Rectifying nozzle type flow divider and method of designing the same

Abstract

The application discloses a rectifying nozzle type flow divider and a design method thereof, which makes refrigerant flow symmetrically and uniformly in the flow divider and provides a theoretical model for the design of the flow divider. The flow divider comprises a cyclone section, a rectifying section and a distribution section connected in sequence; the cyclone section comprises a cyclone pipe and cyclone blades arranged in the cyclone pipe; the cyclone blades are composed of a center column and a spiral plate mounted on the center column. The distribution section comprises a distribution cavity and a plurality of branch pipes; the inlet of each branch pipe is located in the distribution cavity; each branch pipe is provided with a flow dividing sonic nozzle; the plurality of branch pipes and the flow dividing sonic nozzles are distributed along the circumference of the distribution cavity; and the center lines of the branch pipes and the flow dividing sonic nozzles are parallel to the center line of the distribution cavity. The flow divider can make the refrigerant flow symmetrically and uniformly in the flow divider, improve the heat exchange performance of the evaporator and improve the refrigerating capacity of the evaporator.

Description

TECHNICAL FIELD

[0001] The present application relates to the technical field of regulating system, in particular to a rectifying nozzle type flow divider for regulating gas-liquid two-phase refrigerant into annular flow and a design method thereof. BACKGROUND

[0002] In a refrigeration system, the gas-liquid two-phase refrigerant after pressure reduction by an expansion valve is evenly distributed to each flow of an evaporator through a flow divider. Due to the complex mechanism of gas-liquid two-phase flow, the flow rate and dryness of the gas-liquid two-phase flow in each branch of the evaporator are not the same, resulting in phase separation and uneven heat exchange of each branch, which may even cause liquid damage to the compressor. The uneven distribution of the gas-liquid two-phase flow in the flow divider also leads to the insufficient utilization of the heat exchange area of the evaporator, and the phase separation further aggravates the heat exchange capacity of the evaporator.

[0003] The patent document with the publication number CN104457046A and the invention name of "Rectifying nozzle type equal flow flow divider and refrigeration system" discloses a flow divider. The refrigerant is converted into stable annular flow in the rectifier, and flows through the throat of the flow sound speed nozzle to reach the critical state, realizing uniform liquid supply to each flow path of the evaporator.

[0004] The above flow divider has the following defects: the structure that the distribution cavity, the flow sound speed nozzle and the flow pipe are arranged in a circular array with the center line of the distribution cavity as the center increases the resistance of the refrigerant entering the flow sound speed nozzle, reduces the flow rate of the refrigerant, and the flow sound speed nozzle fails to achieve the effect of pressure reduction and speed increase, and the distribution performance is reduced. At the same time, the above flow divider only gives the external structure of the liquid supply pipe, the fluid rectifier and other flow dividers, and does not give a complete mathematical model for how to determine the size of each part of the structure. The design of the flow divider is only based on experience, and cannot be customized for different systems, which affects the refrigeration efficiency of the refrigeration system. Therefore, a new type of flow divider must be designed and developed, and a complete mathematical model for the design of the flow divider must be provided to provide a basis for the design of the flow divider. SUMMARY

[0005] The purpose of the present application is to provide a rectifying nozzle type flow divider for forming annular flow and uniform flow, reducing the resistance of entering the flow sound speed nozzle, improving the flow rate of the refrigerant, making the refrigerant flow symmetrically in the flow divider, and improving the refrigeration capacity.

[0006] Another purpose of the present application is to provide a design method of the rectifying nozzle type flow divider, which provides a theoretical model for the design of the flow divider.

[0007] The application realizes the above-mentioned purposes by adopting the following technical scheme.

[0008] A rectifying nozzle type flow divider comprises a swirl section, a rectifying section and a distribution section connected in sequence; the swirl section comprises a swirl tube and swirl vanes arranged inside the swirl tube, the swirl vanes are composed of a central column and a spiral plate mounted on the central column, the spiral plate is fixedly mounted between the central column and the inner wall of the swirl tube to make fluid rotate and flow; the distribution section comprises a distribution cavity and a plurality of branch pipes, the inlet of each branch pipe is located in the distribution cavity, a flow dividing sonic nozzle is arranged in the inlet of each branch pipe respectively, a plurality of branch pipes and the flow dividing sonic nozzles thereon are distributed along the circumference with the center of the distribution cavity as the center, the center line of each branch pipe and the flow dividing sonic nozzle thereon is parallel to the center line of the distribution cavity and parallel to the direction of incoming flow; the outlet of the swirl section is connected with the inlet of the rectifying section, and the outlet of the rectifying section is connected with the inlet of the distribution cavity.

