Absorber and cold generation system

Abstract

This invention provides an absorber and a cooling / heating system that can suppress variations in the flow rate of refrigerant ejected from multiple outlets in a refrigerant supply pipe. [Solution] The system comprises a plurality of plates 10 arranged alternately to form a first flow path 11 through which a fluid to be heated flows and a second flow path 12 through which an absorbent liquid and a refrigerant flow, an absorbent liquid supply unit 3 located at one end of the plurality of plates 10 and communicating with the second flow path 12, and a refrigerant supply pipe 4 located inside the absorbent liquid supply unit 3 and through which the refrigerant flows. The refrigerant supply pipe 4 is arranged in a direction perpendicular to the plurality of plates 10 and has a plurality of nozzles 41 that eject refrigerant at positions corresponding to the second flow path 2. The system is formed such that, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4, at least one of the cross-sectional area of ​​the refrigerant supply pipe 4 and the cross-sectional area of ​​the plurality of nozzles 41 decreases as one moves downstream of the refrigerant supply pipe 4.

Description

Technical Field

[0001] The present invention relates to an absorber and a cold heat generation system.

Background Art

[0002] Conventionally, an absorption cycle has been used as a cold heat generation system in an air conditioner. In the absorption cycle, an object to be cooled is cooled by evaporating a refrigerant under low pressure. At this time, the refrigerant is absorbed by an absorbent in an absorber so as to enable evaporation of the refrigerant under low pressure.

[0003] In order to improve the performance of the absorber, various studies have been conducted to increase the efficiency of heat exchange in the absorber. For example, Japanese Patent Laid-Open No. 2006-162154 (Patent Document 1) discloses a laminated plate type absorber configured by laminating a plurality of heat transfer plates. According to the technique of Patent Document 1, it is possible to achieve both downsizing and high performance of the absorber.

Prior Art Documents

Patent Documents

[0004]

Patent Document 1

Summary of Invention

Problems to be Solved by the Invention

[0005] In the technique of Patent Document 1, a refrigerant supply pipe is arranged in a direction orthogonal to the laminated plates, and the refrigerant supply pipe is configured to have a plurality of ejection ports along the axial direction. In that case, the flow rate of the refrigerant ejected from the plurality of ejection ports provided in the refrigerant supply pipe may vary in the axial direction of the refrigerant supply pipe due to the influence of the internal pressure of the refrigerant supply pipe. If the flow rate of the refrigerant ejected from the plurality of ejection ports varies, the absorption of the refrigerant by the absorbent cannot be performed efficiently in the laminated plate type absorber.

[0006] Therefore, there is a need to realize an absorber and a cooling system that can suppress variations in the flow rate of refrigerant ejected from multiple outlets provided in the refrigerant supply pipe. [Means for solving the problem]

[0007] The characteristic configuration of the absorber according to the present invention is that it comprises a plurality of plates arranged so that a first flow path through which a heated fluid flows and a second flow path through which an absorbent liquid and a refrigerant flow alternately form between them; an absorbent liquid supply unit arranged at one end of the plurality of plates and communicating with the second flow path; and a refrigerant supply pipe arranged inside the absorbent liquid supply unit and through which the refrigerant flows. The refrigerant supply pipe is arranged in a direction perpendicular to the plurality of plates and has a plurality of nozzles that eject the refrigerant at positions corresponding to the second flow path, and is formed such that at least one of the cross-sectional area of ​​the refrigerant supply pipe and the cross-sectional area of ​​the plurality of nozzles decreases as one moves downstream of the refrigerant supply pipe, depending on the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe.

[0008] The pressure inside a refrigerant supply pipe can be expressed in terms of total pressure, static pressure, and dynamic pressure. Here, static pressure is the pressure when the fluid is at rest, and dynamic pressure is the kinetic energy of the fluid expressed in units of pressure. Total pressure is the sum of static pressure and dynamic pressure, and is always constant in frictionless flow inside the refrigerant supply pipe. Also, inside a refrigerant supply pipe with a closed downstream end, when the static pressure decreases, the dynamic pressure increases and the refrigerant flow accelerates, and when the static pressure increases, the dynamic pressure decreases and the refrigerant flow decelerates.

[0009] Furthermore, if the density of the refrigerant is constant inside the refrigerant supply pipe, and the cross-sectional area of ​​the refrigerant supply pipe and the cross-sectional areas of the multiple outlets are constant, the static pressure tends to be higher downstream. Therefore, the flow rate of refrigerant discharged from the outlets increases downstream in the refrigerant supply pipe. For this reason, it is thought that the flow rate of refrigerant at each outlet can be adjusted by increasing or decreasing at least one of the cross-sectional area of ​​the refrigerant supply pipe and the cross-sectional area of ​​the multiple outlets.

