 13, and therefore, the dryness of the refrigerant is maintained. After passing through the point l, it does not move to the state of the point m, the heat absorption amount in the evaporator (heat absorption part) 11 can be maintained at Q1, and the efficiency can be maintained.
Further, when a liquid refrigerant is present in the tube main body 10A, the liquid refrigerant falls in the tube main body 10A due to its own weight. When the rising liquid refrigerant collides with the falling liquid refrigerant, the speed energy of the rising refrigerant is converted into work energy also by this collision. Therefore, through the verification test, the energy consumed by this collision is obtained, and the energy of this energy is taken into account, and the spring force of the spring 10E, the weight of the resistor 10D, etc. so that the flow velocity of the refrigerant at the outlet 10C becomes almost zero. It is desirable to set

In the evaporator 11, the refrigerant evaporates due to the endothermic heat of isobaric and isothermal expansion (point l to point h in FIG. 2), thereby completing the cycle in FIG.
In the above embodiment, the helical tube 6 and the helical tube 8 are connected in series, but the capacity of the mini heat exchange device 3 is increased, and the helical tube 6 is omitted at the outlet of the heat exchange device 3. Thus, it is possible to connect the spiral tubule 8 directly. In this case as well, the efficiency can be improved by connecting the tubular body 10 to the outlet of the helical thin tube 8.

The configuration of the tube body 10 is not limited to the above configuration, and any configuration may be adopted as long as the configuration can convert the velocity energy of the refrigerant into work energy (or potential energy).
For example, the form of the resistor 10D may be any form such as a sphere, a flat plate, a polygon, or a polygonal cone, and the spring 10E may be omitted. In addition, the tubular bodies 10 are arranged vertically and vertically, but may be arranged in a horizontal arrangement, an oblique arrangement, or the like. When the velocity energy of the refrigerant is not converted into work energy but is converted only into potential energy, the resistor 10D and the spring 10E may be omitted, and the tubes 10 may be arranged vertically. In this case, the velocity energy is converted into potential energy in the process in which the refrigerant having velocity rises in the tube body 10.

FIG. 5 shows another form. An impeller 25 is disposed in the tube body 10, and a generator 27 is connected to a rotating shaft 26 of the impeller 25. The generator 27 is, for example, a power supply circuit for a driving motor of the compressor 1 shown in FIG. (Not shown). In this configuration, the energy efficiency is further improved by rotating the impeller 25 with the velocity energy of the refrigerant, converting the energy into work energy, and collecting the electric power of the generator 27 in the power supply circuit.

6A to 6C are Ph diagrams, respectively.
In FIG. 6A (conventional refrigeration cycle), the change from point 3 to point 4 is an isentropic change, and the flow rate is small due to pipe friction, so the energy equation is expressed as follows.
dh (enthalpy change) = dq (heat quantity from outside) −dwt (work quantity to outside)
Therefore, the changes of points 3 to 4 are as follows.
h ₄ −h ₃ = q ₃₄ (friction heat) −wt ₃₄ (pressure loss)
Since the heat generated by friction is equal to the pressure loss, it is expressed by a change in isenthalpy as shown in the following equation.
h ₄ -h ₃ = 0

In FIG. 6B (using helical tubules 6 and 8), since the influence of tube friction is small and the change in kinetic energy cannot be ignored, the energy equation is expressed as follows.
dh (enthalpy change) + wdw (kinetic energy) = dq (heat quantity from the outside) −dwt (work quantity to the outside)
Therefore, the change from point 3 to point 3 ′ is as follows.
h _{3 ′} −h ₃ + {(w _{3 ′} ) ² − (w ₃ ) ² } / 2 = q _{33 ′} (friction heat) −wt _{33 ′} (pressure loss)
Here, the change from point 3 to point 3 ′ is an adiabatic change (q _{33 ′} = 0), and the speed is the speed of sound (w _{3 ′} = w _C ).
h _{3 ′} −h ₃ + {(w _C ) ² − (w ₃ ) ² } / 2 = 0−wt _{33 ′}
Next, since the speed does not change at points 3 ′ to 4, it is expressed by the following equation.
h ₄ −h _{3 ′} + {(w _C ) ² − (w _C ) ² } / 2 = 0−wt _3′4
Finally, the change from point 4 to point 4 ′ is represented by the following equation because the flow rate is zero due to friction, there is no adiabatic change, and no pressure change.
h _{4 ′} −h ₄ + {0− (wc) ² } / 2 = q _4′4 −wt _4′4 = 0 + v (P _{4 ′} −P ₄ ) = 0
Also, since the kinetic energy due to friction changes to thermal energy,
^{qf = (wc) 2/2} ( frictional heat)
Therefore, h _{4 ′} = h ₄ + qf
Thus, it can be seen that the refrigeration effect is reduced by the frictional heat qf.

