Refrigeration splitters, heat exchanger assemblies and air conditioners

Abstract

This invention provides a refrigeration distributor, a heat exchanger assembly, and an air conditioner. The refrigeration distributor includes a main body and an inlet pipe. An inlet pipe assembly hole is formed at the inlet end of the main body, and multiple branch pipe holes are formed at its outlet end. A mixing chamber is formed within the accommodating cavity of the main body, with its opening facing the inlet pipe assembly hole and extending recessed towards the outlet end of the main body. When projected along the axial direction of the main body, the mixing chamber is located within the inner circumference of the multiple branch pipe holes. The end of the inlet pipe is connected to the inlet pipe assembly hole but does not extend into the mixing chamber. The inlet pipe includes a main body section and a throttling section located downstream of the main body section with a smaller inner diameter than the main body section. By controlling the relationship between the minimum inner diameter D1 of the throttling section and the outer diameter D0 of the main body section and constructing a refrigerant expansion model based on the area ratio S2 / S1, the refrigerant can be reflected and fully mixed into a dispersed state within the mixing chamber. Furthermore, the refrigerant can maintain this dispersed flow pattern throughout the distribution stroke after mixing, thereby greatly improving the uniformity of the distribution.

Description

Technical Field

[0001] This invention relates to the field of refrigerant distribution technology, and particularly to a refrigeration distributor, heat exchanger assembly, and air conditioner. Background Technology

[0002] In compression refrigeration systems, the evaporator is the heat exchange component for cold energy transfer. Its multi-branch structure can improve its heat exchange efficiency, and the performance of the evaporator directly affects the efficiency of the entire refrigeration system. After the liquid refrigerant is converted from a single liquid phase to a two-phase gas-liquid phase by a throttling device, the uniformity of the refrigerant gas-liquid two-phase flow distributed to each branch of the evaporator has a significant impact on the evaporator's heat exchange performance. Experiments have shown that the heat exchange capacity reduction caused by uneven refrigerant distribution can be as high as 25%. To improve the heat exchange efficiency of the evaporator, a distributor needs to be installed before the evaporator to distribute the refrigerant as evenly as possible to each branch of the evaporator. However, due to the limitations of the distributor's structure and its installation conditions, existing distributors generally suffer from poor distribution uniformity, and the uniformity is greatly affected by the installation angle. Only the Venturi distributor with its smooth linearity performs better. However, the smooth linearity also makes the Venturi distributor difficult to manufacture, and the consistency after manufacturing is also difficult to control.

[0003] To address this, some researchers have proposed a reflective distributor that achieves uniform mixing of the gas and liquid phases by reflecting the refrigerant through a reflective cavity. Examples include the refrigerant distributor structure proposed in Chinese patent CN 216204506 U and the air conditioner distributor proposed in Chinese patent CN111919067A. The reflective mixing of the refrigerant within the reflective cavity makes this distributor less affected by the installation angle, allowing it to be well-compatible with different installation angles. Although the reflective distributor outperforms traditional distributors in terms of installation angle, the vertical uniformity of liquid distribution in Chinese patent CN 216204506 U still lags behind that of traditional Venturi distributors, requiring further research to improve its refrigerant distribution performance. In contrast, Chinese patent CN111919067A describes a guide element inserted into a guide groove via a guide hole to form a gap S in the guide space. The refrigerant, after being reflected and redirected by the reversing space, flows into the relatively smaller cross-sectional area of ​​the guide space for refrigerant mixing. In this structure, the guide element inserted into the guide slot inevitably reduces the volume of the guide slot, resulting in insufficient mixing space for the reflected refrigerant. Meanwhile, the pressure inside the guide slot is very high, and the refrigerant, unable to mix in time, is distributed to each branch port through the gap S and distribution space, causing the gas and liquid phases to separate and severely affecting the uniformity of the distribution. Furthermore, the guide space formed by the gap S also leads to excessive refrigerant pressure loss, affecting compressor efficiency; simultaneously, excessive pressure loss also causes noise problems. Summary of the Invention

[0004] In order to overcome at least one deficiency of the prior art, the present invention provides a refrigeration splitter, a heat exchanger assembly, and an air conditioner.

[0005] To achieve the above objectives, the present invention provides a refrigeration distributor, which includes a body and an inlet pipe.

[0006] The body includes an inlet end, an outlet end, and a accommodating cavity connecting the inlet end and the outlet end respectively. An inlet pipe assembly hole is formed on the inlet end of the body, and multiple branch pipe holes for liquid distribution are formed on the outlet end. A mixing cavity is formed within the accommodating cavity of the body, with its opening facing the inlet pipe assembly hole and extending recessed towards the outlet end of the body. When projected along the axial direction of the body, the mixing cavity is located on the inner circumference of the multiple branch pipe holes. The end of the inlet pipe is connected to the inlet pipe assembly hole but does not extend into the mixing cavity. The inlet pipe includes a main body section with a substantially constant inner diameter and wall thickness, and a throttling section located downstream of the main body section with an inner diameter smaller than the outer diameter of the main body section.

[0007] Among them, the minimum inner diameter D1 of the throttling section and the outer diameter D0 of the main body section satisfy: 0.45D0≤D1≤0.95D0, and the cross-sectional area S2 of the inner cavity at the opening of the mixing chamber is greater than the flow cross-sectional area S1 at the minimum inner diameter of the throttling section, and S2 / S1 satisfies the following refrigerant expansion model:

[0008] (S2 / S1) min =-1.26×ln(D0)+4.572;

[0009] (S2 / S1) max =-1.75×ln(D0)+9.757;

[0010] Wherein, (S2 / S1) min The minimum value of S2 / S1 is (S2 / S1). max It is the maximum value of S2 / S1.

[0011] According to an embodiment of the present invention, the depth H1 of the mixing chamber satisfies: 2mm≤H1≤22mm, and the depth H1 of the mixing chamber refers to the vertical distance from the center of the bottom wall of the mixing chamber to the end face of the opening of the mixing chamber.

[0012] According to one embodiment of the present invention, the vertical distance H0 from the end face of the liquid outlet of the inlet pipe to the end face of the opening of the mixing chamber satisfies 3.5mm≤H0≤11.5mm.

[0013] According to one embodiment of the present invention, a gradually expanding section is formed on the inlet pipe, which is located downstream of the minimum inner diameter of the throttling section and has an arc-shaped generatrix on its outer wall. The outer wall of the gradually expanding section is welded to the inner wall of the inlet end of the body. The outlet end face of the inlet pipe is the inlet pipe cross-section at the downstream end of the inlet pipe assembly hole.

[0014] According to one embodiment of the present invention, the throttling section is a throttling orifice section with an inner diameter that remains substantially constant and is smaller than the outer diameter of the main body section;

[0015] Alternatively, the refrigerant distributor may also include a throttling orifice plate disposed within the main body section. The throttling orifice plate has a throttling orifice with an inner diameter smaller than the outer diameter of the main body section. A throttling section is formed where the throttling orifice is located. The throttling orifice is a straight orifice with a basically constant inner diameter or an inner diameter that gradually decreases along the flow direction of the refrigerant in the liquid inlet pipe.

[0016] Alternatively, the throttling section is a Venturi tube section, and the minimum inner diameter D1 of the throttling section is the inner diameter of the straight section at the throat of the Venturi tube section.

[0017] According to an embodiment of the present invention, the refrigerant distributor further includes multiple branch pipes welded and connected to multiple branch pipe holes respectively. Each branch pipe includes a first pipe section and a second pipe section located downstream of the first pipe section with an inner diameter that is smaller than the inner diameter at the downstream end of the first pipe section. The difference Δd between the inner diameter d11 at the downstream end of the first pipe section and the inner diameter d12 at the downstream end of the second pipe section is: 0.1mm≤Δd≤3.5mm. A reflective section is formed on the first pipe section, which bends and extends to one side relative to the center line of the branch pipe hole. The axis at the upstream end of the reflective section and the axis at the downstream end of the reflective section intersect to form an angle θ with 90°≤θ≤175°. Based on the reflective section, the axis of the second pipe section intersects the center line of the branch pipe hole.

[0018] According to an embodiment of the present invention, the refrigerant distributor further includes multiple branch pipes welded and connected to multiple branch pipe holes respectively. Each branch pipe includes a connecting straight section, a jet section located downstream of the connecting straight section with an inner diameter smaller than the inner diameter of the connecting straight section body, and a branch section located downstream of the jet section. An equal-diameter section with an inner diameter basically close to the inner diameter of the connecting straight section body is formed on the branch section. The difference Δd' between the inner diameter d11' of the connecting straight section body and the minimum inner diameter d12' at the jet section is: 0.1mm≤Δd'≤3.5mm.