[0010] Based on the above reasoning, in this configuration, in the absorber, the refrigerant supply pipe is formed such that, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe, at least one of the cross-sectional area of ​​the refrigerant supply pipe and the cross-sectional area of ​​the multiple outlets decreases as it moves downstream, so that the difference in the refrigerant flow rates ejected from the multiple outlets in the absorber is within a predetermined range. In other words, by reducing the cross-sectional area of ​​the refrigerant supply pipe or the cross-sectional area of ​​the multiple outlets, or by reducing the cross-sectional area of ​​both the refrigerant supply pipe and the multiple outlets as it moves downstream, it is considered that the variation in the refrigerant flow rate ejected from the multiple outlets in the refrigerant supply pipe toward the second flow path is suppressed and made uniform. As a result, the absorber can suppress the pressure loss of the refrigerant at the inlet of the second flow path. Consequently, the absorber can efficiently absorb the refrigerant with the absorbent liquid, and the absorption characteristics of the absorber can be improved.

[0011] Preferred embodiments of the present invention will be described below. However, the scope of the present invention is not limited by the examples of preferred embodiments described below.

[0012] A further characteristic feature of the absorber according to the present invention is that the plurality of outlets are formed such that their cross-sectional area decreases as they move downstream of the refrigerant supply pipe, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe.

[0013] By adopting this configuration, it was confirmed that variations in the refrigerant flow rate ejected from multiple outlets in the refrigerant supply pipe toward the second flow path are suppressed and made uniform. As a result, the absorber can suppress the pressure loss of the refrigerant at the inlet of the second flow path. Consequently, the absorber can efficiently absorb the refrigerant with the absorbent liquid, improving the absorption characteristics of the absorber.

[0014] A further characteristic feature of the absorber according to the present invention is that the plurality of nozzles are divided into a plurality of groups toward the downstream side of the refrigerant supply pipe, and the single or plurality of nozzles included in the divided group have the same cross-sectional area, and the plurality of nozzles are formed such that the cross-sectional area of ​​the nozzles included in the divided group decreases as one moves toward the downstream side of the refrigerant supply pipe.

[0015] In order to efficiently suppress variations in the flow rate of the refrigerant being discharged, it is preferable for the cross-sectional area of ​​the multiple outlets provided in the refrigerant supply pipe to decrease continuously toward the downstream side of the refrigerant supply pipe. However, in that case, it is necessary to minimize the difference in cross-sectional area between adjacent outlets along the flow path of the refrigerant supply pipe, which requires high precision in the processing of the multiple outlets. On the other hand, with this configuration, the cross-sectional area of ​​the multiple outlets can be changed in groups, so the difference in cross-sectional area between adjacent outlets along the flow path of the refrigerant supply pipe can be increased. This makes it easier to process the multiple outlets, thus reducing the manufacturing cost of the absorber. Furthermore, since the cross-sectional area of ​​the multiple outlets decreases in groups toward the downstream side of the refrigerant supply pipe, variations in the flow rate of the refrigerant discharged from the multiple outlets can be suppressed to an acceptable degree, and pressure loss of the refrigerant at the inlet of the absorber can be suppressed.

[0016] A further characteristic feature of the absorber according to the present invention is that the refrigerant flowing through the refrigerant supply pipe is one of R1234yf, R1234ze(E), R1233zd(E), and R1224yd(Z).

[0017] This configuration allows the absorber to use a refrigerant with a Global Warming Potential (GWP) of level 1 as the refrigerant flowing through the refrigerant supply pipe, thus contributing to environmental protection.

[0018] The characteristic configuration of the cold energy generation system according to the present invention is that it uses one of the above-mentioned absorbers and is an absorption type or a compression-absorption type that combines a compression type and an absorption type.

[0019] According to this configuration, as the absorber of a compression-absorption hybrid system (an example of a cold heat generation system) that can utilize relatively low-temperature heat such as solar heat and waste heat, a simple plate heat exchanger type absorber can be used while minimizing adverse effects on cycle performance as much as possible. In addition, as an absorption-type cold heat generation system, it can also be used for, for example, refrigerant subcooling of gas heat pump air conditioners.

