Cooling shunt, heat exchanger assembly, and air conditioner

Abstract

To provide a flow divider for cooling capable of improving flow dividing uniformity.SOLUTION: And a liquid feeding pipe including a main body portion 21 having an inside diameter and a wall thickness that are substantially unchanged, and a narrowed portion 22 located downstream of the main body portion 21 and having an inside diameter smaller than an outside diameter D0 of the main body portion 21. The minimum inside diameter D1 of the throttling portion 22 and the outside diameter D0 of the main body portion 21 satisfy 0. 45D0 ≤ D1 ≤ 0. 95D0, the lumen cross-sectional area S2 at the opening of the mixing chamber 106 is greater than the flow cross-sectional area S1 at the minimum inside diameter D1 of the throttling portion 22, and S2 / satisfy ( / ) min = - 1.26 * ln () + 4.572, ( / ) max = - 1.75 * ln () + 9.757. S1 S1 D0 S2 S2 D0 S1.SELECTED DRAWING: Figure 2

Description

[Technical Field]

[0001] The present invention relates to the technical field of refrigerant distribution, and more particularly to cooling flow dividers, heat exchanger assemblies and air conditioners. [Background technology]

[0002] In a compression-type cooling system, the evaporator is the heat exchange component that transfers heat. Its multi-branch structure can improve its heat exchange efficiency, and the performance of the evaporator directly affects the overall efficiency of the cooling system. After the liquid refrigerant transforms from single-phase liquid to two-phase gas / liquid through the throttle element, whether the two-phase gas / liquid refrigerant can be evenly distributed to each branch of the evaporator has a significant impact on the evaporator's heat exchange performance. Experiments have shown that uneven refrigerant distribution can reduce heat exchange capacity by as much as 25%. To improve the evaporator's heat exchange efficiency, a flow divider must be installed before the evaporator to distribute the refrigerant as evenly as possible to each branch of the evaporator. However, due to limitations in the flow divider structure and its installation conditions, most conventional flow dividers have poor flow distribution uniformity, which is significantly affected by the installation angle. Only Venturi-type flow dividers with smooth linearity are superior. However, smooth linearity is not only difficult to achieve, but also difficult to control the consistency after processing.

[0003] To this end, we propose a reflective flow divider that uniformly mixes gas and liquid phases by reflecting the refrigerant through a reflective chamber, such as the air conditioning flow divider structure proposed by the inventors in Chinese Patent No. CN216204506U and the air conditioner distributor proposed in Chinese Patent No. CN111919067A. The reflective mixing of the refrigerant within the reflective chamber minimizes the impact of the installation angle on the flow divider, allowing it to better adapt to different installation angle conditions. While the performance of the reflective flow divider is superior to that of conventional flow dividers in terms of installation angle, the vertical flow distribution uniformity of the reflective flow divider in Chinese Patent No. CN216204506U is somewhat lower than that of conventional Venturi distributors, requiring further research to improve refrigerant distribution performance. In Chinese Patent No. CN111919067A, the guide member is inserted into the guide tank through the guide hole, forming a gap S in the guide space. The refrigerant reflected and switched in the switching space flows into the guide space with a smaller cross-sectional area and is mixed. In this configuration, the guide member inserted into the guide vessel reduces the internal volume of the guide vessel, resulting in insufficient mixing space for the refrigerant after reflection. At this time, the pressure inside the guide vessel is high, and the refrigerant is distributed to each branch port through the gap S and distribution space without being mixed, resulting in the gas-liquid two-phase separation and serious impact on the uniformity of the divided flow. Furthermore, the guide space formed by the gap S creates excessive refrigerant pressure loss, which affects compressor efficiency, and excessive pressure loss also leads to noise problems. Summary of the Invention [Problem to be solved by the invention]

[0004] SUMMARY OF THE INVENTION The present invention provides a cooling flow divider, a heat exchanger assembly, and an air conditioner to overcome at least one of the shortcomings of the prior art. [Means for solving the problem]

[0005] In order to achieve the above object, the present invention provides a refrigerant flow divider, a main body including a liquid supply end, a liquid drain end, and a storage chamber that connects the liquid supply end and the liquid drain end, wherein the liquid supply end has a liquid supply pipe mounting hole, and the liquid drain end has a plurality of branch pipe holes for dividing the liquid, and within the storage chamber is formed a mixing chamber whose opening faces the liquid supply pipe mounting hole and extends recessed toward the liquid drain end of the main body, and when projected along the axial direction of the main body, the mixing chamber includes: a main body located on the inner periphery of the plurality of branch pipe holes; a liquid supply pipe whose end is connected to the liquid supply pipe mounting hole and does not extend into the mixing chamber, the liquid supply pipe including a main body portion whose inner diameter and wall thickness do not change substantially; and a throttle portion located downstream of the main body portion and whose inner diameter is smaller than the outer diameter of the main body; The minimum inner diameter D1 of the throttle portion and the outer diameter D0 of the main body portion satisfy 0.45D0≦D1≦0.95D0, the lumen cross-sectional area S2 at the opening of the mixing chamber is larger than the flow cross-sectional area S1 at the minimum inner diameter of the throttle portion, 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, (S2 / S1) min is the minimum value of S2 / S1, and (S2 / S1) max provides a refrigerant flow divider with the maximum value of S2 / S1.

[0006] In one embodiment of the present invention, the depth H1 of the mixing chamber satisfies 2 mm≦H1≦22 mm, and the depth H1 of the mixing chamber is the vertical distance from the center of the inner bottom wall of the mixing chamber to the end face of the opening of the mixing chamber.

[0007] In one embodiment of the present invention, the vertical distance H0 from the end face of the discharge end of the liquid supply tube to the end face of the opening of the mixing chamber satisfies 3.5 mm≦H0≦11.5 mm.

[0008] In one embodiment of the present invention, the liquid supply pipe is formed with a divergent section located downstream of the minimum inside diameter of the constricted section and having an outer wall generatrix that forms an arc-shaped curve, and the outer wall of the divergent section is tightly welded to the inner wall of the liquid supply end of the body.

[0009] In one embodiment of the present invention, the restrictor portion is a restrictor hole portion whose inner diameter is substantially constant and smaller than the outer diameter of the main body portion, Alternatively, the refrigerant flow divider further includes a throttling hole plate provided in the main body, the throttling hole plate having an inner diameter smaller than the outer diameter of the main body, the throttling hole having a throttling portion formed therein, the throttling hole being a straight hole whose inner diameter does not change substantially, or a collecting hole whose inner diameter gradually becomes smaller along the flow direction of the refrigerant in the liquid supply pipe, Alternatively, the throttle portion is a Venturi tube portion, and the minimum inner diameter D1 of the throttle portion is the inner diameter of a straight throat portion of the Venturi tube portion.

[0010] In one embodiment of the present invention, the refrigerant flow divider further includes a plurality of branch pipes welded to the plurality of branch pipes, respectively. Each branch pipe includes a first pipe section and a second pipe section located downstream of the first pipe section, the second pipe section having an inner diameter smaller than that of the downstream end of the first pipe section. The difference Δd between the inner diameter d11 of the downstream end of the first pipe section and the inner diameter d12 of the downstream end of the second pipe section is in the range of 0.1 mm≦Δd≦3.5 mm. The first pipe section has a reflector section bent and extending to one side of the center line of the branch pipe section, and the axis of the upstream end of the reflector section intersects with the axis of the downstream end of the reflector section at an angle θ which is in the range of 90°≦θ≦175°. The reflector section allows the axis of the second pipe section to intersect with the center line of the branch pipe section.

[0011] In one embodiment of the present invention, the refrigerant flow divider further includes a plurality of branch pipes welded to the plurality of branch pipe holes, each branch pipe including a straight connection portion, a jet portion located downstream of the straight connection portion and having an inner diameter smaller than that of the straight connection portion body, and a branch portion located downstream of the jet portion, wherein the branch portion has an isodiameter portion whose inner diameter is approximately the same as that of the straight connection portion body, and the difference Δd′ between the inner diameter d11′ of the straight connection portion body and the minimum inner diameter d12′ of the jet portion satisfies 0.1 mm≦Δd′≦3.5 mm.

[0012] In one embodiment of the present invention, the refrigerant distributor further includes a partition plate provided in the main body storage chamber, and the area of ​​the partition plate facing the liquid supply pipe mounting hole is recessed toward the liquid discharge end of the main body to form a mixing chamber, and a plurality of partition plate holes are formed on the partition plate plane on the outer periphery of the mixing chamber, distributed in a ring shape around the center line of the main body, and the plurality of partition plate holes distribute the refrigerant mixed in the mixing chamber to the liquid discharge end of the main body.