[0009] A design method of a rectifying nozzle type flow divider, comprising the design of a swirl section;

[0010] In the design of the swirl section, the length of the swirl section is obtained by the following formula (1) to (8):

[0011] (1);

[0012] Wherein, H is the pitch of the swirl vane, R is the outer diameter of the swirl tube of the swirl section, and θ is the helix angle of the swirl vane;

[0013] (2);

[0014] Wherein, V is the linear velocity of the rotating airflow, is the mass flow rate of the refrigerant entering the single flow passage of the swirl vane, ρ is the density of the refrigerant, and A is the contact surface area of the single flow passage of the swirl vane and the refrigerant;

[0015] (3);

[0016] Wherein, is the maximum entrained droplet diameter, is the surface tension of the refrigerant, is the gas phase density of the refrigerant, is the gas phase velocity of the refrigerant; and g is the gravitational constant;

[0017] (4);

[0018] Wherein, is the settling velocity, is the entrained droplet diameter, the specific gravity of the particles, the specific gravity of the gas, μ is the dynamic viscosity of the refrigerant gas phase;

[0019] (5);

[0020] wherein, the time required for the center column surface to reach the spiral plate outer edge, the critical particle diameter, the spiral plate outer edge radius, the center column radius; C is the separation constant;

[0021] (6);

[0022] (7);

[0023] wherein N is the number of swirl vane turns;

[0024] (8);

[0025] wherein L is the length of the swirl section.

[0026] Also included is the design of the straightening section; in the straightening section design, the straightening section length is obtained by the following equation (9):

[0027] (9);

[0028] wherein, the Reynolds number, the wing-side fluid velocity, the wing-side fluid kinematic viscosity, the wing-side fluid preferably wing length, the refrigerant flow rate, the refrigerant kinematic viscosity, the straightening section length.

[0029] Also included is the design of the distribution section; in the distribution section design, the size of the throat of the distribution sonic nozzle section is obtained by equations (10) to (11):

[0030] (10);

[0031] wherein, the sound velocity of the gas-liquid two-phase flow mixture, the specific heat ratio of the two-phase flow mixture, the gas constant, the gas phase temperature; the gas mass fraction, the gas volume fraction;

[0032] (11)

[0033] wherein, is the throat area of the velocity-dividing nozzle, is the refrigerant mass flow rate of the branch pipe, is the density of the gas-liquid mixed refrigerant, is the throat radius of the velocity-dividing nozzle.

[0034] Compared with the prior art, the present application has the following advantages:

[0035] 1. The design method of the flow divider provides a complete mathematical model for the design of the flow-dividing nozzle type flow divider, so that the design of the flow divider is more standardized.

[0036] 2. In the flow divider, the center line of the velocity-dividing nozzle and the branch pipe is parallel to the center line of the distribution chamber and parallel to the direction of the incoming flow. By changing the position of the nozzle installed in the distribution chamber, the resistance of the refrigerant entering the velocity-dividing nozzle is reduced, so that the refrigerant can reach the local sonic speed at the nozzle, effectively suppress the upward transmission of pressure wave oscillation, increase the flow rate of the refrigerant, and make the refrigerant flow symmetrically in the flow divider, so that the flow divider can achieve uniform distribution of the refrigerant in the multi-branch evaporator, greatly improve the heat exchange capacity of the evaporator, and thus improve the refrigeration performance of the entire system.

[0037] 3. The flow divider can bear a greater degree of throttling pressure drop, so that the opening degree of the expansion valve does not have to be too small to achieve the same throttling total pressure drop, and under the same throttling total pressure drop, the flow rate of the liquid refrigerant flowing through the evaporator can be increased, and the working range of the expansion valve can be improved. BRIEF DESCRIPTION OF DRAWINGS

[0038] Figure 1 Fig. 1 shows a front view of the flow-dividing nozzle type flow divider of the present application;

[0039] Figure 2 Fig. 2 shows an A-A sectional view of Figure 1

[0040] Figure 3 Fig. 3 shows a B-B sectional view of Figure 1

[0041] Fig. 4 shows a schematic diagram of the principle of the cold air fan performance test experimental system; Figure 4

[0042] Fig. 5 shows a schematic diagram of the refrigeration capacity of the flow divider obtained in Example 1 and the flow divider used as a comparative example, as a function of the library temperature; Figure 5

[0043] Figure 6 ​​The pressure drop of the shunt obtained in Example 1 and the pressure drop of the shunt as a comparative example are shown in the graph of the change of the pressure drop with the reservoir temperature.