[0020] Further features and advantages of the present invention will become clearer from the following description of exemplary and non-limiting embodiments described with reference to the drawings. Brief Description of the Drawings

[0021] [Figure 1] It is a flow diagram of the air conditioning system according to the embodiment. [Figure 2] It is a cross-sectional view of the absorber according to the embodiment. [Figure 3] It is a diagram showing changes in static pressure when the cross-sectional area of the ejection hole is constant. [Figure 4] It is a diagram showing the flow rate and static pressure with mathematical expressions. [Figure 5] It is a flow diagram of the air conditioning system according to Test 1 and Test 2. [Figure 6] It is a table showing the conditions and results of Test 1. [Figure 7] It is a diagram showing the change in the cross-sectional area of the ejection port in Test 1. [Figure 8] It is a graph showing the results of Test 1. [Figure 9] It is a table showing the conditions and results of Test 2. [Figure 10] It is a diagram showing the change in the cross-sectional area of the ejection port in Test 2. [Figure 11] It is a graph showing the results of Test 2. [Figure 12] It is a flow diagram of an air conditioning system according to another embodiment. Modes for Carrying Out the Invention

[0022] Embodiments of the absorber and cooling system according to the present invention will be described with reference to the drawings. Below, an example will be described in which the absorber according to the present invention is applied to an absorber 1 provided in an air conditioning system 100 for a car air conditioner (an example of a cooling system).

[0023] [Air conditioning system configuration] The air conditioning system 100 according to this embodiment has a configuration in which an absorption refrigeration cycle and a compression refrigeration cycle are integrated in a manner that shares a refrigerant. The air conditioning system 100 has a high-pressure stage 110, a low-pressure stage 120, and a gas-liquid separator 130 (Figure 1).

[0024] The high-pressure stage 110 is equipped with a pump 111, a solution depressurization device 112, a solution heat exchanger 113, a regenerator 114, a condenser 115, a refrigerant expansion device 116, and an absorber 1.

[0025] The low-pressure stage 120 is equipped with a refrigerant expansion device 121, an evaporator 122, and a compressor 123. The compressor 123 is driven by a motor 124. However, the compressor 123 may also be driven by a power source such as an engine.

[0026] The gas-liquid separator 130 is located at the boundary between the high-pressure stage 110 and the low-pressure stage 120. In the high-pressure stage 110, it is located downstream of the refrigerant expansion device 116 and upstream of the absorber 1. In the low-pressure stage 120, it is located downstream of the compressor 123 and upstream of the refrigerant expansion device 121.

[0027] In this embodiment, for example, an HFO-based refrigerant such as R1234yf is used as the refrigerant. The refrigerant is not limited to R1234yf, and may also be R1234zeE, R1233zdE, or R1224ydZ. In this embodiment, the absorbent may be, but is not limited to, ethylene glycol monobutyl ether, ethylene glycol monoisobutyl ether, diethylene glycol diethyl ether, diethylene glycol dibutyl ether, diethylene glycol isopropyl methyl ether, triethylene glycol butyl methyl ether, propylene glycol monopropyl ether, tetraethylene glycol dimethyl ether, and diethylene glycol monoethyl ether acetate.

[0028] [Air conditioning system operation] The following describes the case when the air conditioning system 100 is operating in cooling mode. The evaporator 122 cools the air in the target area. That is, in the evaporator 122, the refrigerant absorbs heat from the air in the target area and evaporates. The refrigerant is compressed in the compressor 123 to become superheated steam, which is then introduced into the gas-liquid separator 130.

[0029] As will be described later, the refrigerant separated from the absorbent liquid (strong solution) that absorbed the refrigerant in the regenerator 114 is condensed in the condenser 115. The condensed refrigerant is depressurized in the refrigerant expansion device 116 to become wet vapor, which is then introduced into the gas-liquid separator 130. In the gas-liquid separator 130, this wet vapor refrigerant is mixed with the aforementioned superheated vapor refrigerant compressed in the compressor 123, and then separated into saturated liquid and saturated vapor.

[0030] The gaseous component (saturated vapor) of the refrigerant separated in the gas-liquid separator 130 flows through the high-pressure stage 110. The gaseous component of the refrigerant is absorbed by the absorbent liquid in the absorber 1. In this embodiment, the absorber 1 is water-cooled as described later, and the heat generated by the mixing of the refrigerant and the absorbent liquid is removed by water cooling. The absorbent liquid (strong solution) that has absorbed the refrigerant is pressurized by the pump 111 and goes to the solution heat exchanger 113, where it is preheated before going to the regenerator 114.

[0031] In the regenerator 114, the absorbent liquid (strong solution) that has absorbed the refrigerant is heated. At this time, some of the refrigerant vaporizes due to the difference in boiling points between the refrigerant and the absorbent liquid. The energy for heating is supplied by waste heat, etc.