[0013] In one embodiment of the present invention, the partition plate divides the storage chamber into a first chamber near the main body supply end and a second chamber near the main body discharge end, the second chamber being an annular chamber surrounding the mixing chamber, and the vertical distance H2 between the outer bottom wall of the mixing chamber and the inner bottom wall of the main body discharge end satisfies 0≦H2≦3 mm.

[0014] In one embodiment of the present invention, the mixing chamber is a constant diameter chamber with approximately the same lumen diameter, or is any one of a tapered chamber, a truncated conical chamber, or a hemispherical chamber whose inner diameter gradually decreases along the direction of refrigerant inflow, or is a combination of two or more of a constant diameter chamber, a tapered chamber, a truncated conical chamber, or a hemispherical chamber.

[0015] According to another aspect, the present invention further provides a heat exchanger assembly including the refrigerant flow divider described above.

[0016] According to another aspect, the present invention further provides an air conditioner including the heat exchanger assembly described above. [Effects of the Invention]

[0017] As described above, the cooling flow diverter of the present invention achieves refrigerant pressure drop acceleration by providing a throttling section in the supply pipe, whose inner diameter is smaller than that of the main body section. Controlling the inner diameter ratio between the two sections ensures sufficient mass flow at the main body inlet, enhancing the refrigerant collision effect within the mixing chamber and improving refrigerant mixing uniformity. Controlling the minimum inner diameter of the throttling section effectively reduces pressure loss and throttling-related noise within the flow diverter. Accurately controlling the internal volume of the mixing chamber by setting the ratio of the lumen cross-sectional area S2 at the opening of the mixing chamber to the flow cross-sectional area S1 at the minimum inner diameter of the throttling section allows the mixing chamber to provide expansion space and distribution channels for the refrigerant, effectively avoiding problems such as excessive refrigerant expansion and a sudden drop in refrigerant flow rate due to an oversized mixing chamber. The input refrigerant has sufficient kinetic energy to impinge on the inner bottom wall of the mixing chamber and undergo reflective mixing, promoting fully dispersed atomization of the two-phase refrigerant and maintaining a dispersed flow during distribution, further improving flow distribution uniformity. Furthermore, based on the fitting and established refrigerant expansion model, the outer diameter D0 and cross-sectional area S1 of the main body are determined, and then the lumen cross-sectional area S2 at the opening of the mixing chamber is calculated. This clarifies the parameter selection of the untested mixing chamber, provides guidance for pre-development design, and further significantly improves the development speed of flow divider products, especially the development of series products.

[0018] In order to make the above and other objects, features and advantages of the present invention more clearly understandable, the following particularly preferred embodiments will be described in detail in conjunction with the drawings. [Brief explanation of the drawings]

[0019] [Figure 1A] FIG. 1 is a structural schematic diagram of a conventional brass flow divider with a venturi structure. [Figure 1B] 1 is a schematic diagram of the structure of a conventional pressure drop type flow divider. [Figure 2] 1 is a structural schematic diagram of a cooling flow shunt according to a first embodiment of the present invention; [Figure 3]FIG. 3 is a structural schematic diagram in which the liquid supply pipe and branch pipe in FIG. 2 have been removed. [Figure 4] FIG. 1 is a diagram illustrating the test principle of the optimized real sample performance test device. [Figure 4A] FIG. 2 is a diagram showing the change in enthalpy value when the refrigerant flows through the electronic expansion valve. [Figure 5] FIG. 3 is a structural schematic diagram of the partition plate in FIG. 2. [Figure 6A] FIG. 10 is a structural schematic diagram of a partition plate according to another embodiment of the present invention. [Figure 6B] FIG. 10 is a structural schematic diagram of a partition plate according to another embodiment of the present invention. [Figure 6C] FIG. 10 is a structural schematic diagram of a partition plate according to another embodiment of the present invention. [Figure 7] FIG. 4 is a structural schematic diagram of a cooling flow shunt according to another embodiment of the present invention. [Figure 8A] FIG. 10 is a structural schematic diagram of a liquid supply pipe according to another embodiment of the present invention. [Figure 8B] FIG. 10 is a structural schematic diagram of a liquid supply pipe according to another embodiment of the present invention. [Figure 9] FIG. 4 is a structural schematic diagram of a cooling flow shunt according to another embodiment of the present invention. [Figure 10] FIG. 4 is a structural schematic diagram of a cooling flow shunt according to a second embodiment of the present invention. [Figure 11] FIG. 10 is a structural schematic diagram of a cooling flow shunt according to a third embodiment of the present invention. [Figure 12] FIG. 10 is a structural schematic diagram of a branch pipe according to a fourth embodiment of the present invention. [Figure 13A] FIG. 10 is a structural schematic diagram of a branch pipe according to another embodiment of the present invention. [Figure 13B] FIG. 10 is a structural schematic diagram of a branch pipe according to another embodiment of the present invention. [Figure 13C] FIG. 10 is a structural schematic diagram of a branch pipe according to another embodiment of the present invention. [Figure 14] 5 is a structural schematic diagram of a refrigerant uniform distributor according to a fifth embodiment of the present invention. [Figure 15A] 4 is a structural schematic diagram of a branch pipe in a refrigerant uniform distributor according to another embodiment of the present invention; FIG. [Figure 15B] 4 is a structural schematic diagram of a branch pipe in a refrigerant uniform distributor according to another embodiment of the present invention; FIG. [Figure 15C] 4 is a structural schematic diagram of a branch pipe in a refrigerant uniform distributor according to another embodiment of the present invention; FIG. [Figure 15D] 4 is a structural schematic diagram of a branch pipe in a refrigerant uniform distributor according to another embodiment of the present invention; FIG. [Figure 15E] 4 is a structural schematic diagram of a branch pipe in a refrigerant uniform distributor according to another embodiment of the present invention; FIG. DETAILED DESCRIPTION OF THE INVENTION

[0020] Example 1 In a conventional reflective flow divider, high-pressure, high-velocity two-phase refrigerant reflects and collides inside the reflective chamber, promoting the mixing of the gas and liquid phases. Due to the reflection of the reflective chamber, this type of flow divider is less affected by gravity, so it has a good flow dividing effect even when not installed vertically. However, when installed vertically, the performance of this flow divider is difficult to reach the flow dividing standard of a conventional brass Venturi flow divider.

[0021] In view of this, as shown in Figure 2, this embodiment provides an air conditioning flow diverter including a main body 1, a feed pipe 2, and a plurality of branch pipes 4. The main body 1 includes a feed end 101, a drain end 102, and a storage chamber 103 connecting the feed end 101 and the drain end 102, respectively. The feed end 101 of the main body is formed with a feed pipe mounting hole 104, and the drain end 102 is formed with a plurality of branch pipe holes 105 for dividing the flow and connecting to the plurality of branch pipes 4, respectively. A mixing chamber 106 is formed within the storage chamber 103 of the main body, with its opening facing the feed pipe mounting hole 104 and extending recessed toward the main body drain end 102. When projected along the axial direction of the main body 1, the mixing chamber 106 is located on the inner periphery of the plurality of branch pipe holes 105. An end of the liquid supply pipe 2 is connected to the liquid supply pipe mounting hole 104 and does not extend into the mixing chamber 106, and the liquid supply pipe 2 includes a main body portion 21 and a throttling portion 22 located downstream of the main body portion 21 and having an inner diameter smaller than that of the main body portion 21. The minimum inner diameter D1 of the throttling portion 22 and the outer diameter D0 of the main body portion 21 satisfy 0.45D0≦D1≦0.95D0, the lumen cross-sectional area S2 at the opening of the mixing chamber 106 is larger than the flow cross-sectional area S1 at the minimum inner diameter of the throttling portion, 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, (S2 / S1) min is the minimum value of S2 / S1, and (S2 / S1) max is the maximum value of S2 / S1.

[0022] In this embodiment, the refrigerant flow distributor further includes a partition plate 3 disposed within the main body storage chamber 103. The partition plate 3 divides the storage chamber 103 into a first chamber 1031 near the main body supply end and a second chamber 1032 near the main body discharge end. The second chamber 1032 is an annular chamber surrounding the mixing chamber 106. The area of ​​the partition plate 3 facing the supply pipe mounting hole 104 is recessed toward the main body discharge end to form the mixing chamber 106. The partition plate plane around the outer periphery of the mixing chamber 106 is formed with a plurality of partition plate holes 31 distributed annularly around the centerline of the main body (i.e., when projected along the axial direction of the main body 1, the plurality of partition plate holes 31 are located on the outer periphery of the mixing chamber 106). The plurality of partition plate holes 31 distribute the refrigerant mixed in the mixing chamber 106 to the second chamber 1032 and then to the plurality of branch pipe holes 105 at the main body discharge end 102 via the second chamber 1032.