On the other hand, in FIG. 6C (combination of the

helical tubules

6 and 8 and the tubular body 10), the changes of the points 3 to 3 ′ are expressed as follows.
h _{3 ′} −h ₃ + {(w _{3 ′} ) ² − (w ₃ ) ² } / 2 = q _{33 ′} −wt _{33 ′}
Here, the change from point 3 to point 3 ′ is an adiabatic change (q _{33 ′} = 0), and the speed is the speed of sound (w _{3 ′} = w _C ).
h _{3 ′} −h ₃ + {(w _C ) ² − (w ₃ ) ² } / 2 = 0−wt _{33 ′}
Next, at points 3 ′ to 4, since external work is performed on the tube 10 and the speed becomes zero, it is expressed by the following equation.
_{_{h 4 -h 3 '+ {0-}} (w C) 2} / 2 = 0-wt 3'4 -wt out
Here, considering that work to the outside is equivalent to kinetic energy,
_{_{^{wt out = (w C) 2}}} /2
Accordingly, the energy equation is as follows, and the refrigerating capacity is not reduced by frictional heat.
_{_{_{h 4 = h 3 '-wt 3'4}}} = h 3 + v (P 4 -P 3')

For work to the outside, there is a method of converting into potential energy as in the following equation.
_{_{^{wt out = (w C) 2}}} /2 = gz
Here, z is the height in the vertical direction.
There is also a method of converting to spring work as in the following equation.
_{_{^{wt out = (w C) 2}}} /2 = k (l 2/2)
Here, l is the displacement of the spring, and k is the spring constant.
Further, there is a method of converting into potential energy and spring work as in the following equation.
_{_{^{wt out = (w C) 2}}} /2 = k (l 2/2) + gz
In addition, there is a method of converting to rotary work as in the following equation.
_{_{^{wt out = (w C) 2}}} /2 = I (ω 2/2)
Here, I is the moment of inertia of the rotating body, and ω is the angular velocity of the rotating body.

In the heat conversion apparatus 30 for condensation during this cycle, a part (5 to 50% by weight) of the refrigerant is liquefied (point i to point j) in the isobaric cooling section (mini heat exchange apparatus 3), and the reduced pressure liquefying section. In (helical tube 6), the refrigerant is accelerated, and the remaining gas refrigerant partially liquefied is substantially liquefied (point j to point k) with decompression and refrigerant enthalpy reduction. In the narrow tube 8), the refrigerant is accelerated and the liquefied refrigerant is supercooled (point k to point l) with decompression and refrigerant enthalpy reduction, so that the COP of the refrigeration cycle is improved. Further, since the refrigerant is depressurized by the condensing heat conversion device 30, there is no need for a depressurization mechanism such as a narrow tube (generally, a capillary tube having an inner diameter of about 0.8 mm) or an expansion valve as in the prior art. The cycle can be simplified. Furthermore, in the reduced pressure liquefaction part (spiral tube 6) and the reduced pressure cooling part (spiral capillary 8), the refrigerant enthalpy, which is thermal energy, is converted into velocity energy, the refrigerant enthalpy is reduced, and the phenomenon of a decrease in static temperature occurs. Therefore, the heat exchange device can be downsized as compared with the case of heat dissipation.
In the present embodiment, the heat conversion device 30 for condensation is composed of an isobaric cooling unit (mini heat exchange device 3), a reduced pressure liquefying unit (spiral tube 6), and a reduced pressure cooling unit (spiral tubule 8). The decompression liquefaction section (spiral tube 6) may be configured by connecting a plurality of spiral tubes in series. In this case, at points j to k in FIG. Become. The reduced-pressure cooling unit (spiral thin tube 8) may also be configured by connecting a plurality of spiral tubes in series. In this case, points k to l in FIG. 2 are cycle lines having a plurality of bending points. .

FIG. 7 shows another embodiment. This cycle is a refrigeration cycle using a so-called ejector that converts pressure energy of refrigerant into velocity energy.
Reference numeral 61 denotes a compressor. A radiator 62 is connected to the compressor 61, and a receiver tank 63 is connected to the radiator 62. An ejector 64 is connected to the receiver tank 63, and the ejector 64 decompresses and expands the refrigerant flowing out of the radiator 62, sucks the gas-phase refrigerant evaporated in the evaporator 65 described later from the suction part 64 A, and expands energy. Is converted into pressure energy to increase the suction pressure of the compressor 61.
The refrigerant flowing out from the ejector 64 is sucked into the compressor 61, thereby forming a refrigerant circulation path. Between the receiver tank 63 and a later-described nozzle 64B of the ejector 64, there is provided a branch channel 66 for guiding the branched refrigerant flow to the suction section 64A, and an evaporator 65 is provided in the branch channel 66. ing.