[0019] According to an embodiment of the present invention, the refrigerant distributor further includes a partition plate disposed in the accommodating cavity of the body. The area of ​​the partition plate opposite to the liquid inlet pipe assembly hole is recessed towards the liquid outlet end of the body to form a mixing cavity. The partition plate plane on the outer periphery of the mixing cavity is formed with a plurality of partition plate holes distributed in a ring around the center line of the body. The plurality of partition plate holes distribute the refrigerant mixed in the mixing cavity to the liquid outlet end of the body.

[0020] According to one embodiment of the present invention, the partition divides the accommodating cavity into a first cavity near the liquid inlet end of the main body and a second cavity near the liquid outlet end of the main body. The second cavity is an annular cavity surrounding the mixing cavity, and the vertical distance H2 between the outer bottom wall of the mixing cavity and the inner bottom wall of the liquid outlet end of the main body satisfies: 0≤H2≤3mm.

[0021] According to one embodiment of the present invention, the mixing chamber is a constant-diameter chamber with a substantially uniform inner diameter; or, the mixing chamber is any one of a conical chamber, a frustum-shaped chamber, or a hemispherical chamber with an inner diameter that gradually decreases along the refrigerant inflow direction; or, the mixing chamber is a combination of multiple constant-diameter chambers, conical chambers, frustum-shaped chambers, or hemispherical chambers.

[0022] On the other hand, the present invention also provides a heat exchanger assembly including the above-mentioned refrigerant distributor.

[0023] On the other hand, the present invention also provides an air conditioner that includes the above-described heat exchanger assembly.

[0024] In summary, the refrigeration distributor provided by this invention achieves refrigerant pressure reduction and speed increase by setting a throttling section with a smaller inner diameter than the main body section on the inlet pipe. By controlling the inner diameter ratio of the two sections, sufficient mass flow rate at the main body inlet is ensured, enhancing the collision effect of the refrigerant in the mixing chamber and improving the mixing uniformity. Furthermore, controlling the minimum inner diameter of the throttling section effectively reduces pressure loss and noise generated by throttling. Simultaneously, setting the ratio of the inner cavity cross-sectional area S2 at the mixing chamber opening to the flow cross-sectional area S1 at the minimum inner diameter of the throttling section precisely controls the volume of the mixing chamber. This ensures that while providing expansion space and distribution channels for the refrigerant, the mixing chamber effectively avoids problems such as excessive refrigerant expansion and sudden drops in refrigerant velocity caused by an excessively large mixing chamber. It ensures that the input refrigerant has sufficient kinetic energy to be incident on the bottom wall of the mixing chamber to participate in reflective mixing, promoting the full atomization of the two-phase refrigerant into a dispersed state. It also ensures that the refrigerant maintains a dispersed flow during distribution, thereby improving the uniformity of the distribution. Furthermore, based on the refrigerant expansion model established by the fitting, after determining the outer diameter D0 and cross-sectional area S1 of the main body section, the inner cavity cross-sectional area S2 at the opening of the mixing cavity can be calculated to clarify the parameter selection of the mixing cavity before testing, providing guidance for the design before development, and thus greatly improving the development speed of the splitter product, especially the development of series products.

[0025] To make the above and other objects, features and advantages of the present invention more apparent and understandable, preferred embodiments are described below in detail with reference to the accompanying drawings. Attached Figure Description

[0026] Figure 1A The diagram shown is a schematic of an existing brass venturi structure shunt.

[0027] Figure 1B The diagram shows a commonly used voltage drop shunt.

[0028] Figure 2 The diagram shown is a schematic diagram of the refrigeration splitter provided in Embodiment 1 of the present invention.

[0029] Figure 3 As shown Figure 2 A schematic diagram of the structure after removing the inlet pipe and branch pipe.

[0030] Figure 4 The diagram shows the testing principle of the optimized physical sample performance testing equipment.

[0031] Figure 4A The figure shows the enthalpy change curve when the refrigerant flows through the electronic expansion valve.

[0032] Figure 5 As shown Figure 2 A schematic diagram of the structure of the partition plate.

[0033] Figure 6A , Figure 6B as well as Figure 6C The diagram shown is a structural schematic of the partition in another embodiment of the present invention.

[0034] Figure 7 The diagram shown is a schematic diagram of a refrigeration splitter provided in another embodiment of the present invention.

[0035] Figure 8A and Figure 8B The diagram shown is a schematic diagram of the inlet pipe in another embodiment of the present invention.

[0036] Figure 9 The diagram shown is a schematic diagram of a refrigeration distributor provided in another embodiment of the present invention.

[0037] Figure 10 The diagram shown is a schematic diagram of the refrigeration splitter provided in Embodiment 2 of the present invention.

[0038] Figure 11 The diagram shown is a schematic diagram of the refrigeration splitter provided in Embodiment 3 of the present invention.

[0039] Figure 12 The diagram shown is a schematic diagram of the branch pipe provided in Embodiment 4 of the present invention.

[0040] Figures 13A to 13C The diagram shown is a structural schematic of a branch pipe provided in another embodiment of the present invention.

[0041] Figure 14 The diagram shown is a schematic diagram of the refrigerant flow equalization distributor provided in Embodiment 5 of this utility model.

[0042] Figures 15A to 15E The diagram shown is a schematic diagram of the branch pipe in a refrigerant flow equalization distributor provided in another embodiment of this utility model. Detailed Implementation

[0043] Example 1

[0044] In traditional reflective splitters, high-pressure, high-speed two-phase refrigerant is reflected and collided within the reflective cavity to promote gas-liquid mixing. The reflection within the cavity makes this type of splitter less affected by gravity, thus exhibiting good splitting performance even when not vertically installed. However, when vertically installed, the performance of this type of splitter is difficult to meet the splitting standards of traditional brass Venturi splitters.

[0045] In view of this, such as Figure 2 As shown, this embodiment provides a refrigeration distributor, which includes a body 1, an inlet pipe 2, and multiple branch pipes 4. The body 1 includes an inlet end 101, an outlet end 102, and a receiving cavity 103 that connects the inlet end 101 and the outlet end 102 respectively. An inlet pipe assembly hole 104 is formed on the inlet end 101 of the body, and multiple branch pipe holes 105 for distributing liquid and connecting to the multiple branch pipes 4 are formed on the outlet end 102 of the body. A mixing cavity 106 is formed in the receiving cavity 103 of the body, with its opening facing the inlet pipe assembly hole 104 and extending recessed towards the outlet end 102 of the body. When projected along the axial direction of the body 1, the mixing cavity 106 is located on the inner periphery of the multiple branch pipe holes 105. The end of the inlet pipe 2 is connected to the inlet pipe assembly hole 104 and does not extend into the mixing cavity 106. The inlet pipe 2 includes a main body section 21 and a throttling section 22 located downstream of the main body section 21 with a smaller inner diameter than the main body section 21. Among them, the minimum inner diameter D1 of the throttling section 22 and the outer diameter D0 of the main body section 21 satisfy: 0.45D0≤D1≤0.95D0, and the inner cavity cross-sectional area S2 at the opening of the mixing chamber 106 is greater than the flow cross-sectional area S1 at the minimum inner diameter of the throttling section, and S2 / S1 satisfies the following refrigerant expansion model:

[0046] (S2 / S1) min =-1.26×ln(D0)+4.572;

[0047] (S2 / S1) max =-1.75×ln(D0)+9.757;

[0048] Wherein, (S2 / S1) min The minimum value of S2 / S1 is (S2 / S1). max It is the maximum value of S2 / S1.

[0049] In this embodiment, the refrigerant distributor further includes a partition 3 disposed within the main body accommodating cavity 103. The partition 3 divides the accommodating cavity 103 into a first cavity 1031 near the liquid inlet end of the main body and a second cavity 1032 near the liquid outlet end of the main body. The second cavity 1032 is an annular cavity surrounding the mixing cavity 106. The area on the partition 3 opposite to the liquid inlet pipe assembly hole 104 is recessed towards the side where the liquid outlet end of the main body is located to form the mixing cavity 106. Multiple partition holes 31 are formed on the partition plane around the outer periphery of the mixing cavity 106 in a ring-shaped distribution around the center line of the main body (i.e., when projected along the axial direction of the main body 1, the multiple partition holes 31 are located on the outer periphery of the mixing cavity 106). The multiple partition holes 31 distribute the refrigerant mixed in the mixing cavity 106 to the second cavity 1032, and then distribute it through the second cavity 1032 to multiple branch pipe holes 105 on the liquid outlet end 102 of the main body.

[0050] The high compatibility of reflective splitters with varying installation angles allows for their application in refrigeration equipment with limited installation space. However, their poor splitting performance in vertical installation significantly limits their application range. Current optimization efforts for reflective splitters primarily focus on the reflection angle and the insertion depth of the splitter branch pipes, but these optimizations offer very limited improvement in splitting uniformity. To address the issue of poor splitting performance of reflective splitters in vertical installation, the inventors conducted extensive analytical experiments.