[0023] The high compatibility of reflective flow diverters with respect to installation angle allows them to be applied in cooling systems with limited installation space, but the drawback of poor flow diversion performance in a vertical installation significantly limits the range of applications of such flow diverters. Optimization of the performance of reflective flow diverters mainly focuses on the reflection angle and the insertion depth of the branch pipes that divert the flow, but these optimizations have very limited effect on improving the uniformity of the flow. To solve the problem of poor flow diversion performance when reflective flow diverters are installed vertically, the inventors conducted extensive analytical tests.

[0024] A commercially available brass venturi-structured flow diverter (hereinafter referred to as the Venturi flow diverter), which has excellent flow uniformity when installed vertically, is used as a control sample (the structure of which is shown in Figure 1A). A CFD (computational fluid dynamics) simulation analysis is performed on the air conditioning flow diverter structure proposed by the inventors in China CN216204506U (hereinafter referred to as the flow diverter sample to be analyzed) and another commercially available common pressure drop type flow diverter (the reference sample, the structure of which is shown in Figure 1B).

[0025] Sample structure parameters: The outer diameter of the main supply pipe of the control sample, reference sample, and the flow divider sample to be analyzed is 16 mm, the number of branch pipe holes on the discharge side is 11, the minimum inner diameter of the constriction part of the flow divider sample to be analyzed is 6 mm, the diameter of the opening of the mixing chamber is 18.8 mm, and the depth H1 of the mixing chamber is 12 mm.

[0026] The software used for the CFD (Computational Fluid Dynamics) simulation analysis is ANSYS software, and the condition settings are shown below.

[0027] Turbulence model: Realizable k-ε model, Network parameters: Tetrahedral unstructured mesh is used. Operating conditions: the working substance is R410A refrigerant, the wall boundary is adiabatic boundary, the inlet dryness is 0.2, and the installation condition is vertical installation; Inlet boundary conditions: inlet total mass flow rate: 440Kg / h, inlet gas phase velocity: 4.57m / s, inlet liquid phase velocity: 0.53m / s, liquid phase volume fraction: 0.1032, turbulence intensity: 5%, hydraulic diameter value: 14.4mm.

[0028] After the simulation analysis, it can be seen that the flow uniformity STD0 of the Venturi flow divider as the control sample is 0.1839g / s, the flow uniformity STD' of the pressure drop type flow divider (reference sample) is 0.4650g / s, and the flow uniformity STD01 of the flow divider sample to be analyzed is 0.4452g / s. The evaluation index of the flow uniformity is the refrigerant mass flow standard deviation STD at the outlet of each branch pipe, and its formula is as follows:

[0029] JPEG2026034369000002.jpg23170

[0030] As can be seen from the above simulation analysis, the flow uniformity STD01 of the flow divider sample to be analyzed in the vertically installed state is slightly better than the flow uniformity STD' of the conventional general pressure drop type flow divider (reference sample). Therefore, the flow divider sample to be analyzed has market application value, but its flow uniformity is inferior to that of the Venturi flow divider (reference sample).

[0031] Based on the simulation analysis results, and after observing the changes in the refrigerant flow conditions in the flow divider sample to be analyzed, it was found that the refrigerant flow velocity and flow rate at the inner bottom wall of the mixing chamber were all low, the degree of mixing collision between the gas and liquid two-phase refrigerant was not severe, and gas and liquid two-phase separation occurred during the process of the refrigerant being reflected and mixed and flowing back along the mixing chamber to be distributed to the discharge side, and some liquid refrigerant also accumulated on the inner bottom wall of the main body's liquid supply end.The observation results revealed that the unreflected refrigerant and the gas-liquid two-phase re-separation during the distribution of the reflected refrigerant were the main factors affecting the flow division uniformity of the flow divider sample to be analyzed.

[0032] In view of this, the inventors designed the following optimization scheme based on the cooling diverter of this embodiment shown in Figure 2. Based on the outer diameter D0 of the main body of the liquid supply pipe, six optimization groups were selected, each of which included six sub-optimization groups with different minimum inner diameters D1 of the throttle section. Each sub-optimization group then generated eight optimized samples based on the inner diameter D2 of the opening of the mixing chamber, resulting in a total of 288 optimized samples. To clearly illustrate the structure of each component, Figure 2 is a structural schematic diagram of an optimized sample with four branch pipes. The structures of the other optimized samples are essentially the same as Figure 2, differing only in the number of branch pipes.

[0033] ANSYS software was used to perform CFD simulation analysis on the 288 optimized samples. The Venturi flow divider shown in Figure 1 was still used as the control sample, and the simulation analysis results of all the optimized samples were compared. Specifically, because the inner diameter D0 of the main body of the liquid feed pipe differed, each optimization group corresponded to a control sample to ensure that the structural parameters were consistent. That is, the outer diameter D0 of the main body of the liquid feed pipe and the number of branch pipes of each control sample were the same as those of the 48 optimized samples in this sub-optimization group. An optimized sample whose flow distribution uniformity STD was equal to or less than the control sample's flow distribution uniformity STD0 was considered a passing optimized sample; otherwise, it was considered a failing optimized sample.

[0034] The turbulence model, network parameters and operating conditions selected for the CFD simulation analysis are all the same as above.

[0035] The boundary conditions for each optimization group are shown in the table below.

[0036] JPEG2026034369000003.jpg69170

[0037] In the first optimization group, the outer diameter D0 of the main body of the liquid supply pipe is 6.35 mm, the number of branch pipes is two, and the inner diameter D3 of the branch pipe is 3 mm. The sub-optimization groups included in this optimization group have the minimum inner diameter D1 of the constriction section set to 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 based on the inner diameter D2 of the opening of the mixing chamber. For example, in the sub-optimization group with D1 = 2.1 mm, the values ​​of D2 are 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.

[0038] In the second optimization group, the outer diameter of the main supply pipe D0 is 9.52 mm, the number of branch pipes is four, and the inner diameter of the branch pipes D3 is 3 mm. In the sub-optimization groups included in this optimization group, the minimum inner diameters of the constriction section D1 are 2.7 mm, 4.3 mm, 5.91 mm, 7.5 mm, 9.0 mm, and 10.6 mm. In each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 of the opening of the mixing chamber.

[0039] In the third optimization group, the outer diameter of the main supply pipe D0 is 12.7 mm, the number of branch pipes is 8, and the inner diameter of the branch pipes D3 is 3 mm. In the sub-optimization groups included in this optimization group, the minimum inner diameter of the constriction section D1 is 3.3 mm, 5.7 mm, 8.1 mm, 10.5 mm, 12.1 mm, and 13.6 mm. In each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 of the opening of the mixing chamber.

[0040] In the fourth optimization group, the outer diameter of the main supply pipe D0 is 16 mm, the number of branch pipes is 11, and the inner diameter of the branch pipes D3 is 3 mm. In the sub-optimization groups included in this optimization group, the minimum inner diameter of the constriction section D1 is 4.5 mm, 7.2 mm, 9.9 mm, 12.5 mm, 15.2 mm, and 16 mm. In each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 of the opening of the mixing chamber.

[0041] In the fifth optimization group, the outer diameter of the main supply pipe D0 is 22 mm, the number of branch pipes is 22, and the inner diameter of the branch pipes D3 is 3 mm. In the sub-optimization groups included in this optimization group, the minimum inner diameters of the constriction sections D1 are 8.2 mm, 9.9 mm, 13.6 mm, 17.2 mm, 20.9 mm, and 22.6 mm. In each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 of the opening of the mixing chamber.

[0042] In the sixth optimization group, the outer diameter of the main supply pipe D0 is 28 mm, the number of branch pipes is 24, and the inner diameter of the branch pipes D3 is 3 mm. In the sub-optimization groups included in this optimization group, the minimum inner diameters of the constriction section D1 are 10.3 mm, 12.6 mm, 17.3 mm, 21.9 mm, 26.6 mm, and 28.9 mm. In each sub-optimization group, eight optimized samples are selected based on the inner diameter D2 of the opening of the mixing chamber.

[0043] The simulation data of the 288 optimized samples are summarized and shown in Table 1.

[0044] The data in Table 1 is analyzed to determine the influence on the uniformity of the divided flow caused by the outer diameter D0 of the main body of the liquid feed pipe, the minimum inner diameter D1 of the restriction, and the inner diameter D2 of the opening of the mixing chamber.