[0032] In the regenerator 114, the absorbent liquid (weak solution) from which some of the refrigerant has been separated is cooled in the solution heat exchanger 113, and then depressurized in the solution depressurization device 112 before returning to the absorber 1.

[0033] The refrigerant separated from the absorbent liquid (strong solution) in the regenerator 114 is condensed in the condenser 115. As described above, the condensed refrigerant is depressurized in the refrigerant expansion device 116 to become wet vapor, which is then introduced into the gas-liquid separator 130.

[0034] However, in situations where there is little waste heat, there may be insufficient energy to heat the absorbent liquid (strong solution) that has absorbed the refrigerant in the regenerator 114. In this case, in addition to the path that includes the absorber 1, pump 111, solution heat exchanger 113, and regenerator 114, a path that includes the compressor 117 is also used. In this case, a portion of the gaseous component of the refrigerant separated in the gas-liquid separator 130 is compressed in the compressor 117 to become superheated steam, which is then introduced into the condenser 115 along with the refrigerant separated from the absorbent liquid (strong solution) in the regenerator 114. The compressor 117 is driven by a motor 118. However, the compressor 117 may also be driven by a power source such as an engine. If there is almost no energy to heat the absorbent liquid (strong solution) that has absorbed the refrigerant in the regenerator 114, all of the gaseous component of the refrigerant separated in the gas-liquid separator 130 is compressed in the compressor 117 to become superheated steam, which is then introduced into the condenser 115.

[0035] [Composition of the absorber] The absorber 1 according to this embodiment comprises a laminated plate heat exchanger 2, an absorbent liquid supply unit 3, a refrigerant supply pipe 4, a first flow path 11, and a second flow path 12 (Figure 2). The laminated plate heat exchanger 2 is in fluid communication with the absorbent liquid supply unit 3 at one end 2a. The absorbent liquid supply unit 3 houses the refrigerant supply pipe 4, and the absorbent liquid supply unit 3 and the refrigerant supply pipe 4 are in fluid communication via an outlet 41 provided in the refrigerant supply pipe 4.

[0036] The laminated plate heat exchanger 2 is composed of a plurality of plates 10 arranged inside a housing 20. The plurality of plates 10 are arranged such that first flow paths 11 and second flow paths 12 are alternately formed between them. The first flow paths 11 and the second flow paths 12 are demarcated by the plurality of plates 10. The first flow paths 11 and the second flow paths 12 are arranged with spacing between them in the longitudinal direction (axial X, direction X in Figure 2) of the refrigerant supply pipe 4, with the surfaces of the plurality of plates 10 intersecting (approximately perpendicular to) the longitudinal direction (axial X, direction X in Figure 2) of the refrigerant supply pipe 4. The first flow paths 11, located at both ends in the longitudinal direction (direction X in Figure 2) of the refrigerant supply pipe 4, may include a part of the housing 20. Cooling water W (an example of a fluid to be heated) flows through the first flow paths 11. Absorbent liquid and refrigerant flow through the second flow paths 12. Then, heat exchange occurs between the cooling water W flowing through the first channel 11 and the fluid (a mixture of refrigerant and absorbent) flowing through the second channel 12 via the plate 10.

[0037] As shown in Figure 2, the refrigerant supply pipe 4 is arranged in a direction perpendicular to the planes of the multiple plates 10 and has multiple nozzles 41 for ejecting refrigerant. The longitudinal position (direction X in Figure 2) of the refrigerant supply pipe 4 where the nozzles 41 are located substantially coincides with the position where the laminated plate heat exchanger 2 is located. The multiple nozzles 41 are arranged side by side with spacing along the axial direction X of the refrigerant supply pipe 4 and are formed at positions corresponding to the second flow path 12. Furthermore, in the refrigerant supply pipe 4, in accordance with the static pressure, each of the multiple nozzles 41 is formed such that the cross-sectional area A decreases continuously or in groups of two or more nozzles 41 as it moves downstream (to the right in Figure 2) in the axial direction X of the refrigerant supply pipe 4. A detailed explanation of static pressure will be given later.

[0038] The refrigerant flowing through the refrigerant supply pipe 4 flows into the absorbent liquid supply unit 3 through the outlet 41, where it is mixed with the absorbent liquid.

[0039] An air-cooling method may be used to cool the absorber 1. In that case, fins are arranged in the first flow path 11. The fins are formed from materials known for use in fins of absorber 1 and have a known shape. The fins may be made of aluminum, for example.