[0051] A brass venturi shunt with superior flow uniformity in a vertically installed state (hereinafter referred to as a venturi shunt) was used as a control sample (its structure is as follows). Figure 1A As shown), the structure of the refrigeration distributor proposed by the inventor in Chinese CN216204506 U (hereinafter referred to as the distributor sample to be analyzed) and another commonly used pressure drop type distributor on the market (refer to the sample, whose structure is as follows) are compared. Figure 1B As shown, CFD (Computational Fluid Dynamics) simulation analysis was performed.

[0052] Sample structural parameters: The outer diameter of the main body of the inlet pipe of the control sample, reference sample and the sample to be analyzed is 16 mm, and the number of holes in the branch pipe on the outlet side is 11. The minimum inner diameter of the throttling section of the sample to be analyzed is 6 mm, the diameter at the opening of the mixing chamber is 18.8 mm, and the depth of the mixing chamber H1 is 12 mm.

[0053] The CFD (Computational Fluid Dynamics) simulation analysis uses ANSYS software, and its conditions are set as follows:

[0054] Turbulence model: Realizable k-ε model;

[0055] Network parameters: Tetrahedral unstructured mesh is used;

[0056] Operating conditions: The working fluid is R410A refrigerant, the wall boundary is an insulating boundary, the inlet dryness is 0.2, and the installation condition is vertical installation;

[0057] Inlet boundary conditions: Total inlet mass flow rate: 440 kg / h; Inlet gas velocity: 4.57 m / s; Inlet liquid velocity: 0.53 m / s; Liquid volume fraction: 0.1032; Turbulence intensity: 5%; Hydraulic diameter: 14.4 mm.

[0058] Simulation analysis revealed the following: the flow uniformity (STD0) of the Venturi splitter (as a control sample) was 0.1839 g / s; the flow uniformity (STD') of the pressure drop type splitter (reference sample) was 0.4650 g / s; and the flow uniformity (STD01) of the splitter under analysis was 0.4452 g / s. The evaluation index for flow uniformity is the standard deviation (STD) of the refrigerant mass flow rate at each branch outlet, expressed as follows:

[0059]

[0060] in, The average refrigerant mass flow rate at all branch outlets, m j Let n be the refrigerant mass flow rate at the outlet of the j-th branch pipe, and n be the number of branch pipes.

[0061] The simulation analysis above shows that the flow uniformity STD01 of the liquid separator sample to be analyzed under vertical installation is slightly better than the flow uniformity STD' of the commonly used pressure drop type liquid separator (reference sample). Therefore, the liquid separator sample to be analyzed has market application value; however, its flow uniformity is worse than that of the Venturi liquid separator (reference sample).

[0062] Based on simulation analysis, observations of the refrigerant flow pattern within the sample of the liquid separator revealed that: the refrigerant velocity and flow rate at the bottom wall of the mixing chamber were very low, indicating infrequent mixing and collision of the gas-liquid two-phase refrigerant; furthermore, after reflection mixing, gas-liquid phase separation occurred again during the refrigerant's return distribution along the mixing chamber to the outlet side, and even some liquid refrigerant deposited on the bottom wall of the inlet end of the main body. These observations suggest that the failure of some refrigerant to participate in reflection and the subsequent re-separation of the gas-liquid phases during distribution are the main factors affecting the uniformity of liquid distribution in the sample of the liquid separator.

[0063] In view of this, the inventor based on Figure 2The refrigeration distributor design provided in this embodiment is optimized as follows: Six optimization groups are selected based on the outer diameter D0 of the main body section of the inlet pipe. Each optimization group contains six sub-optimization groups with different minimum inner diameters D1 of the throttling section. Within each sub-optimization group, eight optimization samples are formed based on the inner diameter D2 of the mixing chamber opening, resulting in a total of 288 optimization samples. To clearly illustrate the structure of each component, see attached... Figure 2 This is a schematic diagram of the optimized sample structure with four branches. The structures of other optimized samples are similar. Figure 2 They are basically the same, the difference being the number of branches.

[0064] CFD simulation analysis was performed on 288 optimized samples using ANSYS software. The Venturi splitter shown in Figure 1 was used as the control sample, and the simulation analysis results of all optimized samples were compared. Specifically, due to the different inner diameter D0 of the main body section of the inlet pipe, each optimization group had a corresponding control sample to ensure consistent structural parameters; that is, the outer diameter D0 of the main body section of the inlet pipe and the number of branch pipes of each control sample were the same as the 48 optimized samples in that sub-optimization group. Optimized samples with a split uniformity STD less than or equal to the split uniformity STD0 of the control sample were considered qualified optimized samples; otherwise, they were considered unqualified optimized samples.

[0065] The turbulence model, network parameters, and operating conditions selected for the CFD simulation analysis are the same as those described above.

[0066] The boundary conditions for each optimization group are shown in the table below:

[0067]

[0068] In the first optimization group: the outer diameter of the main section of the inlet pipe is D0 = 6.35 mm, there are two branch pipes, and the inner diameter of the branch pipe is D3 = 3 mm; in the sub-optimization groups included in this optimization group, the minimum inner diameter of the throttling section is D1 = 2.1 mm, 2.9 mm, 3.7 mm, 4.5 mm, 6.0 mm, and 7.5 mm; in each sub-optimization group, eight optimized samples are selected according to the inner diameter D2 at the opening of the mixing chamber; for example, in the sub-optimization group where D1 = 2.1 mm, D2 = 1.3 mm, 1.8 mm, 2.3 mm, 2.8 mm, 3.3 mm, 3.8 mm, 4.3 mm, and 4.8 mm.

[0069] In the second optimization group: the outer diameter of the main section of the inlet pipe is D0=9.52mm, the number of branch pipes is four, and the inner diameter of the branch pipe is D3=3mm; the minimum inner diameter of the throttling section in the sub-optimization groups included in this optimization group is D1=2.7mm, 4.3mm, 5.91mm, 7.5mm, 9.0mm, 10.6mm; in each sub-optimization group, eight optimized samples are selected according to the inner diameter D2 at the opening of the mixing chamber.

[0070] In the third optimization group: the outer diameter of the main section of the inlet pipe is D0=12.7mm, the number of branch pipes is eight, and the inner diameter of the branch pipe is D3=3mm; in the sub-optimization groups included in this optimization group, the minimum inner diameter of the throttling section is D1=3.3mm, 5.7mm, 8.1mm, 10.5mm, 12.1mm, and 13.6mm; in each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 at the opening of the mixing chamber.

[0071] In the fourth optimization group: the outer diameter of the main section of the inlet pipe is D0=16mm, the number of branch pipes is eleven, and the inner diameter of the branch pipe is D3=3mm; the minimum inner diameter of the throttling section in the sub-optimization groups included in this optimization group is D1=4.5mm, 7.2mm, 9.9mm, 12.5mm, 15.2mm, 16mm; in each sub-optimization group, eight optimized samples are selected according to the inner diameter D2 at the opening of the mixing chamber.

[0072] In the fifth optimization group: the outer diameter of the main section of the inlet pipe is D0=22mm, the number of branch pipes is twenty-two, and the inner diameter of the branch pipe is D3=3mm; the minimum inner diameter of the throttling section in the sub-optimization groups included in this optimization group is D1=8.2mm, 9.9mm, 13.6mm, 17.2mm, 20.9mm, and 22.6mm; eight optimized samples are selected in each sub-optimization group based on the inner diameter D2 at the opening of the mixing chamber.

[0073] In the sixth optimization group: the outer diameter of the main section of the inlet pipe is D0=28mm, the number of branch pipes is twenty-four, and the inner diameter of the branch pipe is D3=3mm; the minimum inner diameter of the throttling section in the sub-optimization groups included in this optimization group is D1=10.3mm, 12.6mm, 17.3mm, 21.9mm, 26.6mm, 28.9mm; eight optimized samples are selected in each sub-optimization group based on the inner diameter D2 at the opening of the mixing chamber.

[0074] The simulation data of 288 optimized samples are summarized in Appendix Table 1.

[0075] To investigate the effects of the outer diameter D0 of the main body of the inlet pipe, the minimum inner diameter D1 of the throttling section, and the inner diameter D2 of the mixing chamber opening on the uniformity of the flow distribution, the data in Appendix 1 were analyzed.