[0045] First, we observed the effect of the minimum inner diameter D1 of the throttle section on the flow distribution uniformity for each optimization group. We found that, regardless of the change in the inner diameter D2 of the mixing chamber opening, the flow distribution uniformity of the optimized samples was inferior to that of the control sample, as seen in the data for the first and sixth optimization groups. In the cooling flow distributor of this embodiment, the throttle section 22 is designed to accelerate the refrigerant entering the main body 21 by reflecting it at a high velocity onto the inner bottom wall of the mixing chamber 106. However, a minimum inner diameter D1 that is too small significantly reduces the input refrigerant mass flow rate. The smaller the refrigerant mass flow rate, the weaker the degree of reflection and collision at the inner bottom wall of the mixing chamber 106, further severely affecting the mixing strength of the gas-liquid two-phase refrigerant. While a throttle can increase the refrigerant flow velocity and improve the reflection mixing strength, a minimum inner diameter D1 that is too small results in the refrigerant mass flow rates being too close, reducing the degree of mixing of the two-phase flow. If the minimum inner diameter D1 of the throttle portion is too small, the refrigerant pressure loss will be too large, which will worsen the separation of the gas and liquid phases.

[0046] Therefore, further analysis of the relationship between the minimum inner diameter D1 of the throttle portion of each optimized group and the feed pipe mounting hole D0 reveals that the flow distribution uniformity of the optimized sample is superior to that of the control real sample only when the minimum inner diameter D1 of the throttle portion of each sub-optimized group satisfies 0.45D0≦D1≦0.95D0. Preferably, D1 / D0 may be set to a ratio of 0.5, 0.55, 0.6, 0.65, 0.7, 0.78, etc., but the present invention is not limited thereto. In other embodiments, D1 / D0 may be other ratios between 0.45 and 0.95.

[0047] Then, based on the fixed minimum inner diameter D1 of the throttle section, the effect of the inner diameter D2 of the mixing chamber opening of each sub-optimization group on flow division uniformity is analyzed. In the cooling flow divider of this embodiment, the high-pressure, high-velocity refrigerant inevitably expands and decelerates due to inertia after entering the mixing chamber 106 with an enlarged cross-section through the throttle section 22. The decrease in flow velocity affects the kinetic energy of the refrigerant. If the refrigerant flow velocity is too low, some refrigerant will not be able to be reflected from the inner bottom wall of the mixing chamber 106, resulting in insufficient mixing of the gas-liquid two-phase flow and poor atomization of the two-phase flow, resulting in dispersed flow. In gas-liquid two-phase refrigerant, the flow of the liquid refrigerant is mainly dependent on inertia and gravity. If the refrigerant flow rate is too low, the inertia of the liquid refrigerant will be smaller than its gravity. As a result, the liquid phase in the mixed refrigerant will be easily separated from the gas phase again due to the influence of gravity. Furthermore, during the distribution process after mixing, some of the liquid refrigerant cannot be distributed to the discharge side and will gradually accumulate on the inner wall of the main body's liquid supply end, further affecting the uniformity of the distribution.

[0048] As can be seen from the above analysis, both the expansion and deceleration of the refrigerant are caused by changes in the cross-section of the refrigerant flow path. Therefore, the relationship between the cross-sectional area S2 at the opening of the mixing chamber 106 and the cross-sectional area S1 at the downstream end of the throttle section 22 for each sub-optimization group in Table 1 was analyzed. The analysis results showed that when the ratio of the cross-sectional area S2 at the opening of the mixing chamber 106 to the cross-sectional area S1 at the downstream end of the throttle section 22 (hereinafter referred to as the cross-sectional area ratio S2 / S1 for convenience of explanation) is within a certain range, the flow distribution uniformity STD of the optimized sample is significantly better than the flow distribution uniformity STD0 of the control sample. However, if the cross-sectional area S2 / S1 is too large or too small, the flow distribution uniformity will be reduced.

[0049] The upper and lower limit values ​​of the cross-sectional area ratio S2 / S1 for each sub-optimization group in Table 1 are compiled to form Table 2. Further analysis of the data in Table 2 reveals that the upper limit values ​​of the cross-sectional area ratio S2 / S1 for each optimization group all have a certain relationship with the feed pipe mounting hole D0 for that optimization group. Since the lower limit is limited to S2 / S1 > 1, analysis of the optimization group with D0 between 6.35 mm and 16 mm reveals that there is also a certain relationship between the lower limit value of the cross-sectional area ratio S2 / S1 and the feed pipe mounting hole D0. Based on this analysis, the data in Table 2 is fitted and calculated, and the following refrigerant expansion model is constructed: (S2 / S1) min =-1.26×ln(D0)+4.572, (S2 / S1) max =-1.75×ln(D0)+9.757, (S2 / S1) min is the minimum value of the cross-sectional area ratio S2 / S1, (S2 / S1) max is the maximum value of the cross-sectional area ratio S2 / S1.

[0050] By establishing a refrigerant expansion model, the cooling diverter of this embodiment can accurately set the range of the cross-sectional area ratio S2 / S1 based on the outer diameter D0 of the feed pipe main body. The upper limit of the cross-sectional area ratio S2 / S1 limits the volumetric expansion and deceleration of the refrigerant after it enters the mixing chamber 106. This ensures that the input refrigerant has sufficient kinetic energy to be reflected by the inner bottom wall of the mixing chamber 106 even after expansion and deceleration. This intensifies the collision of the gas-liquid two-phase refrigerant, atomizing the two-phase refrigerant into a dispersed flow pattern. Furthermore, the high refrigerant flow velocity further increases the inertial force of the liquid refrigerant after mixing, making it stronger than gravity. This prevents the liquid refrigerant from agglomerating and sedimenting during the distribution process after mixing. This ensures that the two-phase refrigerant maintains a uniformly mixed dispersed flow pattern during the distribution process after mixing, further achieving uniform refrigerant distribution.

[0051] The minimum space required for refrigerant expansion and mixing and the minimum cross-sectional area required for refrigerant distribution after mixing are determined by the lower limit of the cross-sectional area ratio S2 / S1. Specifically, the mixing chamber 106 provides a collision space for the reflected two-phase refrigerant. If this space is too small (especially in product specifications with a small outer diameter of the feed pipe main body), the two-phase refrigerant will not be able to mix sufficiently and will be rapidly distributed among the multiple partition plate holes 31 under the action of high pressure within the chamber, resulting in insufficient refrigerant mixing and evenly distributed refrigerant distribution. Furthermore, the high-pressure, high-velocity refrigerant entering the mixing chamber 106 occupies part of the cross-sectional area at the opening of the mixing chamber 106. The small opening cross-sectional area S2 of the mixing chamber 106 makes it difficult for the reflected and mixed refrigerant to pass through and be distributed among the multiple partition plate holes 31. In other words, the effective distribution flow path at the opening of the mixing chamber 106 is too narrow. A narrow distribution channel results in a sudden increase in the flow resistance of the refrigerant. This increase in flow resistance affects the flow velocity of the refrigerant entering through the feed pipe mounting hole 104 and further affects the refrigerant's reflective mixing effect. Furthermore, the increase in flow resistance results in excessive pressure loss in the flow divider, which further impacts the performance of the heat exchanger assembly connected to the downstream end of the flow divider. Specifically, if the pressure loss in the flow divider is too large, the average evaporation temperature of the evaporator is too low. If the condensation temperature remains unchanged, the average evaporation temperature decreases, increasing the compression ratio of the compressor and further increasing compressor power. Therefore, the cooling flow divider of this embodiment determines the lower limit of the cross-sectional area ratio S2 / S1 based on a refrigerant expansion model, and further determines the minimum diameter D2 at the opening of the mixing chamber 106. min is determined, providing an appropriate mixing space for the refrigerant after reflection, controlling the distribution flow path of the refrigerant, and further improving the uniformity of the refrigerant distribution flow.

[0052] After obtaining the simulation analysis data of the optimized samples, corresponding optimized real samples are fabricated based on the structural parameters of some of the optimized samples to verify the reliability of the simulation analysis data. The optimized real samples are then subjected to operational tests to determine their actual flow distribution uniformity STD. Specifically, three sub-optimization groups are selected from each optimization group, and three optimized samples are selected from each sub-optimization group to fabricate real samples for operational performance tests. There are a total of six real test groups, each of which includes three sub-real test groups, and each sub-real test group includes three optimized real samples, for a total of 54 optimized real samples. Similarly, the control samples corresponding to each optimization group are used as control real samples for operational performance tests.

[0053] A flow divider performance test device was used to conduct a flow division uniformity test on 54 optimized real samples. Figure 4 shows the principle of the flow divider performance test device used in this embodiment, which includes multiple ball valves 61, 24 sets of finned evaporators 62, check valves 63, L-finned evaporators 64, four-way switching valves 65, compressors 66, economizers 67, economizer electronic expansion valves 68, evaporator main electronic expansion valves 69, evaporator auxiliary electronic expansion valves 70, gas-liquid flow dividers 71, sleeve evaporators 72, reservoirs 73, total discharge temperature sensors 74, total return temperature sensors 75, thermal buffer water tanks 76, and cooling towers 77. The evaporator main electronic expansion valves 69 and the evaporator auxiliary electronic expansion valves 70 are connected in parallel and then connected to the feed pipe 2 of the optimized real sample 80, and the multiple branch pipes 4 of the optimized real sample 80 are connected to all or part of the finned evaporators 62. The test principle of this shunt performance test device is as follows: [1] Device structure and layout concept This diverter performance test device is composed of a combination of 24 loop evaporators, each of which uses a fin structure (i.e., a fin evaporator). The heat exchange area, heat exchange tubes, and fin configuration of each loop evaporator are perfectly matched, and each loop is uniformly arranged 360 degrees around the blower. Therefore, by default, the air flow rate through each loop evaporator is the same, and similarly, it can be roughly considered that the heat exchange rate of each loop evaporator is the same.