[0040] Referring to Figures 3 and 4, the total pressure (Ptin), static pressure (Ps1, etc.), and dynamic pressure inside the refrigerant supply pipe 4 will be explained. Static pressure is the pressure when a fluid is at rest. Dynamic pressure is the kinetic energy of a fluid expressed in units of pressure. Total pressure is the sum of static pressure and dynamic pressure, and is always constant in frictionless flow inside the refrigerant supply pipe 4. Furthermore, inside the refrigerant supply pipe 4 with its downstream end closed, when the static pressure decreases, the dynamic pressure increases and the refrigerant flow accelerates, and when the static pressure increases, the dynamic pressure decreases and the refrigerant flow decelerates.

[0041] Here, the dynamic pressure is (1 / 2)*(ρ(density))*(V(flow velocity)) 2 This is represented by the following: Inside the refrigerant supply pipe 4, the refrigerant flow rate decreases as you move downstream in the axial direction X, so the refrigerant flow velocity gradually decreases. As a result, in the refrigerant supply pipe 4, the dynamic pressure tends to be lower and the static pressure to be higher as you move downstream (see Figure 3).

[0042] Furthermore, if the density of the refrigerant is constant inside the refrigerant supply pipe 4, and the cross-sectional area A of the refrigerant supply pipe 4 and the cross-sectional area A of the outlet 41 are constant in the axial direction X, the static pressure will be higher downstream. Therefore, the flow rate of refrigerant ejected from the outlet 41 will be higher downstream of the refrigerant supply pipe 4. For this reason, it is considered possible to adjust the refrigerant flow rate at each outlet 41 by changing at least one of the cross-sectional area A of the refrigerant supply pipe 4 and the cross-sectional area A of the outlet 41 in the axial direction X of the refrigerant supply pipe 4.

[0043] In Figure 4, the flow rate G of the refrigerant flowing inside the refrigerant supply pipe 4 with a constant cross-sectional area, the total pressure Ptin, the static pressure Ps1 of the first basin R1, the dynamic pressure Pv1 of the first basin R1, the static pressure Ps2 of the second basin R2, the dynamic pressure Pv2 of the second basin R2, the density ρ, the refrigerant discharge flow rate g from the outlet 41, and the static pressure Psout at the outlet of the refrigerant supply pipe 4 are shown, respectively. Furthermore, the cross-sectional area of ​​the first outlet 41-1 is denoted as A1, and the cross-sectional area of ​​the second outlet 41-2 is denoted as A2.

[0044] In the refrigerant supply pipe 4, the flow rate G1 of the refrigerant flowing through the first basin R1 on the outermost side, the static pressure Ps1 of the first basin R1, the dynamic pressure Pv1 of the first basin R1, and the refrigerant discharge flow rate g1 from the first outlet 41-1 can each be expressed by the following equations (1) to (3). Flow rate G1=G0-g0=G0 (1) Static pressure Ps1=Ptin-Pv1=Ptin-1 / 2*ρ*V1 2 (2) Refrigerant jet flow rate g1=C1*A1*{2*(Ps1-Psout)*ρ} 0.5 (3) Here, G0 represents the flow rate of the refrigerant flowing into the first basin R1, and g0 represents the flow rate of the refrigerant ejected before it flows into the first basin R1 (flow rate 0). C1 is the flow coefficient. The flow coefficient C1 is a correction factor used to represent the actual flow rate, taking into account losses due to contraction and friction, relative to the theoretical fluid velocity.

[0045] In the refrigerant supply pipe 4, the flow rate G2 of the refrigerant flowing through the second basin R2, which is downstream of the first basin R1, the static pressure Ps2 of the second basin R2, and the refrigerant ejection flow rate g2 from the second outlet 41-2 of the second basin R2 can each be expressed by the following equations (4) to (6). Flow rate G2=G1-g1 (4) Static pressure Ps2=Ptin-Pv2=Ptin-1 / 2*ρ*V2 2 (5) Refrigerant jet flow rate g2=C2*A2*{2*(Ps2-Psout)*ρ} 0.5 (6) Here, C2 is the flow coefficient.