[0044] Figure 7 The pressure drop of the shunt obtained in Example 1 and the pressure drop of the shunt as a comparative example are shown in the graph of the change of the pressure drop with the reservoir temperature.

[0045] Figure 8 The pressure drop of the shunt obtained in Example 1 and the pressure drop of the shunt as a comparative example are shown in the graph of the change of the pressure drop with the reservoir temperature.

[0046] In the figure: 1, compressor; 2, oil separator; 3, plate heat exchanger; 4, water pump; 5, liquid accumulator; 6, stop valve; 7, electromagnetic valve; 8, sight glass; 9, electronic expansion valve; 10, electric heater; 11, shunt; 12, air cooler; 13, humidifier; 14, gas-liquid separator; 15, water-cooled condenser; 16, plate heat exchanger; 17, refrigerant mass flow meter; 18, constant-temperature water tank; 19, water tank heater; 20, air-cooled condenser; 21, drying filter; 22, filter; T, temperature measuring point; P, pressure measuring point; F1, cyclone section; F1-1, cyclone pipe; F1-2, spiral plate; F1-3, center column; F2, straightening section; F3, distribution section; F3-1, distribution cavity; F3-2, branch pipe; F3-3, shunt sonic nozzle. DETAILED DESCRIPTION

[0047] The application will be further described in detail below in combination with the accompanying drawings and specific examples.

[0048] The schematic diagram of the straightening nozzle type shunt of the application is shown in Figures 1-3As shown, it comprises a swirl section F1, a straightening section F2 and a distribution section F3 connected in sequence. The swirl section F1 comprises a swirl tube F1-1 and swirl vanes arranged inside the swirl tube, wherein the swirl vanes are composed of a central column F1-3 and a spiral plate F1-2 mounted on the central column, the spiral plate F1-2 is fixedly mounted between the central column F1-3 and the inner wall of the swirl tube F1-1 to make the fluid rotate. The central column F1-3 is connected and fixed with the spiral plate F1-2 structure, and the spiral plate F1-2 is fixed to the inner wall of the swirl tube F1-1 by welding or other means. The swirl section F1 can make the gas-liquid two-phase refrigerant rotate and flow, generate centrifugal force to separate the gas-liquid two-phase, and form a ring flow type which is beneficial to uniform distribution. The straightening section F2 is an arc straightener, which is used to eliminate the centrifugal force of the ring flow formed by the swirl section and process it into stable ring flow. The distribution section F3 comprises a distribution cavity F3-1 and a plurality of branch pipes F3-2, the inlet of each branch pipe F3-2 is located in the distribution cavity, and a splitter sonic nozzle F3-3 is arranged in the inlet of each branch pipe F3-2, respectively. A plurality of branch pipes F3-2 and splitter sonic nozzles F3-3 are uniformly distributed along the circumference with the center of the distribution cavity F3-1 as the center, the center line of each branch pipe F3-2 and splitter sonic nozzle F3-3 is parallel to the center line of the distribution cavity F3-1 and parallel to the flow direction.

[0049] The outlet of the swirl section F1 is connected with the inlet of the straightening section F2, the outlet of the straightening section F2 is connected with the inlet of the distribution cavity F3-1, the inlet of the splitter sonic nozzle F3-3 is located in the distribution cavity F3-1, and the outlet of the splitter sonic nozzle F3-3 is connected with the branch pipe F3-2. The two-phase flow refrigerant after throttling enters the splitter, the centrifugal force is generated in the swirl section F1 by the swirl vanes to make the two-phase flow form a ring flow, then the ring flow passes through the straightening section F2, the centrifugal force of the ring flow is eliminated by the arc straightener to form a stable ring flow, and then enters the distribution cavity F3-1. In the distribution cavity F3-1, the stable ring flow enters each splitter sonic nozzle F3-3, respectively, the two-phase flow is accelerated to local sonic speed by the splitter sonic nozzle F3-3, enters each branch pipe F3-2, and makes each branch pipe F3-2 not affected by the pressure wave oscillation caused by the non-uniform heat exchange in the downstream to the upstream transmission, which influences the distribution capacity of the splitter, so that the refrigerant flows symmetrically and uniformly in the splitter, and the uniformity of the liquid distribution of the splitter is ensured.

[0050] The design method of the straightening nozzle type splitter of the application comprises swirl section design, straightening section design and distribution section design.