[0054] [2] Test theory calculations It is known that the heat exchange rate of an evaporator can be calculated from the enthalpy values ​​of the fluid at the inlet and outlet of the evaporator using the theoretical formula: Heat exchange rate = refrigerant mass flow rate x (outlet enthalpy value - inlet enthalpy value). Based on this, the difference in refrigerant mass flow rate of each loop evaporator can be back-calculated to obtain the difference percentage, which can be used to judge the theoretical suitability of the test element.

[0055] For single-phase refrigerants, the enthalpy value can be obtained by looking up a temperature table and a pressure table.

[0056] Because the refrigerant at the evaporator inlet is in a two-phase gas-liquid state, its enthalpy value cannot be calculated. This problem can be solved by adding a throttle component. This test device includes two electronic expansion valves (throttle components), including the evaporator main electronic expansion valve 69 and the evaporator auxiliary electronic expansion valve 70. By utilizing the enthalpy throttling characteristics of the throttle components, the inlet enthalpy values ​​of the two electronic expansion valves (throttle components), including the evaporator main electronic expansion valve 69 and the evaporator auxiliary electronic expansion valve 70, are measured. The inlet enthalpy values ​​of the two electronic expansion valves (throttle components) are equal to the inlet enthalpy value of the evaporator. By ensuring that the refrigerant before the two electronic expansion valves (throttle components) is a pure liquid (subcooled liquid), the enthalpy value can be obtained by looking up the corresponding refrigerant property table based on the temperature before the two electronic expansion valves (throttle components).

[0057] If it is ensured that the outlets of each loop evaporator are all superheated gas refrigerant, the enthalpy value can be obtained by looking up a physical property table for the corresponding refrigerant based on the refrigerant temperature and pressure at the outlets.

[0058] [3] Isoenthalpy squeezing As shown in Figure 4A, the abscissa h is the process enthalpy value, the ordinate lg(P) is the logarithmic value of pressure, and the 4→1 process is the enthalpy change curve when the refrigerant flows through the electronic expansion valve (throttle component). Because there is no heat exchange with the outside during the process of flowing through the throttle component, and the refrigerant does not exchange heat with any fluid, the 4→1 process is perpendicular to the abscissa, and it can be seen that the enthalpy value does not change during the throttle process.

[0059] [4] Enthalpy Value Creation The total discharge temperature sensor 74 and total return temperature sensor 75 are used to measure the temperatures at the inlet and outlet of the evaporator during testing, and the corresponding saturation pressure is calculated using enthalpy value calculation software (e.g., a refrigerant calculator), thereby determining the subcooled (superheated) state of the refrigerant at the inlet and outlet. By inputting the inlet and outlet temperature and pressure values ​​measured during testing, the corresponding enthalpy value can be obtained, and the difference in refrigerant mass flow rate of each loop evaporator can be back-calculated based on the theoretical calculation formula above.

[0060] The data of the 54 optimized actual samples after operation tests using the flow diverter performance test device is summarized in Table 3. In Table 3, the flow diversion uniformity of the operation tests of the optimized actual samples is represented by the STD test, and the flow diversion uniformity of the control actual sample is represented by the STD0 test. Note that for the actual test group with an outer diameter D=16 mm of the feed pipe main body, the operation test data STD01 test of the optimized actual sample of the air conditioning flow diverter structure proposed by the inventors in China CN216204506U (i.e., the actual sample of the flow diverter to be analyzed) is further summarized for reference.

[0061] As can be seen from Table 3, <1> In each physical test group, D1 of the first sub-physical test group is less than 0.45D0, and the flow distribution uniformity STD tests of all optimized physical samples in this sub-test are all inferior to the flow distribution uniformity STD0 test of the control physical sample.

[0062] <2> In the sub-physical test group where 0.45D0 ≤ D1 ≤ 0.95D0 (including the second sub-physical test group and the third sub-physical test group), S2 > S1 and only the optimized physical samples whose area ratio S2 / S1 satisfies the refrigerant expansion model have their flow distribution uniformity STD test superior to the flow distribution uniformity STD0 test of the corresponding control physical sample.

[0063] <3> In the physical test group with D0 = 16 mm, the optimized physical samples with flow distribution uniformity superior to the control physical sample have their flow distribution uniformity STD test far superior to the flow distribution uniformity STD01 test of the physical sample of the flow divider to be analyzed before improvement.

[0064] The operation test results of 54 optimized physical samples are consistent with the simulation analysis results, indicating that in this embodiment, it is reliable to perform sample optimization analysis with ANSYS simulation analysis software to establish the refrigerant expansion model.

[0065] The operation performance test results of the optimized physical samples verify the reliability of the simulation analysis. Furthermore, it is necessary to verify the obtained refrigerant expansion model by fitting and determine the reliability of this model at different outer diameters D0 of the liquid supply pipe main body. Specifically, when the outer diameter D0 of the liquid supply pipe main body is selected as 19 mm, and calculated based on the expansion coefficient k = -1.26 and b = 4.572, the lower limit value (S2 / S1)min of the cross-sectional area ratio S2 / S1 is 0.86. When calculated based on the expansion coefficient k = -1.75 and b = 9.757, the upper limit value (S2 / S1)max of S2 / S1 is 4.60. Since the cross-sectional area ratio S2 / S1 must also satisfy > 1, verification samples that pass within the range of 1 < S2 / S1 ≤ 4.60 are designed, and verification samples that do not pass within the range of S2 / S1 > 4.60 are designed to form the first verification group.

[0066] Similarly, the outer diameter D0 of the liquid supply pipe main body is selected as 32 mm. As a result of calculation based on the expansion coefficients k = -1.26 and b = 4.572, the lower limit value (S2 / S1)min of the cross-sectional area ratio S2 / S1 is 0.21. As a result of calculation based on the expansion coefficients k = -1.75 and b = 9.757, the upper limit value (S2 / S1)max of S2 / S1 is 3.69. Since the cross-sectional area ratio S2 / S1 > 1 must also be satisfied, verification samples that passed within the range of 1 < S / S1 ≤ 3.69 are designed, and verification samples that did not pass within the range of S2 / S1 > 3.69 are designed to form a second verification group.

[0067] The outer diameter D0 of the liquid supply pipe main body in the first verification group is 19 mm. The minimum inner diameters D1 of the throttle portions in the sub-verification groups included in the first verification group are 8.6 mm, 10.5 mm, 12.4 mm, 14.3 mm, 16.2 mm, and 18.1 mm, and eight verification samples are selected from each sub-verification group. Six of them are verification samples that passed, and their area ratio satisfies 1 < S2 / S1 ≤ 4.60. The other two are verification samples that did not pass, and their area ratio satisfies S2 / S1 > 4.60.

[0068] The outer diameter D0 of the liquid supply pipe main body in the second verification group is 32 mm. The minimum inner diameters D1 of the throttle portions in the sub-verification groups included in the first verification group are 14.4 mm, 17.6 mm, 20.8 mm, 24.0 mm, 27.2 mm, and 30.9 mm, and eight verification samples are selected from each sub-verification group. Six of them are verification samples that passed, and their area ratio satisfies 1 < S2 / S1 ≤ 3.69. The other two are verification samples that did not pass, and their area ratio satisfies S2 / S1 > 3.69.

[0069] ANSYS simulation analysis software is adopted to perform simulation analysis on 96 verification samples included in the two verification groups. The turbulent flow model, network parameters, and operating conditions selected for the simulation analysis are all the same as above.

[0070] The boundary conditions for each validation group are shown in the table below.

[0071] JPEG2026034369000004.jpg33170

[0072] The simulation data for the 96 optimized samples are summarized and shown in Table 4.

[0073] Similarly, a comparison analysis was performed on all verification samples in each verification group, using the simulation analysis data of a Venturi flow divider with the same feed pipe main body inner diameter D0, the same number of branch pipes, and the same branch pipe size as the control sample. The comparison analysis revealed that the flow distribution uniformity STD of the six successful verification samples in each sub-verification group was superior to the flow distribution uniformity STD0 of the corresponding control sample, while the flow distribution uniformity STD of the two unsuccessful verification samples was inferior to the flow distribution uniformity STD0 of the control sample. The verification data in Table 4 reflect the reliability of the refrigerant expansion model established by fitting in this example.