[0046] (Calculation 1) As shown in Figure 5, a compression-absorption hybrid system 100A was used, in which the refrigerant expansion device 116, gas-liquid separator 130, and compressor 117 of the air conditioning system 100 in Figure 1 were omitted, and the compressor 123 was connected to the absorber 1. The cross-sectional area of ​​the refrigerant supply pipe 4 was kept constant, and the cross-sectional area of ​​the outlet, the flow rate of refrigerant discharged from the outlet, etc. were estimated. The operating conditions are as follows. Evaporation temperature: 5°C, Condensation / absorption temperature: 30°C, Superheating degree: 5K, Supercooling degree: Zero Compression efficiency: 75%, Pump efficiency: 50%, Solution heat exchange efficiency: 85% Regeneration heat Qg / Evaporation heat Qe = 1.5 Compression in the high-pressure stage: None, Expansion: One stage, Absorbent liquid: Organic solvent system (one of the absorbent liquids exemplified above that is suitable for this system is used) Number of refrigerant divisions (number of outlets in the axial direction) = 20, Number of outlets in the circumferential direction of the refrigerant supply pipe: 2 Outlet flow coefficient = 0.6 (fixed) Medium-pressure refrigerant: R1234ze(E)

[0047] Figures 6 and 7 show the settings related to the hole diameter of the nozzle 41 in Calculation 1. Figure 8 shows the calculation results of Calculation 1.

[0048] (Setting 1: Optimization) To ensure that the refrigerant flow rate (refrigerant discharge flow rate) from multiple outlets 41 arranged in parallel along the axial direction X of the refrigerant supply pipe 4 is uniform, the diameter of the holes (cross-sectional area A) of the outlets 41 are continuously reduced toward the downstream side in the axial direction X of the refrigerant supply pipe 4, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4. Specifically, as shown in Figure 6, the diameter of the holes of the outlets 41 is continuously reduced toward the downstream side in units of less than 0.1 mm. This allows for accurate suppression of fluctuations in the refrigerant discharge flow rate in the axial direction X of the refrigerant supply pipe 4.

[0049] (Setting 2: Uniform) The diameter of each hole (cross-sectional area A) of the multiple outlets 41, which are arranged in parallel toward the downstream side in the axial direction X of the refrigerant supply pipe 4, is made uniform.

[0050] (Setting 3: Machining adjustment, considering machining accuracy) The multiple outlets 41 are divided into multiple groups along the axial direction X of the refrigerant supply pipe 4, and the outlets 41 included in each group have the same cross-sectional area A. Furthermore, the cross-sectional area A of the outlets 41 included in each group decreases in a stepwise manner as the group moves downstream along the axial direction X of the refrigerant supply pipe 4. Specifically, as shown in Figure 6, the hole diameter of the outlet 41 is set to 3.8 mm at position 1-2, 3.7 mm at position 3-5, 3.6 mm at position 6-9, 3.5 mm at position 10-18, and 3.4 mm at position 19-20.

[0051] (Setting 4: Uniform (High allowable pressure loss)) The diameter of each of the multiple outlets 41 arranged in parallel toward the downstream side in the axial direction X of the refrigerant supply pipe 4 (cross-sectional area A) is made smaller than setting 2 and uniform.

[0052] (Calculation results for Calculation 1) As shown in Figure 8, in setting 1, the refrigerant flow rate ejected from the multiple outlets 41 located in the axial direction X of the refrigerant supply pipe 4 became uniform. Similarly, in setting 3, a tendency for the refrigerant flow rate ejected from the multiple outlets 41 located in the axial direction X of the refrigerant supply pipe 4 to become uniform was observed. On the other hand, in setting 2, the difference in refrigerant flow rate ejected from the outlets 41 between the upstream and downstream sides of the refrigerant supply pipe 4 becomes large, preventing the full utilization of the absorber 1's performance. Furthermore, in setting 4, although the difference in refrigerant flow rate ejected from the multiple outlets 41 between the upstream and downstream sides of the refrigerant supply pipe 4 is smaller than in setting 2, the pressure loss is greater than in settings 1-3, leading to a decrease in the overall system performance.

[0053] (Calculation 2) Using the same compression-absorption hybrid system 100A (Figure 5) as in Calculation 1, the cross-sectional area of ​​the refrigerant supply pipe 4 was kept constant, and the refrigerant was changed to calculate the cross-sectional area of ​​the nozzle, the refrigerant flow rate discharged from the nozzle, etc. The operating conditions were as follows: Evaporation temperature: 5°C, Condensation / absorption temperature: 30°C, Superheating degree: 5K, Supercooling degree: Zero Compression efficiency: 75%, Pump efficiency: 50%, Solution heat exchange efficiency: 85% Regeneration heat Qg / Evaporation heat Qe = 1.35 Compression in high-pressure section 110: None, Expansion: One stage, Absorbent liquid: Organic solvent system (one of the absorbent liquids exemplified above that is suitable for this system is used) Number of refrigerant divisions (number of outlets in the axial direction) = 20, Number of outlets in the circumferential direction of the refrigerant supply pipe: 2 Outlet flow coefficient = 0.6 (fixed) Low-pressure refrigerant: R1233zd(E)

[0054] Figures 9 and 10 show the settings related to the hole diameter (cross-sectional area) of the nozzle 41 in Calculation 2. Figure 11 shows the calculation results of Calculation 2.