[0076] First, the influence of the minimum inner diameter D1 of the throttling section on the flow uniformity within each optimization group was observed. It was found that after the minimum inner diameter D1 of the throttling section exceeded a certain range, regardless of how the inner diameter D2 at the opening of the mixing chamber changed, the flow uniformity of the optimized samples was worse than that of the control sample, as shown in the data of the first and sixth optimization groups. In the refrigeration splitter provided in this embodiment, the purpose of setting the throttling section 22 is to reduce the pressure and increase the speed of the refrigerant input into the main section 21, hoping that the refrigerant can be injected into the inner bottom wall of the mixing chamber 106 at a higher speed for reflection. However, an excessively small minimum inner diameter D1 of the throttling section will significantly reduce the input refrigerant mass flow rate; the smaller the refrigerant mass flow rate, the weaker its reflection and collision on the inner bottom wall of the mixing chamber 106 will be, thus seriously affecting the mixing intensity of the gas-liquid two-phase refrigerant. In other words, although throttling can increase the refrigerant flow rate to improve the intensity of reflective mixing, an excessively small minimum inner diameter D1 of the throttling section will also result in an excessively low refrigerant mass flow rate, which will weaken the mixing degree of the two-phase flow. In addition, an excessively small minimum inner diameter D1 of the throttling section will also lead to excessive refrigerant pressure loss, which will exacerbate the separation of the gas and liquid phases.

[0077] Therefore, further analysis of the relationship between the minimum inner diameter D1 of the throttling section and the assembly hole D0 of the inlet pipe within each optimization group reveals that the flow uniformity of the optimized sample is superior to that of the control sample only when the minimum inner diameter D1 of the throttling section within the sub-optimization group satisfies 0.45D0≤D1≤0.95D0. Preferably, D1 / D0 can be set to ratios of 0.5, 0.55, 0.6, 0.65, 0.7, 0.78, etc. However, this invention does not impose any limitations on this. In other embodiments, D1 / D0 can also be other ratios within the range of 0.45 to 0.95.

[0078] Secondly, the influence of the inner diameter D2 at the opening of the mixing chamber in each sub-optimization group on the uniformity of the flow distribution is analyzed by fixing the minimum inner diameter D1 of the throttling section. In the refrigeration splitter provided in this embodiment, after the high-pressure and high-speed refrigerant enters the cross-sectionally enlarged mixing chamber 106 from the throttling section 22, the refrigerant will inevitably expand and decrease in speed due to inertia. The decrease in flow velocity will affect the kinetic energy of the refrigerant. If the refrigerant flow velocity is too low, some refrigerant will not be able to enter the bottom wall of the mixing chamber 106 to participate in reflection, and the degree of gas-liquid two-phase mixing will not be intense, resulting in the two-phase flow being difficult to atomize and flowing in a diffuse manner. Furthermore, in two-phase refrigerants, the flow of the liquid phase mainly relies on inertial force and gravity. If the refrigerant flow rate is too low, the inertial force of the liquid phase will be less than its gravity. After mixing, the liquid phase in the refrigerant is easily affected by gravity and separates from the gas phase again. This leads to some liquid refrigerant gradually depositing on the inner wall of the liquid inlet end of the body during the distribution process after mixing, and thus cannot be distributed to the liquid outlet side, thereby affecting the uniformity of the distribution.

[0079] As the above analysis shows, both the expansion and deceleration of the refrigerant are caused by changes in the cross-sectional area of ​​the refrigerant flow channel. Therefore, the relationship between the cross-sectional area S2 at the opening of the mixing chamber 106 and the flow cross-sectional area S1 at the downstream end of the throttling section 22 in each sub-optimization group in Appendix 1 is analyzed. The analysis reveals that when the ratio of the cross-sectional area S2 at the opening of the mixing chamber 106 to the flow cross-sectional area S1 at the downstream end of the throttling section 22 (hereinafter referred to as the cross-sectional area ratio S2 / S1 for ease of description) is within a certain range, the flow uniformity STD of the optimized sample is significantly better than that of the control sample. However, when the cross-sectional area S2 / S1 is too large or too small, the flow uniformity decreases.

[0080] Appendix 2 summarizes the upper and lower limits of the cross-sectional area ratio S2 / S1 within each sub-optimization group in Appendix 1. Further analysis of the data in Appendix 2 reveals that the upper limit of the cross-sectional area ratio S2 / S1 within each optimization group is related to the inlet pipe assembly hole D0 within that group. Regarding the lower limit, due to the constraint that S2 / S1 > 1, analysis of optimization groups with D0 ranging from 6.35mm to 16mm shows that the lower limit of the cross-sectional area ratio S2 / S1 is also related to the inlet pipe assembly hole D0. Based on this analysis, the following refrigerant expansion model is constructed after fitting the data in Appendix 2:

[0081] (S2 / S1) min =-1.26×ln(D0)+4.572;

[0082] (S2 / S1) max =-1.75×ln(D0)+9.757;

[0083] Wherein, (S2 / S1) min The minimum value of the cross-sectional area ratio S2 / S1 is (S2 / S1). max This represents the maximum value of the cross-sectional area ratio S2 / S1.

[0084] The establishment of the refrigerant expansion model allows the refrigeration distributor provided in this embodiment to precisely set the range of the cross-sectional area ratio S2 / S1 based on the outer diameter D0 of the main body section of the inlet pipe. The upper limit of the cross-sectional area ratio S2 / S1 limits the degree of volume expansion and deceleration of the refrigerant after entering the mixing chamber 106, ensuring that the input refrigerant still has sufficient kinetic energy to be incident on the bottom wall of the mixing chamber 106 after expansion and deceleration, thus accelerating the collision between the gas and liquid phases and causing the two phases of refrigerant to atomize into a diffuse flow pattern. At the same time, the high refrigerant flow rate also makes the inertial force of the liquid phase refrigerant after mixing greater than its gravity. During the distribution process after mixing, the liquid phase refrigerant is not easily affected by gravity and will not aggregate and deposit, ensuring that the two phases of refrigerant can always maintain a uniformly mixed diffuse flow pattern during the distribution process after mixing, thereby achieving uniform distribution of refrigerant.

[0085] The lower limit of the cross-sectional area ratio S2 / S1 determines the minimum space required for refrigerant expansion and mixing, as well as the minimum flow channel cross-sectional area required for refrigerant distribution after mixing. Specifically, the mixing chamber 106 provides collision space for the two-phase refrigerant after reflection. If this space is too small (especially in product specifications with a small outer diameter of the main body section of the inlet pipe), the two-phase refrigerant will not have enough time to mix fully before being rapidly distributed into the multiple baffle holes 31 under the action of high pressure in the chamber. Insufficient refrigerant mixing will affect the uniformity of flow distribution. Furthermore, the high-pressure, high-speed refrigerant incident into the mixing chamber 106 will occupy part of the cross-sectional area at the opening of the mixing chamber 106. An excessively small cross-sectional area S2 at the opening of the mixing chamber 106 will make it difficult for the reflected and mixed refrigerant to be distributed into the multiple baffle holes 31 through this area, that is, the effective distribution flow channel at the opening of the mixing chamber 106 is too narrow. An excessively narrow distribution flow channel will cause a sharp increase in refrigerant flow resistance. The increase in flow resistance will affect the refrigerant flow velocity input through the inlet pipe assembly hole 104, thereby affecting the refrigerant reflection and mixing effect. On the other hand, increased flow resistance can also lead to excessive pressure loss in the distributor, thereby affecting the performance of the heat exchanger components connected downstream of the distributor. Specifically, excessive pressure loss in the distributor results in an excessively low average evaporation temperature in the evaporator. With the condensation temperature remaining constant, the decrease in the average evaporation temperature leads to an increase in the compressor's pressure ratio, which in turn increases the compressor's power. Therefore, the refrigeration distributor provided in this embodiment determines the lower limit of the cross-sectional area ratio S2 / S1 using a refrigerant expansion model, thereby determining the minimum diameter D2 at the opening of the mixing chamber 106. min This provides a suitable mixing space for the refrigerant after reflection, while controlling the refrigerant distribution channel, thereby improving the uniformity of refrigerant distribution.

[0086] After obtaining the simulation analysis data of the optimized samples, to verify the reliability of the simulation analysis data, corresponding optimized physical samples were fabricated based on the structural parameters of some optimized samples. These optimized physical samples were then subjected to on-machine testing to determine their actual flow uniformity (STD) test. Specifically, three sub-optimization groups were selected within each optimization group, and three optimized samples were selected within the specifications of each sub-optimization group for physical fabrication and on-machine performance testing. A total of six physical test groups were conducted, each consisting of three sub-physical test groups and each sub-physical test group containing three optimized physical samples, for a total of 54 optimized physical samples. Similarly, the control sample corresponding to each optimization group was used as the control sample for the on-machine performance test.