[0051] In the swirl section design, the length of the swirl section is obtained by the following formulas (1) to (8):

[0052] (1);

[0053] wherein H is the pitch of the swirl vane, R is the outer diameter of the swirl section swirl tube, determined according to the swirl plate outer edge radius and the swirl tube wall thickness, and θ is the helix angle of the swirl vane;

[0054] (2);

[0055] wherein V is the linear velocity of the rotating gas flow, is the refrigerant mass flow entering the single flow passage of the swirl vane, ρ is the refrigerant density, and A is the contact surface area of the single flow passage of the swirl vane with the refrigerant;

[0056] The formula for calculating the diameter of the liquid droplets that can be entrained by the gas phase is wherein is the diameter of the entrained liquid droplets, is the surface tension of the refrigerant, is the gas phase density of the refrigerant, is the gas phase velocity of the refrigerant; is the Weber number, which is quantitatively explained by the parameter correlation that the Weber number is an important physical quantity determining the maximum entrained liquid droplet size. If the Weber number exceeds a critical value, the liquid droplets will break up. For liquid droplets falling freely under the influence of gravity, it is found that the critical Weber number is between 20 and 30. If a larger estimated value is used, the relationship between the maximum falling diameter of the liquid droplets and the liquid droplet velocity can be obtained, forming the maximum liquid droplet diameter entrained by the gas phase of the annular flow is:

[0057] ;

[0058] Here, the critical Weber number is taken as 30, and thus the calculation formula for the maximum liquid droplet diameter is:

[0059] (3);

[0060] The Stokes settling velocity describes the uniform velocity at which a small spherical particle settles in a viscous fluid under force balance, reflecting the quantitative influence of the particle's own properties (density, size), fluid properties (density, viscosity) on the settling speed.

[0061] In a gravitational field, the basic calculation formula for the Stokes settling velocity is wherein is the density of the particle (unit: kg / m³), is the density of the gas (unit: kg / m³), r is the radius of the particle (unit: m), and the diameter dWith radius r、 Heavy gamma With density rho The replacement can be expressed in formula (4) sedimentation velocity, the value 18 in which is generated in the formula replacement process. The basic formula of Stokes sedimentation velocity is derived based on Stokes law in fluid mechanics and particle force balance analysis, which is based on the force balance principle of spherical particles under laminar flow conditions. The core is to analyze the gravity, buoyancy and viscous resistance of the particle when it is sinking in the viscous fluid, establish the balance equation and solve the velocity;

[0062] (4);

[0063] Wherein, The settling velocity is, The diameter of the entrained droplet is, The particle gravity is, The gravity of the gas is, and μ is the dynamic viscosity of the refrigerant gas phase;

[0064] The time required for the center column surface to the outer edge of the spiral plate is, and its expression is:

[0065] ,

[0066] Substitute In formula (4) into the above formula to obtain formula (5) as follows:

[0067] (5);

[0068] Wherein, The time required for the center column surface to the outer edge of the spiral plate is, The critical particle size is, The outer edge radius of the spiral plate is flexibly selected according to the required cyclone section inlet pipe diameter, The center column radius is flexibly selected according to the actual processing convenience and the requirement of strength, which should not be too large, otherwise it will reduce the flow area, and C is the separation constant, which is obtained by formula (6):

[0069] (6);

[0070] Wherein, V is the linear velocity of the rotating gas flow, and H is the pitch of the cyclone blade;

[0071] (7);

[0072] Wherein, N is the number of cyclone blade turns;

[0073] (8);

[0074] Wherein, L is the length of the cyclone section.

[0075] Formula (8) represents the calculation formula of the length of the cyclone section L , wherein 0.002 m is an empirical value in the design formula, and means that the actual designed length of the cyclone section is more than the length of the spiral blade by 0.002 m, which is a process allowance size for conveniently welding the spiral blade to the wall of the cyclone section, so as to ensure that the welding is firm by fully filling the welding gap with welding material.

[0076] In use, according to the actual inlet liquid supply pipe diameter after throttling, the outer edge radius of the spiral plate, the center column radius, the wall thickness of the cyclone pipe, the wall thickness of the spiral plate, the spiral plate rise angle theta are preferably determined, so as to facilitate welding and control the overall size. Then, according to the working conditions in the refrigeration system, the relevant physical property parameters and the inlet flow parameters of the gas-liquid two-phase refrigerant after throttling are found for design calculation. In sequence, the cyclone pipe outer diameter R , the pitch of the cyclone blade H , the linear velocity of the rotating gas flow, the maximum entrained droplet diameter, the settling velocity, the separation constant, the required time from the center column surface to the outer edge of the spiral plate, the number of cyclone blade turns, and the length of the cyclone section are calculated to obtain, and thus the design of the cyclone section part is completed.