[0055] (Setting 1: Optimization) The settings will be the same as in Calculation 1. However, the hole diameter (cross-sectional area) of nozzle 41 will be different from that in Calculation 1.

[0056] (Setting 2: Uniform) The settings will be the same as in Calculation 1. However, the hole diameter (cross-sectional area) of nozzle 41 will be different from that in Calculation 1.

[0057] (Setting 3: Machining adjustment, considering machining accuracy) The settings will be the same as in Calculation 1. However, the hole diameter (cross-sectional area) of the nozzle 41 will be different from that in Calculation 1. Specifically, as shown in Figure 9, the hole diameter of the nozzle 41 will be set to 4.8 mm at position 1-2, 4.7 mm at position 3-5, 4.6 mm at position 6-8, 4.5 mm at position 9-14, and 4.4 mm at position 15-20.

[0058] (Setting 4: Uniform (High allowable pressure loss)) The settings will be the same as in Calculation 1. However, the hole diameter (cross-sectional area) of nozzle 41 will be different from that in Calculation 1.

[0059] (Calculation results for Calculation 2) As shown in Figure 11, the results were almost the same as those of Calculation 1 shown in Figure 8.

[0060] [Effects of calculations 1 and 2] As shown in Setting 1, by continuously reducing the cross-sectional area A of the multiple outlets 41 toward the downstream side in the axial direction X of the refrigerant supply pipe 4 in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4, it was confirmed that the refrigerant flow rate ejected from the multiple outlets 41 of the refrigerant supply pipe 4 in the absorber 1 can be made uniform with reduced variation. As a result, the absorber 1 can suppress the pressure loss of the refrigerant at the inlet of the second flow path 12. Consequently, the absorber 1 can efficiently absorb the refrigerant with the absorbent liquid, improving the absorption characteristics of the absorber 1.

[0061] In order to efficiently suppress variations in the flow rate of the discharged refrigerant, it is preferable for the cross-sectional area of ​​the multiple outlets 41 provided in the refrigerant supply pipe 4 to be continuously reduced toward the downstream side of the refrigerant supply pipe 4. However, in that case, it is necessary to reduce the difference in cross-sectional area of ​​adjacent outlets 41 along the flow path of the refrigerant supply pipe 4, which requires high precision in processing the multiple outlets 41. On the other hand, according to setting 3, the cross-sectional area of ​​the multiple outlets 41 can be changed in groups, so the difference in cross-sectional area of ​​adjacent outlets 41 along the flow path of the refrigerant supply pipe 4 can be made larger than in setting 1. This makes it easier to process the multiple outlets 41, and thus the manufacturing cost of the absorber 1 can be reduced. In addition, since the cross-sectional area of ​​the multiple outlets 41 decreases in groups toward the downstream side of the refrigerant supply pipe 4, variations in the flow rate of the refrigerant discharged from the multiple outlets 41 can be suppressed to an acceptable degree, and the pressure loss of the refrigerant at the inlet of the second flow path 12 of the absorber 1 can be suppressed.

[0062] [Other Embodiments] Finally, other embodiments of the absorber and cooling system according to the present invention will be described. Note that the configurations disclosed in each of the following embodiments can be applied in combination with configurations disclosed in other embodiments, as long as this does not create a contradiction.

[0063] In the above embodiment, an example was described in which, in the absorber 1, the multiple outlets 41 are formed such that their cross-sectional area decreases as they move downstream of the refrigerant supply pipe 4, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4, which has a constant cross-sectional area. Alternatively, the absorber 1 may be formed such that the cross-sectional area of ​​the refrigerant supply pipe 4 decreases as it moves downstream, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4. As a result, the dynamic pressure tends to increase and the static pressure tends to decrease as it moves downstream in the flow direction of the refrigerant inside the refrigerant supply pipe 4, accelerating the flow of the refrigerant. As a result, even if the cross-sectional area of ​​the multiple outlets 41 in the absorber 1 is kept constant, it is considered that the refrigerant flow rate ejected from the multiple outlets 41 towards the second flow path 12 in the refrigerant supply pipe 4 will have reduced variation and become more uniform. Furthermore, the absorber 1 may be formed such that the cross-sectional area of ​​the refrigerant supply pipe 4 and the cross-sectional area of ​​the multiple outlets 41 decrease as they move downstream of the refrigerant supply pipe 4, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe 4.