[0087] The flow uniformity of 54 optimized physical samples was tested using a flow divider performance testing device. Figure 4The diagram shows the schematic of the flow divider performance testing equipment used in this embodiment. This equipment includes multiple ball valves 61, twenty-four sets of finned evaporators 62, a one-way valve 63, an L-shaped finned evaporator 64, a four-way reversing valve 65, a compressor 66, an economizer 67, an economizer electronic expansion valve 68, an evaporator electronic expansion valve main valve 69, an evaporator electronic expansion valve auxiliary valve 70, a gas-liquid flow divider 71, a shell-and-tube evaporator 72, a liquid receiver 73, a total outlet water temperature sensor 74, a total return water temperature sensor 75, an insulated flushing water tank 76, and a cooling tower 77. The evaporator electronic expansion valve main valve 69 and the evaporator electronic expansion valve auxiliary valve 70 are connected in parallel to the inlet pipe 2 of the optimized physical sample 80. Multiple branch pipes 4 of the optimized physical sample 80 are connected to all or part of the finned evaporators 62. The testing principle of this performance testing equipment is as follows:

[0088] [1] Equipment structure and layout concept

[0089] This distributor performance testing equipment consists of twenty-four loop evaporators, each with a finned structure (i.e., a finned evaporator). The heat exchange area, heat exchange tubes, and fin arrangement of each loop evaporator are completely identical, and each loop is uniformly arranged around the fan in a 360° surround. Therefore, it is assumed that the airflow through each loop evaporator is consistent, and similarly, the heat exchange capacity of each loop evaporator can be roughly considered consistent.

[0090] [2] Test theoretical calculation

[0091] The heat exchange capacity of an evaporator can be calculated from the enthalpy values ​​of the fluids at the inlet and outlet of the evaporator. The theoretical formula is: Heat exchange capacity = Refrigerant mass flow rate × (Outlet enthalpy - Inlet enthalpy). Based on this, the difference in refrigerant mass flow rate in each loop of the evaporator can be deduced, the percentage difference can be obtained, and the theoretical compliance of the test element can be judged.

[0092] The enthalpy value of a single-phase refrigerant can be obtained by looking up the temperature and pressure in a table.

[0093] Since the refrigerant at the evaporator inlet is in a two-phase gas-liquid state, its enthalpy cannot be calculated. This problem can be solved by setting up throttling devices. This test equipment has two electronic expansion valves (throttling devices), including the main evaporator electronic expansion valve 69 and the auxiliary evaporator electronic expansion valve 70. Utilizing the isoenthalpic throttling characteristic of the throttling devices, the enthalpy value before the two electronic expansion valves (throttling devices) is measured to be equivalent to the enthalpy value at the evaporator inlet. Assuming that the refrigerant before the two electronic expansion valves (throttling devices) is a pure liquid (subcooled liquid), the enthalpy value can be obtained by referring to the corresponding refrigerant property table using the temperature before the two electronic expansion valves (throttling devices).

[0094] Assuming that the evaporator outlet of each loop is filled with superheated gaseous refrigerant, its enthalpy can be obtained by referring to the property table of the corresponding refrigerant based on the outlet refrigerant temperature and pressure.

[0095] [3] Isoenthalpic throttling

[0096] like Figure 4A As shown, the horizontal axis h represents the process enthalpy, and the vertical axis lg(P) represents the logarithmic pressure value. The 4→1 process is the enthalpy change curve when the refrigerant flows through the electronic expansion valve (throttling component). During the process of flowing through the throttling component, there is no heat exchange with the outside world, and the refrigerant does not exchange heat with any fluid. Therefore, it can be seen that the 4→1 process is perpendicular to the horizontal axis, and the enthalpy value does not change during the throttling process.

[0097] [4] Enthalpy Value Inquiry

[0098] The evaporator inlet and outlet temperatures are measured experimentally using the total outlet water temperature sensor 74 and the total return water temperature sensor 75. The corresponding saturation pressure can be looked up using enthalpy lookup software (such as a refrigerant calculator) to determine the subcooling (heating) state of the inlet and outlet refrigerant. By inputting the measured inlet and outlet temperature and pressure values, the corresponding enthalpy values ​​can be obtained. Based on the theoretical calculation formula mentioned above, the difference in refrigerant mass flow rate in each loop of the evaporator can be deduced.

[0099] The data from 54 optimized physical samples after testing on the distributor performance testing equipment are summarized in Appendix Table 3. In Appendix Table 3, the distribution uniformity of the optimized physical samples is represented by the STD test, and the distribution uniformity of the control physical samples is represented by the STD0 test. In addition, in the physical test group with an outer diameter D=16mm of the main body of the inlet pipe, the STD01 test data of the optimized physical sample of the refrigeration distributor structure proposed by the inventor in Chinese CN216204506 U (i.e., the physical sample of the distributor to be analyzed) is also summarized as a reference.

[0100] From Appendix 3, we can obtain:

[0101] ① In each physical test group, D1 of the first sub-physical test group is less than 0.45D0. The split uniformity STD test of all optimized physical samples in this sub-test is worse than the split uniformity STD0 test of the control physical sample.

[0102] ② For the sub-physical test group (including the second and third sub-physical test groups) with 0.45D0≤D1≤0.95D0, the optimized physical sample with S2>S1 and area ratio S2 / S1 satisfying the refrigerant expansion model can outperform the corresponding control physical sample in the STD0 test of the split uniformity test.

[0103] ③ In the physical test group with D0=16mm, the optimized physical sample, which has better split uniformity than the control physical sample, also has a much better split uniformity STD test than the original sample of the liquid separator to be analyzed, which has a better split uniformity STD01 test.

[0104] The test results of 54 optimized physical samples were consistent with the simulation analysis results, indicating that the use of ANSYS simulation analysis software to optimize samples and establish a refrigerant expansion model in this embodiment is reliable.

[0105] The performance test results of the optimized physical samples verified the reliability of the simulation analysis. Based on this, the fitted refrigerant expansion model needs to be validated to determine its reliability under different outer diameters D0 of the main body of the inlet pipe. Specifically, an outer diameter D0 = 19 mm was selected for the main body of the inlet pipe. Based on the expansion coefficients k = -1.26 and b = 4.572, the lower limit of the cross-sectional area ratio S2 / S1 (S2 / S1)min = 0.86 was calculated; the upper limit of S2 / S1 (S2 / S1)max = 4.60 was calculated based on the expansion coefficients k = -1.75 and b = 9.757. Since the cross-sectional area ratio S2 / S1 > 1 must also be satisfied, qualified verification samples were designed within the range of 1 < S2 / S1 ≤ 4.60, and unqualified verification samples were designed within the range of S2 / S1 > 4.60 to form the first verification group.

[0106] Similarly, selecting an outer diameter D0 of 32mm for the main body of the inlet pipe, and calculating the lower limit (S2 / S1)min = 0.21 based on the expansion coefficients k = -1.26 and b = 4.572, and the upper limit (S2 / S1)max = 3.69 based on the expansion coefficients k = -1.75 and b = 9.757, the lower limit (S2 / S1)max = 3.69. Since the cross-sectional area ratio S2 / S1 > 1 must also be satisfied, qualified verification samples are designed within the range of 1 < S2 / S1 ≤ 3.69, and unqualified verification samples are designed within the range of S2 / S1 > 3.69 to form the second verification group.

[0107] In the first verification group, the outer diameter of the main body of the inlet pipe is D0=19mm. The minimum inner diameter of the throttling section in the sub-verification groups is D1=8.6mm, 10.5mm, 12.4mm, 14.3mm, 16.2mm, and 18.1mm. Eight verification samples are selected in each sub-verification group, of which six are qualified verification samples with an area ratio of 1<S2 / S1≤4.60; the other two are unqualified verification samples with an area ratio of S2 / S1>4.60.

[0108] In the second verification group, the outer diameter of the main body of the inlet pipe is D0=32mm. The minimum inner diameter of the throttling section in the sub-verification groups is D1=14.4mm, 17.6mm, 20.8mm, 24.0mm, 27.2mm, and 30.9mm. Eight verification samples are selected from each sub-verification group. Six of them are qualified verification samples, and their area ratio satisfies 1<S2 / S1≤3.69. The other two are unqualified verification samples, and their area ratio satisfies S2 / S1>3.69.

[0109] The 96 verification samples in the two verification groups were simulated and analyzed using ANSYS simulation analysis software. The turbulence model, network parameters, and operating conditions selected for the simulation analysis were the same as those described above.

[0110] The boundary conditions for each validation group are shown in the table below:

[0111]

[0112] The simulation data of 96 optimized samples are summarized in Appendix Table 4.

[0113] Similarly, simulation analysis data of a Venturi splitter with identical inlet pipe inner diameter D0, branch pipe number, and branch pipe specifications were used as a control sample for comparative analysis of all verification samples within each verification group. The comparative analysis showed that the split uniformity STD of the six qualified verification samples in each sub-verification group was better than the split uniformity STD0 of the corresponding control sample, while the split uniformity STD of the two unqualified verification samples was worse than the split uniformity STD0 of the control sample. The verification data in Appendix Table 4 reflects the reliability of the refrigerant expansion model fitted and established in this embodiment.