[0077] In the design of the straightening section, the length of the straightening section is obtained by formula (9) as follows:

[0078] (9)

[0079] According to the similarity criterion and the similarity criterion number, the Reynolds number of the straightening section can be equal to the Reynolds number of the wing, wherein, is the Reynolds number, is the fluid velocity on the wing side, is the fluid kinematic viscosity on the wing side, is the preferred wing length on the wing side; is the refrigerant flow rate, is the refrigerant kinematic viscosity, is the length of the straightening section.

[0080] In use, the length of the straightening section is calculated according to the preferred fluid velocity on the wing side, the kinematic viscosity on the wing side, the preferred wing length, the refrigerant flow rate, and the refrigerant kinematic viscosity.

[0081] In the design of the distribution section, the length of the distribution cavity should not be too short or too long, otherwise the uniformity of distribution will be affected, and generally it is less than 0.2 L . The radius of the distribution cavity is determined according to the number of branch pipes. When the number of branch pipes is ≤8, the radius of the distribution cavity is consistent with the radius of the cyclone pipe; when the number of branch pipes is >8, the radius of the distribution cavity is determined according to the minimum radius of the branch pipes evenly distributed on the cavity, and is consistent with the minimum radius.

[0082] The throat size of the flow dividing sonic nozzle section is obtained by equations (10) to (11):

[0083] (10);

[0084] wherein, is the sound speed of the gas-liquid two-phase flow mixture, is the specific heat ratio of the two-phase flow mixture, is the gas constant, and T is the gas phase temperature, is the mass fraction, is the volume fraction.

[0085] (11);

[0086] wherein, S is the throat cross-sectional area of the flow dividing sonic nozzle, is the refrigerant mass flow of the branch pipe, is the density of the gas-liquid mixture refrigerant, is the throat radius of the flow dividing sonic nozzle.

[0087] The sound speed of the gas-liquid two-phase flow mixture is calculated according to the physical property parameters and flow parameters of the refrigerant during use, and the throat radius of the flow dividing sonic nozzle is calculated based on this. Embodiment 1

[0088] Take refrigerant R22 as an example for illustration. The design conditions of embodiment 1 are shown in Table 1:

[0089] Table 1 Design conditions of the flow regulating nozzle type flow divider of embodiment 1:

[0090]

[0091] The physical property parameters and flow parameters of refrigerant R22 under the design conditions are found, and the parameters such as =0.01057 kg / s, =0.016166 N / m, =7.9819 kg / m 3 , =4.25 m / s, =13450.491 (kg*m / s²) / m³, =78.30 (kg*m / s²) / m³ are brought into equations (1) to (8), the outer radius of the spiral plate is selected to be 0.025 m, the spiral angle is selected to be 50°, the thickness of the swirl section pipe wall is selected to be 2 mm, and the outer diameter of the swirl section is calculated based on this R=0.027 m; the thickness of the spiral plate is selected as 2 mm and the radius of the central column is 0.025 m. Based on the above data, the corresponding pitch is calculated H =0.187 m, number of swirl blades N=0.9 turns, and length of swirl section L=0.1703 m.

[0092] According to the similarity criterion and similarity number, the air speed =30 m / s, preferred wing length =0.3 m, kinematic viscosity of air at 20°C =14.8*10 -6 m 2 / s, the speed of refrigerant R22 =2.56 m / s, the kinematic viscosity of refrigerant R22 at -28°C =2.16*10 -7 m 2 Substitute / s into formula (9) to calculate the length of the rectifier section =0.05132 m, and the wall thickness of the rectifying section is selected to be equal to that of the swirl section, which is 2 mm.

[0093] In the design of the distribution section, the length of the distribution cavity is selected as 0.025 m, the wall thickness is selected as 2 mm, the number of branches is 6, and the radius of the distribution cavity is selected as 0.025 m.

[0094] Specific heat ratio of two-phase flow mixture of refrigerant R22 at -28℃ is 1.246, the gas constant is 96.16J / (kg*K), mass fraction is 0.287, volume fraction is 0.986, and the parameters are substituted into the nozzle section calculation formula (10) to obtain the sound velocity of the gas-liquid two-phase flow mixture The refrigerant mass flow rate at each nozzle inlet is 93.13 m / s. =0.00528 kg / s, density of mixed phase =24.77 kg / m 3 , the sound velocity of gas-liquid two-phase flow mixture Substituting into formula (11), the design radius of the split sonic nozzle throat is obtained: is 0.0008 m.