[0074] By establishing a highly reliable refrigerant expansion model, the cooling diverter of this embodiment can accurately calculate the cross-sectional area ratio S2 / S1 based on the specified outer diameter D0 of the feed pipe main body during product design. Furthermore, by matching the relationship between the outer diameter D0 of the feed pipe main body and the minimum inner diameter D1 of the throttle section, the cross-sectional area S2 of the mixing chamber opening can be calculated to achieve excellent flow uniformity. Furthermore, the inner diameter D2 of the mixing chamber opening required for product design and prototyping can be determined. This provides selection guidance for product series design, significantly shortens the R&D cycle for cooling diverter products, and reduces R&D costs. Preferably, after determining the upper limit (S2 / S1)max and the lower limit (S2 / S1)min, the area ratio S2 / S1 can be designed closer to the upper limit (S2 / S1)max if the space within the storage chamber 103 allows.

[0075] In this embodiment, the depth H1 of the mixing chamber satisfies 2 mm≦H1≦22 mm, and the depth H1 of the mixing chamber 106 is the vertical distance from the end face of the opening of the mixing chamber to the center of the inner bottom wall of the mixing chamber. Preferably, the depth H1 of the mixing chamber 106 is set to any one of 8 mm, 10 mm, 12 mm, 14 mm, 16 mm, 18 mm, and 20 mm, but the present invention is not limited thereto. In other embodiments, the depth H1 of the mixing chamber may be another value within the range of 2 mm≦H1≦22 mm. Setting the depth H1 of the mixing chamber allows control of the internal volume of the mixing chamber 106 and further promotes the degree of mixing of the two-phase refrigerant.

[0076] 2 to 5, the mixing chamber 106 is a cylindrical chamber with approximately equal lumen diameters, but the present invention is not limited thereto. In other embodiments, the shape of the lumen of the mixing chamber 106 may be a combination of one or more of a truncated cone (shown in FIG. 6A), a cone (shown in FIG. 6B), a prism, a truncated pyramid, or a pyramid, or a combination of any of the above lumen shapes with a partial sphere, such as a combination of a cylindrical shape and a partial sphere (shown in FIG. 6C), a combination of a truncated cone and a partial sphere, or a combination of a prism and a partial sphere.

[0077] In this embodiment, as shown in FIG. 2 , the vertical distance H0 from the end face of the discharge end of the supply pipe 2 to the end face of the opening of the mixing chamber 106 is set to satisfy the relationship 3.5 mm≦H0≦11.5 mm. Preferably, the vertical distance H0 is set to 5 mm or 8 mm, but the present invention is not limited thereto. In other embodiments, the vertical distance H0 may be other values ​​within the range 3.5 mm≦H0≦11.5 mm. The vertical distance H0 affects the distribution flow path from the opening of the mixing chamber 106 to the partition plate holes 31. If the vertical distance H0 is too small, the distribution flow path becomes too narrow, increasing the flow resistance of the refrigerant to the partition plate holes 31. In particular, when the discharge end of the supply pipe 2 extends into the main body 1, the refrigerant forms a vortex between the outer wall of the inlet end of the supply pipe 2 and the inner wall of the main body supply end 101. This vortex further pushes the distribution flow path, rapidly increasing the flow resistance. In view of this, in this embodiment, the vertical distance H0 is set to 3.5 mm or greater to increase the distribution flow path from the opening of the mixing chamber 106 to the partition plate holes 31, promoting smooth distribution of the refrigerant into the multiple partition plate holes 31 after mixing and reducing pressure loss during distribution. If the vertical distance H0 is too large, the refrigerant's injection stroke toward the inner bottom wall of the mixing chamber 106 increases, causing the refrigerant to lack sufficient kinetic energy for reflection, further worsening the degree of reflective mixing. Therefore, the upper limit of the vertical distance H0 must be controlled to ensure that the refrigerant has sufficient kinetic energy to collide with the inner bottom wall of the mixing chamber after entering the main body 1 and achieve reflective mixing.

[0078] In this embodiment, the throttle portion 22 is a throttle hole portion whose inner diameter is almost constant and smaller than the inner diameter of the main body portion 21, and the minimum inner diameter D1 of the throttle portion is the inner diameter of the throttle portion. Specifically, in this embodiment, the throttle portion 22 is integrally molded at the downstream end of the main body portion 21, but the present invention is not limited to this. In another embodiment, as shown in FIG. 7 , the refrigerant flow divider may further include a throttle hole plate 2A provided within the main body portion 21. The throttle hole plate 2A has a throttle hole 2A1 whose inner diameter is smaller than the outer diameter of the main body portion, and the throttle portion 2A1 is formed in the throttle hole 2A1. The throttle hole may be a straight hole whose inner diameter is almost constant or a collecting hole whose inner diameter gradually decreases in the flow direction of the refrigerant in the liquid supply pipe. In this configuration, the inner diameter of the throttle hole 2A1 at the downstream end is the minimum inner diameter D1 of the throttle portion. Alternatively, as shown in Fig. 8A, the constriction 22 is configured in a Venturi tube section, and the minimum inner diameter D1 of the constriction is the inner diameter of the straight throat portion of the Venturi tube section. Furthermore, in another embodiment, as shown in Fig. 8B, the downstream of the liquid supply pipe main body 21 may be configured to have multiple Venturi tube sections, and in this configuration, the inner diameter of the straight throat portion of the Venturi tube section closest to the liquid supply pipe mounting hole 104 is the minimum inner diameter D1 of the constriction.

[0079] In this embodiment, the end of the restriction portion 22 is welded to the interior of the supply pipe mounting hole 104 and is the discharge end of the supply pipe 2, but the present invention is not limited to this. In other embodiments, the supply pipe may be integrally formed with the main body in the supply pipe mounting hole.

[0080] In this embodiment, the mixing chamber 106 is a recess formed in the partition plate 3, in an area facing the supply pipe mounting hole 104, extending toward the main body discharge end 102. The partition plate 3 also divides the storage chamber 103 within the main body into a first chamber 1031 and a second chamber 1032, and the high-velocity refrigerant discharged through the partition plate hole 31 expands and mixes again in the second chamber 1032. Furthermore, as shown in FIG. 9 , the vertical distance H2 between the outer bottom wall of the mixing chamber 106 and the inner bottom wall of the main body discharge end is set to satisfy the relationship 0≦H2≦3 mm. The internal volume of the second chamber 1032 can be controlled by setting the vertical distance H2 according to the determined depth H1 of the mixing chamber 106 and the inner diameter of the main body. This effectively prevents re-gas-liquid separation of the dispersed refrigerant due to the second chamber 1032 being too large, and further improves flow uniformity. Preferably, H2 is set to 0, i.e., the outer bottom wall of the mixing chamber 106 abuts against the inner bottom wall of the main body drain end, but the present invention is not limited thereto. In another embodiment, the outer bottom wall of the mixing chamber 106 does not abut against the inner bottom wall of the main body drain end, in which case the vertical distance H2 is 3 mm or less.

[0081] In this embodiment, as shown in FIG. 1 , the main body 1 includes a cylindrical body 11, an end cap 12, and a liner 13. The cylindrical body 11 is integrally molded and has an open end. A plurality of flow-division holes are formed in the bottom of the cylindrical body 11. The liner 13 is tightly welded to the inner bottom surface of the cylindrical body 11. The liner 13 and the cylindrical body 11 together form the end wall of the discharge end of the main body 1. The inner surface of the liner 13 forms the inner bottom wall of the discharge end of the main body. The vertical distance H2 is the distance from the outer bottom wall of the mixing chamber 106 to the inner surface of the liner 13. The liner 13 has a plurality of liner holes coaxial with the plurality of flow-division holes. The liner holes and the corresponding flow-division holes form branch pipe holes 105, but the present invention is not limited to this.

[0082] In this embodiment, the partition plate and the inner bottom wall of the main body's drain end enclose the second chamber, but the present invention is not limited to this. In other embodiments, the main body does not need to have a second chamber, and the multiple branch pipes may extend into the main body and directly connect to the corresponding partition plate holes, i.e., the refrigerant may be directly distributed from the partition plate holes into the multiple branch pipes. Furthermore, in this embodiment, the molding method of the mixing chamber is not limited.

[0083] Example 2 This embodiment is basically the same as the first embodiment and its variations, but differs in the way the mixing chamber 106 is formed and the structure of the main body 1.