[0114] The establishment of a high-confidence refrigerant expansion model enables the refrigeration distributor provided in this embodiment to accurately calculate the cross-sectional area ratio S2 / S1 based on a given outer diameter D0 of the main body of the inlet pipe during product design. Simultaneously, by combining the relationship between the outer diameter D0 of the main body of the inlet pipe and the minimum inner diameter D1 of the throttling section described above, the cross-sectional area S2 at the opening of the mixing chamber with excellent flow uniformity can be calculated, thereby obtaining the inner diameter D2 at the opening of the mixing chamber required for product design and trial production. This provides selection guidance for product series design, significantly shortening the product development cycle and reducing R&D costs for refrigeration distributors. Preferably, after determining the upper limit (S2 / S1)max and lower limit (S2 / S1)min, and provided the space within the accommodating cavity 103 allows, the area ratio S2 / S1 can be set to be closer to the upper limit (S2 / S1)max.

[0115] In this embodiment, the mixing chamber depth H1 satisfies: 2mm ≤ H1 ≤ 22mm. The depth H1 of the mixing chamber 106 refers to the vertical distance from the end face of the mixing chamber opening to the center of the bottom wall of the mixing chamber. Preferably, the depth H1 of the mixing chamber 106 is set to any one of 8mm, 10mm, 12mm, 14mm, 16mm, 18mm, and 20mm. However, the present invention does not impose any limitation on this. In other embodiments, the depth H1 of the mixing chamber can also be other values ​​within the range of 2mm ≤ H1 ≤ 22mm. Setting the mixing chamber depth H1 enables control of the internal volume of the mixing chamber 106 to further promote the mixing degree of the two-phase refrigerant.

[0116] In this embodiment, as Figures 2 to 5 As shown, the mixing chamber 106 is a cylindrical chamber with a substantially uniform inner diameter. However, the present invention does not limit this. In other embodiments, the inner shape of the mixing chamber 106 may also be a frustum-shaped cone (e.g., Figure 6A As shown), conical (such as) Figure 6B (as shown), any one or more combinations of prisms, frustums, or pyramids; or a combination of the above-mentioned internal cavity shapes and partial spherical shapes, such as a combination of cylindrical and partial spherical shapes (e.g. Figure 6C (as shown), combinations of frustums and partial spheres, combinations of prisms and partial spheres, etc.

[0117] In this embodiment, as Figure 2 As shown, the vertical distance H0 between the outlet end face of the inlet pipe 2 and the opening end face of the mixing chamber 106 satisfies 3.5mm ≤ H0 ≤ 11.5mm. Preferably, the vertical distance H0 is 5mm or 8mm. However, the present invention does not limit this. In other embodiments, the vertical distance H0 can also be other values ​​within the range of 3.5mm ≤ H0 ≤ 11.5mm. The vertical distance H0 affects the distribution channel from the opening of the mixing chamber 106 to the partition hole 31. An excessively small vertical distance H0 will cause the distribution channel to be too narrow, increasing the flow resistance of the refrigerant entering the partition hole 31. In particular, when the outlet end of the inlet pipe 2 extends into the body 1, the refrigerant forms a vortex between the outer wall of the inlet end of the inlet pipe 2 and the inner wall of the inlet end 101 of the body. This vortex will further compress the distribution channel, resulting in a sharp increase in flow resistance. In view of this, this embodiment sets a vertical distance H0 ≥ 3.5 mm to increase the distribution flow channel from the opening of the mixing chamber 106 to the partition hole 31, thereby facilitating the smooth distribution of the mixed refrigerant into the multiple partition holes 31 and reducing pressure loss during distribution. Furthermore, an excessively large vertical distance H0 will also increase the travel distance of the refrigerant sprayed onto the bottom wall of the mixing chamber 106, resulting in insufficient kinetic energy for reflection and thus worsening the degree of reflection mixing. Therefore, it is necessary to control the upper limit of the vertical distance H0 to ensure that the refrigerant has sufficient kinetic energy to impact the bottom wall of the mixing chamber after entering the body 1 for reflection mixing.

[0118] In this embodiment, the throttling section 22 is a throttling orifice section with a substantially constant inner diameter that is smaller than the inner diameter of the main body section 21. The minimum inner diameter D1 of the throttling section is the inner diameter of the throttling orifice section. Specifically, in this embodiment, the throttling section 22 is integrally formed at the downstream end of the main body section 21. However, the present invention does not impose any limitations on this. In other embodiments, such as Figure 7 As shown, a refrigerant distributor can also be provided, including a throttling orifice plate 2A disposed within the main body section 21. The throttling orifice plate 2A has a throttling orifice 2A1 with an inner diameter smaller than the outer diameter of the main body section, forming a throttling section 22 at the location of the throttling orifice 2A1. The throttling orifice can be a straight orifice with a substantially constant inner diameter or a collecting orifice with an inner diameter that gradually decreases along the flow direction of the refrigerant in the inlet pipe. In this structure, the inner diameter of the downstream end of the throttling orifice 2A1 is the minimum inner diameter D1 of the throttling section. Alternatively, as... Figure 8A As shown, the throttling section 22 is a Venturi tube section, and the minimum inner diameter D1 of the throttling section is the inner diameter at the straight section of the upper throat of the Venturi tube section. Further, in other embodiments, such as... Figure 8B As shown, the downstream of the main body section 21 of the inlet pipe can also be provided with multiple Venturi tube sections. In this structure, the inner diameter of the straight throat of the Venturi tube section closest to the assembly hole 104 of the inlet pipe is used as the minimum inner diameter D1 of the throttling section.

[0119] In this embodiment, the end of the throttling section 22 is welded to the inlet pipe assembly hole 104, and the end of the throttling section 22 is the outlet end of the inlet pipe 2. However, the present invention does not limit this in any way. In other embodiments, the inlet pipe may also be integrally formed with the body at the location of the inlet pipe assembly hole.

[0120] In this embodiment, the mixing chamber 106 is a concave cavity formed by stretching the area on the partition 3 opposite to the liquid inlet pipe assembly hole 104 towards the side where the liquid outlet end 102 of the main body is located. Simultaneously, the partition 3 divides the accommodating cavity 103 within the main body into a first cavity 1031 and a second cavity 1032, where the high-speed refrigerant output from the partition hole 31 expands and mixes again in the second cavity 1032. Further, as... Figure 9 As shown, the vertical distance H2 between the outer bottom wall of the mixing chamber 106 and the inner bottom wall of the liquid outlet end of the main body satisfies: 0 ≤ H2 ≤ 3 mm. Based on the determined depth H1 of the mixing chamber 106 and the inner diameter of the main body, the setting of the vertical distance H2 achieves the control of the internal volume of the second cavity 1032, effectively avoiding the re-occurrence of gas-liquid separation of the dispersed refrigerant due to the second cavity 1032 being too large, thereby improving the uniformity of the distribution. Preferably, H2 is set to equal 0, that is, the outer bottom wall of the mixing chamber 106 abuts against the inner bottom wall of the liquid outlet end of the main body. However, the present invention does not impose any limitation on this. In other embodiments, the outer bottom wall of the mixing chamber 106 does not abut against the inner bottom wall of the liquid outlet end of the main body, but in this case, the vertical distance H2 is also less than or equal to 3 mm.

[0121] In this embodiment, as shown in FIG1, the main body 1 includes a cylinder 11, an end cap 12, and a liner 13. The cylinder 11 is integrally formed and has a single-end open shape, with multiple diversion holes formed at the bottom of the cylinder 11. The liner 13 is welded to the inner bottom surface of the cylinder 11, forming the liquid outlet end wall of the main body 1 together with the cylinder 11. The inner surface of the liner 13 serves as the inner bottom wall of the liquid outlet end of the main body, and the vertical distance H2 is the distance from the outer bottom wall of the cavity 106 to the inner surface of the liner 13. Multiple liner holes coaxial with the multiple diversion holes are formed on the liner 13, and the liner holes and the corresponding diversion holes together form the branch pipe hole 105. However, the present invention does not limit this in any way.

[0122] Although this embodiment uses the partition plate and the inner bottom wall of the liquid outlet of the main body to form a second cavity as an example, the present invention does not limit this. In other embodiments, the second cavity may not be required in the main body, and multiple branch pipes may extend into the main body and connect directly to the corresponding partition plate holes, that is, the refrigerant is directly distributed from the partition plate holes to the multiple branch pipes. Furthermore, this embodiment does not limit the method of forming the mixing chamber.

[0123] Example 2

[0124] This embodiment is basically the same as Embodiment 1 and its variations, except that the mixing cavity 106 is formed in a different way and the structure of the body 1 is also different.