Further, on the upstream side of the refrigerant flow of the evaporator 65, the refrigerant sucked into the evaporator 65 is decompressed to reliably reduce the pressure (evaporation pressure) in the evaporator 65 and the refrigerant flowing into the evaporator 65. A fixed throttle 67 such as a capillary tube is provided as a throttle means for adjusting the flow rate (cooling capacity generated in the evaporator 65).

The ejector 64 sucks the gas-phase refrigerant evaporated by the nozzle 64B and the evaporator 65 that convert the pressure energy (pressure head) of the high-pressure refrigerant flowing out of the radiator 62 into velocity energy (speed head) to decompress and expand the refrigerant. 64A, a mixing unit that mixes the refrigerant injected from the nozzle 64B and the refrigerant sucked from the evaporator 65 while sucking the refrigerant from the suction unit 64A by a high-speed refrigerant flow (jet flow) injected from the nozzle 64B And a diffuser section 64C for converting the velocity energy of the refrigerant flowing out from the mixing section into pressure energy and increasing the pressure of the refrigerant.
The tube body 10 is connected to the outlet of the diffuser portion 64 C, and the compressor 61 is connected to the tube body 10 via an accumulator 68. The tubular body 10 has substantially the same structure as that of the above embodiment, and the description thereof is omitted. Note that it is desirable that the diffuser portion 64C and the tube body 10 are arranged vertically along a vertical line.
At the outlet of the diffuser portion 64C, the refrigerant has a large velocity energy. If the refrigerant flows into the downstream refrigerant pipe 69 while maintaining this large velocity energy, friction heat is generated in the refrigerant pipe 69, and accordingly, The dryness of the refrigerant is increased, the amount of heat absorbed in the evaporator 65 is reduced, and the efficiency is reduced accordingly.

In the present embodiment, the outlet refrigerant of the diffuser portion 64C flows from below into the tubular bodies (energy conversion means) 10 arranged vertically along the vertical line, and the spring of the spring 10E is within the tubular body 10. Against the force, the resistor 10D is pushed away and flows out from above, during which the velocity energy of the refrigerant is converted into work energy (or potential energy).
That is, the velocity energy is released as work energy (or potential energy), and the flow velocity becomes substantially zero at the outlet 10 C of the tube body 10.
According to this configuration, since the flow velocity of the refrigerant becomes substantially zero at the outlet 10C of the tube body 10, frictional heat is not generated in the refrigerant pipe 69, and the dryness of the refrigerant is maintained. The heat absorption amount at 65 can be maintained, and the efficiency can be maintained.
The configuration of the tube body 10 is not limited to the above configuration, and it is sufficient that the velocity energy of the refrigerant can be converted into work energy (or potential energy).

DESCRIPTION OF

SYMBOLS

1,61

Compressor

2, 4, 10 Refrigerant piping 3 Mini heat exchanger 3-1 Mini fan 6 Spiral tube 8 Spiral thin tube 10

Tubing

11, 65

Evaporator

13, 62 Condenser 64 Ejector

Claims

A compressor,
A radiator,
Refrigerant liquefaction means for converting the pressure energy of the refrigerant into velocity energy to liquefy the refrigerant;
Energy conversion means for converting the velocity energy of the refrigerant from the refrigerant liquefaction means into potential energy or work energy while maintaining dryness;
A refrigeration system characterized in that a heat absorber is connected in order with a refrigerant pipe in order.
The refrigeration system of claim 1,
The refrigeration system, wherein the refrigerant liquefying means is a spiral tube that depressurizes the refrigerant and reduces the enthalpy by using a choke phenomenon caused by acceleration of the refrigerant.
The refrigeration system according to claim 1 or 2,
The refrigeration system, wherein the energy conversion means includes a tube body that contains a resistor, and converts the velocity energy of the refrigerant into work energy when the refrigerant exceeds the resistor in the tube body.
The refrigeration system according to claim 3,
A refrigeration system, wherein the resistor in the tube is spring-biased toward the inlet side of the tube.
The refrigeration system according to claim 1 or 2,
The refrigeration system characterized in that the energy conversion means includes a tube body containing a rotating body, and converts the velocity energy of the refrigerant into work energy when the refrigerant rotates the rotating body in the tube body.
The refrigeration system of claim 1,
The refrigerant liquefying means is provided vertically so that the refrigerant flows from the bottom to the top,
The refrigeration system, wherein the energy conversion means converts the velocity energy of the refrigerant into potential energy.
The refrigeration system of claim 1,
The refrigeration system, wherein the refrigerant liquefying means is an ejector that converts pressure energy of refrigerant into velocity energy.