[0064] In the above embodiment, an example was described in which the absorber 1 is provided in an air conditioning system 100 (refrigeration system) configured as a compression-absorption cycle that combines compression and absorption. However, the absorber according to the present invention may also be provided in a purely absorption cycle.

[0065] In the above embodiment, an example was shown in which two nozzles 41 are provided in the circumferential direction in the refrigerant supply pipe 4. However, there may be one nozzle 41 or three or more nozzles 41 provided in the circumferential direction. The circumferential positions of the multiple nozzles 41 provided in the circumferential direction in the refrigerant supply pipe 4 are not particularly limited and are only required to be positions from which refrigerant is supplied toward the second flow path 12.

[0066] In the above embodiment, a configuration in which the compressor 117 is not provided in the high-pressure stage 110 (Figure 5) was described as an example, but the cooling system according to the present invention may also have a compressor 117 (Figure 1). Furthermore, as shown in Figure 12, the configuration of the air conditioning system 100 in Figure 1 may be partially modified to create a cooling system 100B that supplies refrigerant from the compressor 123 to the absorber 1 without going through the gas-liquid separator 130.

[0067] In the above embodiment, an example was shown in which the refrigerant supply pipe 4, which is located inside the absorbent liquid supply section 3 and through which the refrigerant flows, is a cylindrical flow path in the absorber 1. However, the refrigerant supply pipe 4 is not limited to a cylindrical flow path; it may be a flow path other than a cylindrical one, as long as it is a flow path through which the refrigerant is supplied.

[0068] With regard to other configurations, the embodiments disclosed herein are illustrative in all respects, and it should be understood that the scope of the present invention is not limited thereto. Those skilled in the art will readily understand that modifications can be made as appropriate without departing from the spirit of the invention. Therefore, other embodiments modified without departing from the spirit of the invention are naturally included within the scope of the present invention. [Industrial applicability]

[0069] This invention can be used, for example, in a cooling system. [Explanation of Symbols]

[0070] 1: Absorber 2: Flat multi-hole heat transfer tube 2a: One end of a flattened multi-hole heat transfer tube 2b: Other end of the flattened multi-hole heat transfer tube 10: Plate 11: First channel 12: Second channel 3: Absorbent liquid supply unit 31: Piping 4: Refrigerant supply pipe 41: Spout 100: Air conditioning system (cooling and heating system) 110: High-voltage section 111: Pump 112: Solution vacuum device 113: Solution heat exchanger 114: Regenerator 115: Condenser 116: Refrigerant expansion device 117: Compressor 118: Motor 120: Low-pressure section 121: Refrigerant expansion device 122: Evaporator 123: Compressor 124: Motor 130: Gas-liquid separation equipment A: Cross-sectional area of ​​the nozzle X: Longitudinal direction (axial direction) of the refrigerant supply pipe

Claims

1. Multiple plates are arranged such that a first channel through which the heated fluid flows and a second channel through which the absorbent liquid and refrigerant flow are alternately formed between them, An absorbent liquid supply unit is positioned at one end of a plurality of plates and communicates with the second flow path with the fluid, The absorbent liquid supply unit is equipped with a refrigerant supply pipe through which the refrigerant flows, The refrigerant supply pipe is arranged in a direction perpendicular to the plurality of plates and has a plurality of nozzles for ejecting the refrigerant at positions corresponding to the second flow path, and is formed such that, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe, at least one of the cross-sectional area of ​​the refrigerant supply pipe and the cross-sectional area of ​​the plurality of nozzles decreases as one moves downstream of the refrigerant supply pipe.

2. The absorber according to claim 1, wherein the plurality of outlets are formed such that their cross-sectional area decreases as they move downstream of the refrigerant supply pipe, in accordance with the static pressure in the flow direction of the refrigerant inside the refrigerant supply pipe.

3. The multiple nozzles are divided into multiple groups facing downstream of the refrigerant supply pipe. The single or multiple nozzles included in the divided group have the same cross-sectional area, The absorber according to claim 2, wherein the plurality of nozzles are formed such that the cross-sectional area of ​​the nozzles included in the grouped nozzles decreases as they move downstream of the refrigerant supply pipe.

4. The absorber according to claim 1, wherein the refrigerant flowing through the refrigerant supply pipe is one of R1234yf, R1234ze(E), R1233zd(E), and R1224yd(Z).

5. A cold energy generation system that uses an absorber according to any one of claims 1 to 4, and is an absorption type, or a compression-absorption type that combines a compression type and an absorption type.

Images