[0125] Specifically, such as Figure 10 As shown, in this embodiment, the body 1' includes an end cap 12' and a liner 13'. One end of the end cap 12' has a liquid inlet pipe assembly hole ( Figure 10 The inlet pipe 2 is installed inside the inlet pipe assembly hole (therefore it is not marked). The liner 13' is embedded in the other end of the end cap 12', and multiple branch pipe holes 105' are formed on the liner 13'. The mixing chamber 106' is a recessed cavity formed on the liner 13'. The dispersed refrigerant after being mixed by the reflection of the mixing chamber 106' is directly distributed into the multiple branch pipe holes 105'. Although Figure 2 and Figure 10 The number of liner plates is always one. However, the present invention does not limit this. In other embodiments, multiple liner plates may be provided, which are stacked and welded to the inner wall of the cylinder, the outer wall of the cylinder, or embedded in the open end of the end cap.

[0126] Example 3

[0127] This embodiment is basically the same as Embodiment 1 and its variations, except that the structures of the liquid inlet assembly hole 104 and the liquid inlet pipe 4 are different.

[0128] like Figure 11As shown, in this embodiment, the inlet pipe assembly hole 104 is a through hole for a straight section without welding, and the throttling section 22 is a Venturi tube section with the minimum inner diameter D1 of the throttling section being the inner diameter at the straight section of the upper throat of the Venturi tube section. A gradually expanding section 23 is formed on the Venturi tube section, located downstream of the minimum inner diameter D1 of the throttling section, with an arc-shaped generatrix on its outer wall. The gradually expanding section 23 extends into the accommodating cavity through the inlet pipe assembly hole 104, and its outer wall is welded to the inner wall of the inlet end 101 of the main body in a surface-contact manner. Definition: The extension line of the inner wall generatrix of the straight section at the upper throat of the Venturi tube section intersects the virtual extension surface of the inner wall of the inlet end 101 at the inlet pipe assembly hole 104 at position K. The cross-section of the inlet pipe 2 passing through position K is the outlet end face of the inlet pipe 2. Similarly, in this embodiment, the vertical distance H0 between the end face of the liquid inlet pipe and the end face of the opening of the mixing chamber 106 satisfies 3.5mm≤H0≤11.5mm.

[0129] However, the present invention does not limit the specific structure of the inlet pipe assembly hole. In other embodiments, when the inlet pipe assembly hole has a welded assembly straight section, the inlet pipe may also include a straight section (such as the throat of a Venturi tube section) connected to the welded assembly straight section and a gradually expanding section located downstream of the minimum inner diameter of the throttling section with an arc-shaped generatrix on its outer wall. Alternatively, when the throttling section is a throttling orifice section with a substantially constant inner diameter and smaller than the outer diameter of the main body section, a gradually expanding section with an arc-shaped generatrix on its outer wall is provided downstream of the throttling orifice section so that the outer wall of the gradually expanding section can be welded to the inner wall of the inlet end of the main body in a surface contact manner.

[0130] In this embodiment, the number of branch pipes 4 is ten. However, the present invention does not impose any limitation on this.

[0131] Example 4

[0132] This embodiment is basically the same as Embodiment 1 and its variations, except that the structure of branch pipe 4 is different.

[0133] In this embodiment, as Figure 12 As shown, each branch pipe 4 includes a first pipe section 41 and a second pipe section 42 located downstream of the first pipe section 41, with its inner diameter decreasing relative to the downstream end of the first pipe section 41. The difference Δd between the inner diameter d11 at the downstream end of the first pipe section 41 and the inner diameter d12 at the downstream end of the second pipe section 42 is 0.1mm≤Δd≤3.5mm. A reflective section 411 is formed on the first pipe section 41, which bends and extends to one side relative to the center line of the branch pipe hole 105. The axis at the upstream end of the reflective section 411 and the axis at the downstream end of the reflective section intersect to form an angle θ, with 90°≤θ≤175°. Based on the reflective section 411, the axis of the second pipe section 42 intersects the center line of the branch pipe hole 105.

[0134] The larger inner diameter of the first pipe section 41 increases the refrigerant flow rate into each branch pipe 4 and reduces the refrigerant distribution resistance, while the relatively smaller inner diameter of the second pipe section 42 increases the refrigerant flow velocity. Furthermore, the inner diameter difference Δd precisely controls the rate of refrigerant depressurization and acceleration caused by the second pipe section 42, preventing excessive refrigerant pressure loss due to excessive acceleration. The reflector section 411 ensures that the second pipe section 42 is no longer coaxial with the centerline of the branch pipe hole. When a pressure wave exists downstream and oscillates upstream, the reflector section 411 reflects and absorbs part of the pressure wave and changes the propagation direction of the remaining pressure wave, causing it to attenuate rapidly. This effectively prevents the downstream pressure wave from affecting the upstream branch pipe hole 105, further improving distribution performance.

[0135] In this embodiment, the branch pipe 4 further includes a third pipe section 43 welded to the first pipe section 41, and the second pipe section 42 is located on the third pipe section 43. Specifically, as shown... Figure 12 As shown, the third pipe section 43 has a sleeve connection portion 431 that is sleeved over the end of the first pipe section 41 and has a flared structure. Downstream of the sleeve connection portion 431, a second pipe section 42 with a tapered inner diameter is formed. However, the present invention does not limit this in any way. In other embodiments, the end of the first pipe section 41 may also be provided with a sleeve connection portion 431 welded to the third pipe section 43. In this case, the second pipe section is still a tapered inner diameter structure located downstream of the sleeve connection portion 431, such as... Figure 13A As shown.

[0136] Figure 13B and Figure 13C The diagram shown is a structural schematic of a branch pipe provided in another embodiment of this utility model. Figure 13B In the middle, the first pipe section 41, the second pipe section 42, and the third pipe section 43 are formed as a single unit; Figure 13C In this design, the first pipe segment 41 and the second pipe segment 42 are integrally formed, while the third pipe segment 43 is welded to the end of the second pipe segment 42. In other embodiments, the second pipe segment can also be formed directly at the joint between the first and second pipe segments; furthermore, multiple second pipe segments can be provided at the downstream end of the first pipe segment. The third pipe segment can be configured as a straight pipe or a bend depending on the actual application.

[0137] Example 5

[0138] This embodiment is basically the same as Embodiment 1 and its variations, except that branch pipe 4 has a different structure.

[0139] like Figure 14As shown, in this embodiment, each branch pipe 4 includes a connecting straight section 41', a jet section 42' located downstream of the connecting straight section 41' with an inner diameter smaller than the inner diameter d11' of the connecting straight section body 411', and a branch section 43' located downstream of the jet section 42'. The branch section 43' has a constant diameter section 431' with an inner diameter substantially close to the inner diameter d11' of the connecting straight section body. The difference Δd' between the inner diameter d11' of the connecting straight section body and the minimum inner diameter d12' at the jet section 42' is: 0.1mm ≤ Δd' ≤ 3.5mm. The connecting straight section body 411' refers to the pipe section on the connecting straight section 41' where both the inner and outer diameters remain substantially unchanged.

[0140] The refrigerant flow distributor provided in this example increases the refrigerant flow velocity within each branch pipe 4 by adding a jet section 42' to each branch pipe 4 to meet the performance requirements of the downstream heat exchanger components. Furthermore, the inner diameter difference Δd' precisely controls the degree of pressure reduction and acceleration of the refrigerant by the jet section 42', avoiding excessive refrigerant pressure loss due to excessive acceleration. Further, the inner diameter d13' of the equal-diameter section 431' on the branch section 43' is approximately close to the inner diameter d11' of the connecting straight section body 411' to reduce the flow resistance of the refrigerant within the branch pipe 4, achieving balanced control of refrigerant flow velocity and pressure loss.

[0141] In this embodiment, the jet section 42' is integrally formed on the upstream constriction section of the branch section 43', and the two are integrally formed and welded together to the connecting straight section 41'. However, this utility model does not limit this in any way. In other embodiments, such as Figure 15A As shown, the jet section 42' can also be integrally formed downstream of the connecting straight section 41', and then welded to the branch section 43'. Or, as... Figure 15B As shown, the straight section 41', the jet section 42', and the branch section 43' are integrally formed.

[0142] Alternatively, the jet section 42' may be a jet orifice plate placed at the junction of the straight section 41' and the branch section 43'. Specifically, as shown... Figure 15C As shown, the upstream end of branch segment 43' is flared and connected to the downstream end of connecting straight segment 41', with jet section 42' placed at the flared end of branch segment 43'. Similarly, when the flared end of connecting straight segment 41' is connected to the upstream end of branch segment 43', jet section 42' can also be placed at the flared end of connecting straight segment 41', as shown. Figure 15D As shown.

[0143] Alternatively, the jet section 42' can be configured as a straight pipe segment, with its two ends respectively connected to the connecting straight section 41' and the branch section 43', such as... Figure 15E As shown. In Figure 14 , Figure 15A as well as Figure 15EIn the middle, the jet section 42' has a maintaining length L42 with a substantially constant inner diameter. However, this utility model does not impose any limitations on this.