[0084] Specifically, as shown in FIG. 10 , in this embodiment, the main body 1′ includes an end cap 12′ and a liner 13′. A liquid supply pipe mounting hole is formed at one end of the end cap 12′ (not shown in FIG. 10 because the liquid supply pipe 2 is inserted into the liquid supply pipe mounting hole). A liner 13′ is embedded in the other end of the end cap 12′, and multiple branch pipe holes 105′ are formed in the liner 13′. A mixing chamber 106′ is a recess formed in the liner 13′, and the dispersed refrigerant reflected and mixed in the mixing chamber 106′ is directly distributed into the multiple branch pipe holes 105′. While the number of liners in both FIGS. 2 and 10 is one, this is not a limitation of the present invention. In other embodiments, multiple liners may be provided, and the multiple liners may be stacked and then welded to the inner or outer wall of the cylinder or fitted into the open end of the end cap.

[0085] Example 3 This embodiment is basically the same as the first embodiment and its variations, but differs in the structure of the liquid supply pipe mounting hole 104 and the liquid supply pipe 4.

[0086] As shown in FIG. 11 , in this embodiment, feed pipe attachment hole 104 is a straight through-hole without a welded attachment portion, and constricted portion 22 is a Venturi tube portion. The minimum inner diameter D1 of the constricted portion is the inner diameter of the straight throat portion of the Venturi tube portion. The Venturi tube portion has a divergent portion 23 located downstream of the minimum inner diameter D1 of the constricted portion and having an arc-shaped outer wall generatrix. Divergent portion 23 projects into the storage chamber through feed pipe attachment hole 104 and is tightly welded to the inner wall of feed end 101 of the main body so that its outer wall is in surface contact. The definition is as follows: An extension of the inner wall generatrix of the straight throat portion of the Venturi tube portion intersects with an imaginary extension plane of feed pipe attachment hole 104 on the inner wall of feed end 101, and the cross section of feed pipe 2 passing through position K is the end face of the discharge end of feed pipe 2. Similarly, in this embodiment, the vertical distance H0 from the end face of the discharge end of the liquid supply pipe to the end face at the opening of the mixing chamber 106 is set to satisfy 3.5 mm≦H0≦11.5 mm.

[0087] However, the present invention is not limited to the specific structure of the liquid feed pipe attachment hole. In another embodiment, when the liquid feed pipe attachment hole has a straight weld attachment portion, the liquid feed pipe may be configured to include a straight portion (e.g., the throat of the Venturi tube portion) connected to the straight weld attachment portion and a divergent portion located downstream of the minimum inner diameter of the constriction portion and having an outer wall generatrix that forms an arc-shaped curve. Alternatively, when the constriction portion is a constriction hole portion whose inner diameter is substantially constant and smaller than the outer diameter of the main body portion, a divergent portion whose outer wall generatrix forms an arc-shaped curve is provided downstream of the constriction hole portion so that the outer wall of the divergent portion can be tightly welded to the inner wall of the liquid feed end of the main body so as to be in surface contact with the inner wall of the liquid feed end.

[0088] In this embodiment, the number of branch pipes 4 is 10, but the present invention is not limited to this number.

[0089] Example 4 This embodiment is basically the same as the first embodiment and its variations, but differs in the structure of the branch pipe 4.

[0090] 12 , each branch pipe 4 includes a first pipe section 41 and a second pipe section 42 located downstream of the first pipe section 41 and having an inner diameter smaller than that of the downstream end of the first pipe section 41, and the difference Δd between the inner diameter d11 of the downstream end of the first pipe section 41 and the inner diameter d12 of the downstream end of the second pipe section 42 is in the range of 0.1 mm≦Δd≦3.5 mm. The first pipe section 41 is formed with a reflecting section 411 that bends and extends to one side of the center line of the branch pipe hole 105, and the axis of the upstream end of the reflecting section 411 intersects the axis of the downstream end of the reflecting section at an angle θ that is 90°≦θ≦175°. The reflecting section 411 causes the axis of the second pipe section 42 to intersect with the center line of the branch pipe hole 105.

[0091] The first pipe section 41, with its larger inner diameter, increases the refrigerant flow rate distributed to each branch pipe 4 and reduces refrigerant distribution resistance, while the second pipe section 42, with its smaller inner diameter, improves the refrigerant flow velocity. Furthermore, the inner diameter difference Δd accurately controls the refrigerant pressure drop acceleration of the second pipe section 42, preventing excessive refrigerant pressure loss due to over-acceleration. Due to the configuration of the reflecting section 411, the second pipe section 42 is not coaxial with the centerline of the branch pipe. When downstream pressure waves vibrate upstream, the reflecting section 411 reflects and absorbs part of the pressure wave, redirecting the remaining pressure wave and quickly attenuating it. This effectively prevents the downstream pressure wave from affecting the upstream branch pipe 105, further improving distribution performance.

[0092] In this embodiment, the branch pipe 4 further includes a third pipe section 43 welded to the first pipe section 41, where the second pipe section 42 is located. Specifically, as shown in FIG. 12 , the third pipe section 43 has a sleeve connection section 431 fitted onto the end of the first pipe section 41 and having a flared structure, and the second pipe section 42, whose inner diameter is tapered, is formed downstream of the sleeve connection section 431, but the present invention is not limited thereto. In other embodiments, the end of the first pipe section 41 may be configured to be fitted onto and welded to the sleeve connection section 431 of the third pipe section 43. In this case, as shown in FIG. 13A , the second pipe section remains positioned downstream of the sleeve connection section 431 and has a tapered inner diameter.

[0093] 13B and 13C are schematic diagrams of the structure of a branch pipe according to another embodiment of the present invention. In FIG. 13B, the first pipe section 41, the second pipe section 42, and the third pipe section 43 are integrally molded. In FIG. 13C, the first pipe section 41 and the second pipe section 42 are integrally molded, and the third pipe section 43 is welded to the end of the second pipe section 42. In other embodiments, the second pipe section may be directly formed at the sleeve connection point between the first pipe section and the second pipe section, and multiple second pipe sections may be provided at the downstream end of the first pipe section. The third pipe section may be provided on a straight pipe or a curved pipe depending on the actual application.

[0094] Example 5 This embodiment is basically the same as the first embodiment and its variations, except that the branch pipe 4 has a different structure.

[0095] As shown in FIG. 14 , in this embodiment, each branch pipe 4 includes a straight connection section 41′, a jet section 42′ located downstream of the straight connection section 41′ and having an inner diameter smaller than the inner diameter d11′ of the straight connection section main body 411′, and a branch section 43′ located downstream of the jet section 42′. The branch section 43′ includes a constant-diameter section 431′ whose inner diameter is approximately the same as the inner diameter d11′ of the straight connection section main body. The difference Δd′ between the inner diameter d11′ of the straight connection section main body and the minimum inner diameter d12′ of the jet section 42′ is 0.1 mm≦Δd′≦3.5 mm. The straight connection section main body 411′ is a pipe section in which the inner and outer diameters of the straight connection section 41′ are almost constant.

[0096] In the refrigerant distributor according to this embodiment, the addition of a jet section 42' to each branch pipe 4 improves the refrigerant flow velocity within the branch pipe 4, satisfying the performance requirements of the downstream heat exchanger assembly. Furthermore, the inner diameter difference Δd' accurately controls the pressure drop acceleration of the refrigerant due to the jet section 42', preventing excessive refrigerant pressure loss due to excessive acceleration. Furthermore, the branch section 43' is provided with a constant diameter section 431', whose inner diameter d13' is nearly equal to the inner diameter d11' of the straight connection section main body 411', reducing the flow resistance of the refrigerant within the branch pipe 4 and achieving a balance between the refrigerant flow velocity and pressure loss.

[0097] In this embodiment, the jet portion 42' is a constriction portion integrally molded upstream of the branch portion 43', and the two are welded to the straight connecting portion 41' after being integrally molded, but the present invention is not limited to this. In another embodiment, as shown in Figure 15A, the jet portion 42' may be integrally molded downstream of the straight connecting portion 41' and then welded to the branch portion 43'. Alternatively, as shown in Figure 15B, the straight connecting portion 41', the jet portion 42', and the branch portion 43' are integrally molded.

[0098] Alternatively, the jet section 42' may be a jet hole plate placed at the sleeve connection portion of the straight connecting section 41' and the branching section 43'. Specifically, as shown in Fig. 15C, the upstream end of the branching section 43' is flared and then sleeve-connected to the downstream end of the straight connecting section 41', and the jet section 42' may be placed at the flared portion of the branching section 43'. Similarly, as shown in Fig. 15D, when the flared portion at the downstream end of the straight connecting section 41' is externally connected to the upstream end of the branching section 43', the jet section 42' may be placed at the flared portion at the downstream end of the straight connecting section 41'.

[0099] Alternatively, as shown in Fig. 15E, the jet portion 42' is provided as a straight pipe portion, and both ends thereof are sleeve-connected to the straight connecting portion 41' and the branching portion 43', respectively. In Figs. 14, 15A, and 15E, the jet portion 42' has a maintenance length L42 during which the inner diameter remains substantially unchanged, but the present invention is not limited thereto.

[0100] The present invention is not limited to any particular configuration of the jet, and any other jet structure that can achieve a reduction in the inner diameter relative to the inner diameter of the straight connection body is within the scope of the present invention.

[0101] As described above, the cooling flow diverter of the present invention achieves refrigerant pressure drop acceleration by providing a throttling section in the supply pipe, whose inner diameter is smaller than that of the main body section. Controlling the inner diameter ratio between the two sections ensures sufficient mass flow at the main body inlet, enhancing the refrigerant collision effect within the mixing chamber and improving refrigerant mixing uniformity. Controlling the minimum inner diameter of the throttling section effectively reduces pressure loss and throttling-related noise within the flow diverter. Accurately controlling the internal volume of the mixing chamber by setting the ratio of the lumen cross-sectional area S2 at the opening of the mixing chamber to the flow cross-sectional area S1 at the minimum inner diameter of the throttling section allows the mixing chamber to provide expansion space and distribution channels for the refrigerant, effectively avoiding problems such as excessive refrigerant expansion and a sudden drop in refrigerant flow rate due to an oversized mixing chamber. The input refrigerant has sufficient kinetic energy to impinge on the inner bottom wall of the mixing chamber and undergo reflective mixing, promoting fully dispersed atomization of the two-phase refrigerant and maintaining a dispersed flow during distribution, further improving flow distribution uniformity.

[0102] Furthermore, based on the refrigerant expansion model established through fitting, the outer diameter D0 of the main body and the minimum cross-sectional area S1 of the restriction section can be determined, and then the lumen cross-sectional area S2 of the opening of the mixing chamber can be calculated. This clarifies the parameter selection of the mixing chamber before testing and provides guidance for pre-development design, significantly improving the development speed of flow divider products, especially the development of series products.

[0103] Although the present invention has been disclosed as above by means of preferred embodiments, it is not used to limit the present invention, and any person skilled in the art can make slight modifications and refinements without departing from the spirit and scope of the present invention, so the protection scope of the present invention should be in accordance with the protection scope set forth in the claims.

[0104] In the table below, the units of the outer diameter D0 of the main body of the liquid supply pipe, the minimum inner diameter D1 of the restriction, and the inner diameter D2 of the opening of the mixing chamber are all in mm, the units of the refrigerant mass flow standard deviations STD, STD0, STD test, STD0 test, and STD01 test at the outlet of each branch pipe of the optimized sample, control sample, optimized real sample, control real sample, and the real sample of the flow divider to be analyzed are all in g / s, and the cross-sectional area ratio S2 / S1 is dimensionless.

[0105] Table 1 JPEG2026034369000005.jpg214170JPEG2026034369000006.jpg219170JPEG2026034369000007.jpg215170

[0106] Table 2 JPEG2026034369000008.jpg71157

[0107] Table 3 JPEG2026034369000009.jpg189170

[0108] Table 4 JPEG2026034369000010.jpg214170

Claims

1. A refrigerant flow divider, a main body including a supply end, a drain end, and a storage chamber that connects the supply end and the drain end, wherein the supply end has a supply pipe mounting hole, the drain end has a plurality of branch pipe holes for dividing the flow, and the storage chamber has a mixing chamber whose opening faces the supply pipe mounting hole and extends recessed toward the main body drain end, and when projected along the axial direction of the main body, the mixing chamber is located on the inner periphery of the plurality of branch pipe holes; a liquid supply pipe, an end of which is connected to the liquid supply pipe mounting hole and which does not extend into the mixing chamber, the liquid supply pipe including a main body portion whose inner diameter and wall thickness do not change substantially, and a throttle portion located downstream of the main body portion and whose inner diameter is smaller than the outer diameter of the main body portion; The minimum inner diameter D1 of the throttle portion and the outer diameter D0 of the main body portion satisfy 0.45D0≦D1≦0.95D0, the lumen cross-sectional area S2 at the opening of the mixing chamber is larger than the flow cross-sectional area S1 at the minimum inner diameter of the throttle portion, 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、 (S2 / S1) ｍｉｎ is the minimum value of S2 / S1, and (S2 / S1) ｍａｘ is the maximum value of S2 / S1.

2. 2. The refrigerant flow divider according to claim 1, wherein the depth H1 of the mixing chamber satisfies 2 mm≦H1≦22 mm, and the depth H1 of the mixing chamber is the vertical distance from the center of the inner bottom wall of the mixing chamber to the end face of the opening of the mixing chamber.

3. 2. The refrigerant flow divider according to claim 1, wherein a vertical distance H0 from an end face of the discharge end of the liquid supply pipe to an end face of the opening of the mixing chamber satisfies 3.5 mm≦H0≦11.5 mm.

4. 2. The refrigerant flow divider according to claim 1, wherein the liquid supply pipe has a divergent section located downstream of the narrowed section at a minimum inner diameter thereof and having an outer wall generatrix that forms an arc-shaped curve, and the outer wall of the divergent section is tightly welded to the inner wall of the liquid supply end of the main body.

5. the throttle portion is a throttle hole portion whose inner diameter is substantially constant and smaller than the outer diameter of the main body portion, Alternatively, the refrigerant flow divider further includes a throttling hole plate provided in the main body, the throttling hole plate having an inner diameter smaller than the outer diameter of the main body, the throttling hole having a throttling portion, the throttling hole being a straight hole whose inner diameter does not change substantially, or a collecting hole whose inner diameter gradually becomes smaller along the flow direction of the refrigerant in the liquid supply pipe, Alternatively, the refrigerant flow divider according to claim 1, wherein the throttle portion is a Venturi tube portion, and the minimum inner diameter D1 of the throttle portion is the inner diameter of a straight throat portion of the Venturi tube portion.

6. the refrigerant flow divider further includes a plurality of branch pipes welded to the plurality of branch pipe holes, each branch pipe including a first pipe portion and a second pipe portion located downstream of the first pipe portion and having an inner diameter smaller than that of a downstream end of the first pipe portion, a difference Δd between an inner diameter d11 of the downstream end of the first pipe portion and an inner diameter d12 of the downstream end of the second pipe portion being 0.1 mm≦Δd≦3.5 mm; 2. The refrigerant flow divider according to claim 1, wherein the first pipe section has a reflector section bent and extending to one side of a center line of the branch pipe hole, and an axis of the upstream end of the reflector section intersects an axis of the downstream end of the reflector section at an angle θ, the angle being 90°≦θ≦175°.

3. The refrigerant flow divider according to claim 1, wherein the reflector section causes the axis of the second pipe section to intersect with the center line of the branch pipe hole.

7. The refrigerant flow divider further includes a plurality of branch pipes welded to the plurality of branch pipe holes, each branch pipe including a straight connection portion, a jet portion located downstream of the straight connection portion and having an inner diameter smaller than that of the straight connection portion body, and a branch portion located downstream of the jet portion; 2. The refrigerant flow divider according to claim 1, wherein the branch section has a constant diameter section whose inner diameter is approximately the same as the inner diameter of the straight connection section body, and the difference Δd' between the inner diameter d11' of the straight connection section body and the minimum inner diameter d12' of the jet section satisfies 0.1 mm≦Δd'≦3.5 mm.

8. 2. The refrigerant flow distributor according to claim 1, further comprising a partition plate provided in the main body accommodating chamber, wherein an area of ​​the partition plate facing the liquid supply pipe mounting hole is recessed toward the liquid discharge end of the main body to form a mixing chamber, and a plurality of partition plate holes are formed on the partition plate plane on the outer periphery of the mixing chamber, the plurality of partition plate holes being distributed in an annular shape around the center line of the main body, and the plurality of partition plate holes distribute the refrigerant mixed in the mixing chamber to the liquid discharge end of the main body.

9. 9. The refrigerant flow divider of claim 8, wherein the partition plate divides the storage chamber into a first chamber near the main body supply end and a second chamber near the main body discharge end, the second chamber being an annular chamber surrounding the mixing chamber, and the vertical distance H2 between the outer bottom wall of the mixing chamber and the inner bottom wall of the main body discharge end satisfies 0≦H2≦3 mm.

10. 9. The refrigerant flow divider according to claim 8, wherein the mixing chamber is a constant diameter chamber having lumen diameters that are approximately the same, or a tapered chamber, a truncated conical chamber, or a hemispherical chamber having an inner diameter that gradually decreases along the inflow direction of the refrigerant, or a combination of a constant diameter chamber, a tapered chamber, a truncated conical chamber, or a hemispherical chamber.

11. A heat exchanger assembly comprising the refrigerant flow divider of claim 1.

12. An air conditioner comprising the heat exchanger assembly of claim 11.

Images