[0144] This utility model does not limit the formation method of the jet section, and other jet section structures that can achieve a reduction in the inner diameter of the relatively connected straight section body are all within the protection scope of this utility model.

[0145] In summary, the refrigeration distributor provided by this invention achieves refrigerant pressure reduction and speed increase by setting a throttling section with a smaller inner diameter than the main body section on the inlet pipe. By controlling the inner diameter ratio of the two sections, sufficient mass flow rate at the main body inlet is ensured, enhancing the collision effect of the refrigerant in the mixing chamber and improving the mixing uniformity. Furthermore, controlling the minimum inner diameter of the throttling section effectively reduces pressure loss and noise generated by throttling. Simultaneously, setting the ratio of the inner cavity cross-sectional area S2 at the mixing chamber opening to the flow cross-sectional area S1 at the minimum inner diameter of the throttling section precisely controls the volume of the mixing chamber. This ensures that while providing expansion space and distribution channels for the refrigerant, the mixing chamber effectively avoids problems such as excessive refrigerant expansion and sudden drops in refrigerant velocity caused by an excessively large mixing chamber. It ensures that the input refrigerant has sufficient kinetic energy to be incident on the bottom wall of the mixing chamber to participate in reflective mixing, promoting the full atomization of the two-phase refrigerant into a dispersed state. It also ensures that the refrigerant maintains a dispersed flow during distribution, thereby improving the uniformity of the distribution.

[0146] Furthermore, based on the refrigerant expansion model established by the fitting, after determining the outer diameter D0 of the main section and the minimum cross-sectional area S1 of the throttling section, the inner cavity cross-sectional area S2 at the opening of the mixing cavity can be calculated to clarify the parameter selection of the mixing cavity before testing, providing guidance for the design before development, and thus greatly improving the development speed of the splitter product, especially the development of series products.

[0147] Although the present invention has been disclosed above by way of preferred embodiments, it is not intended to limit the present invention. Any person skilled in the art may make some modifications and refinements without departing from the spirit and scope of the present invention. Therefore, the scope of protection of the present invention shall be determined by the scope of protection claimed in the claims.

[0148] In the following appendix: the units for the outer diameter D0 of the main body of the inlet pipe, the minimum inner diameter D1 of the throttling section, and the inner diameter D2 at the opening of the mixing chamber are all mm; the units for the standard deviation of refrigerant mass flow rate STD, STD0, STD test, STD0 test, and STD01 test at the outlet of each branch pipe of the optimized sample, control sample, optimized physical sample, control physical sample, and physical sample of the analytical separator are all g / s; the cross-sectional area ratio S2 / S1 is dimensionless.

[0149] Appendix 1

[0150]

[0151]

[0152]

[0153] Appendix 2

[0154]

[0155] Appendix 3

[0156]

[0157] Appendix 4

[0158]

Claims

1. A refrigerant distributor, characterized in that, include: The body includes an inlet end, an outlet end, and a accommodating cavity that connects the inlet end and the outlet end respectively. An inlet pipe assembly hole is formed on the inlet end of the body, and multiple branch pipe holes for liquid distribution are formed on the outlet end. A mixing cavity is formed in the accommodating cavity of the body, with its opening facing the inlet pipe assembly hole and extending recessed towards the outlet end of the body. When projected along the axial direction of the body, the mixing cavity is located on the inner periphery of the multiple branch pipe holes. The inlet pipe has its end connected to the inlet pipe assembly hole and does not extend into the mixing chamber. The inlet pipe includes a main body section with a basically unchanged inner diameter and wall thickness, and a throttling section located downstream of the main body section with an inner diameter smaller than the outer diameter of the main body section. Among them, the minimum inner diameter D1 of the throttling section and the outer diameter D0 of the main body section satisfy: 0.45D0≤D1≤0.95D0, and the cross-sectional area S2 of the inner cavity at the opening of the mixing chamber is greater than the flow cross-sectional area S1 at the minimum inner diameter of the throttling section, and S2 / S1 satisfies the following refrigerant expansion model: (S2 / S1) min =-1.26×ln(D0)+4.572; (S2 / S1) max =-1.75×ln(D0)+9.757; Wherein, (S2 / S1) min The minimum value of S2 / S1 is (S2 / S1). max The maximum value of S2 / S1; The depth H1 of the mixing chamber satisfies: 2mm≤H1≤22mm. The depth H1 of the mixing chamber refers to the vertical distance from the center of the bottom wall of the mixing chamber to the end face of the opening of the mixing chamber.

2. The refrigerant distributor according to claim 1, characterized in that, The vertical distance H0 from the outlet end face of the inlet pipe to the opening end face of the mixing chamber satisfies 3.5mm≤H0≤11.5mm.

3. The refrigerant distributor according to claim 1, characterized in that, The inlet pipe has a gradually expanding section located downstream of the minimum inner diameter of the throttling section and with an arc-shaped generatrix on its outer wall. The outer wall of the gradually expanding section is welded to the inner wall of the inlet end of the main body.

4. The refrigerant distributor according to claim 1, characterized in that, The throttling section is a throttling orifice section with an inner diameter that remains basically unchanged and is smaller than the outer diameter of the main body section; Alternatively, the refrigerant distributor may also include a throttling orifice plate disposed within the main body section. The throttling orifice plate has a throttling orifice with an inner diameter smaller than the outer diameter of the main body section. A throttling section is formed at the location of the throttling orifice. The throttling orifice is a straight orifice with a basically constant inner diameter or an inner diameter that gradually decreases along the flow direction of the refrigerant in the liquid inlet pipe. Alternatively, the throttling section is a Venturi tube section, and the minimum inner diameter D1 of the throttling section is the inner diameter of the straight section at the throat of the Venturi tube section.

5. The refrigerant distributor according to claim 1, characterized in that, The refrigerant distributor also includes multiple branch pipes welded to multiple branch pipe holes. Each branch pipe includes a first pipe section and a second pipe section located downstream of the first pipe section with a smaller inner diameter relative to the downstream end of the first pipe section. The difference Δd between the inner diameter d11 at the downstream end of the first pipe section and the inner diameter d12 at the downstream end of the second pipe section is 0.1mm≤Δd≤3.5mm. A reflective section is formed on the first pipe section, which bends and extends to one side relative to the center line of the branch pipe hole. The axis at the upstream end of the reflective section and the axis at the downstream end of the reflective section intersect to form an angle θ with 90°≤θ≤175°. Based on the reflective section, the axis of the second pipe section intersects the center line of the branch pipe hole.

6. The refrigerant distributor according to claim 1, characterized in that, The refrigerant distributor also includes multiple branch pipes welded to multiple branch pipe holes. Each branch pipe includes a connecting straight section, a jet section located downstream of the connecting straight section with an inner diameter smaller than that of the connecting straight section, and a branch section located downstream of the jet section. The branch section has an equal diameter section with an inner diameter that is basically close to that of the connecting straight section. The difference Δd' between the inner diameter d11' of the connecting straight section and the minimum inner diameter d12' at the jet section is: 0.1mm≤Δd'≤3.5mm.

7. The refrigerant distributor according to claim 1, characterized in that, The refrigerant distributor also includes a partition plate disposed in the accommodating cavity of the main body. The area on the partition plate opposite to the liquid inlet pipe assembly hole is recessed towards the liquid outlet end of the main body to form a mixing cavity. Multiple partition plate holes are formed on the plane of the partition plate on the outer periphery of the mixing cavity in a ring distribution around the center line of the main body. The multiple partition plate holes distribute the refrigerant mixed in the mixing cavity to the liquid outlet end of the main body.

8. The refrigerant distributor according to claim 7, characterized in that, The partition divides the accommodating cavity into a first cavity near the liquid inlet end of the main body and a second cavity near the liquid outlet end of the main body. The second cavity is an annular cavity surrounding the mixing cavity, and the vertical distance H2 between the outer bottom wall of the mixing cavity and the inner bottom wall of the liquid outlet end of the main body satisfies: 0≤H2≤3mm.

9. The refrigerant distributor according to claim 7, characterized in that, The mixing chamber is a constant-diameter chamber with a basically uniform inner diameter; or, the mixing chamber is any one of a conical chamber, a frustum-cone chamber, or a hemispherical chamber with an inner diameter that gradually decreases along the refrigerant inflow direction; or, the mixing chamber is a combination of multiple constant-diameter chambers, conical chambers, frustum-cone chambers, or hemispherical chambers.

10. A heat exchanger assembly, characterized in that, Includes the refrigerant distributor according to any one of claims 1 to 9.

11. An air conditioner, characterized in that, Includes the heat exchanger assembly as described in claim 10.

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