[0095] The above method results in an outer diameter R of the swirl tube F1-1 of 0.027 m, a wall thickness of 0.002 m, and a swirl section length L of 0.1703 m. The outer edge radius of the spiral plate F1-2 is 0.025 m, the number of turns N is 0.9, and the pitch H is 0.187 m. The radius of the central column F1-3 is 0.002 m. The length of the straightening section F2 is 0.05132 m, and the wall thickness is 0.002 m. The length of the distribution chamber F3-1 is 0.025 m, the inner diameter is 0.025 m, and the wall thickness is 0.002 m. There are six branch pipes F3-2 with an outer diameter of 0.009 m. The throat radius of the split sonic nozzle F3-3 is 0.0008 m.

[0096] The rectifier nozzle-type flow splitter designed in Example 1 was installed on a chiller performance test platform to conduct flow splitting performance testing. This platform primarily consists of a constant temperature environmental chamber, a test chamber, a test chamber cooling unit, a constant temperature environmental chamber cooling unit, a cooling water system, and a cooling water system cooling unit. The main parameters of the equipment and instruments involved in the experiment are shown in Tables 2 and 3.

[0097] Table 2 Main equipment and parameters of laboratory cooling device

[0098]

[0099] Table 3 Experimental test equipment and parameters

[0100]

[0101] The air cooler performance test bench is built according to GB / T 25129-2010 "Air coolers for refrigeration". Figure 4 The diagram below shows the schematic diagram of the chiller performance test system. To study the effects of different types of flow dividers on chiller system performance, a Venturi-type flow divider, a rectifying nozzle-type liquid divider described in Patent Publication No. CN104457046A, and the conventional rectifying nozzle-type flow divider obtained in Example 1 were sequentially installed in the system for testing. The flow diversion performance of the flow dividers was analyzed using the chiller's cooling capacity and pressure drop as evaluation indicators.

[0102] During the experiment, the constant temperature environment chamber was kept constant at 0°C, and the test chamber temperatures were set at -24°C, -22°C, -18°C, and -16°C, respectively. The difference between the evaporating temperature and the test chamber temperature was controlled to be 10°C. Experiments were conducted with different flow dividers at these five test chamber temperatures. The experimental operating conditions are shown in Table 4. During the experiment, the evaporating temperature was controlled by adjusting the opening of the electronic expansion valve to maintain superheat within the range of 3°C to 8°C to prevent liquid hammer in the compressor.

[0103] Table 4 Experimental conditions

[0104]

[0105] The schematic diagram of the cooling fan performance test system is as follows: Figure 4 As shown. The test room system mainly consists of three parts: a test room cooling unit, an electric heater and a control system. Among them, the diverter 11 is a Venturi-type diverter, the original rectifying nozzle-type liquid divider involved in the patent publication number CN104457046A, and the current rectifying nozzle-type diverter obtained in this embodiment 1. The low-temperature, low-pressure refrigerant gas is compressed by the compressor to become a high-temperature, high-pressure refrigerant, which is condensed into a liquid phase through a water-cooled condenser and stored in a liquid reservoir. The high-pressure liquid is then adiabatically throttled by an electronic expansion valve, and then diverted into six paths by the diverter, respectively entering each circuit of the air cooler. After heat exchange, it becomes a low-temperature, low-pressure refrigerant gas, and finally separated by a gas-liquid separator and enters the compressor to complete the cycle.

[0106] Experimental results analysis:

[0107] Figure 5 The cooling capacity of the system with different types of flow splitters installed at different test chamber temperatures was calculated. Within the test chamber temperature range of -24°C to -16°C, the cooling capacity increased with increasing test chamber temperature. Under the same operating conditions, the cooling capacity of the system equipped with the existing rectifying nozzle-type flow splitter of Example 1 was higher than that of the system equipped with the original rectifying nozzle-type liquid splitter and Venturi-type flow splitter (patent publication number CN104457046A).

[0108] Comparing the three types of flow dividers, the finned evaporator equipped with the existing rectifying nozzle-type flow divider of Example 1 showed the most significant improvement in cooling capacity compared to the Venturi-type flow divider. The test chamber temperatures were -24°C, -22°C, -20°C, -18°C, and -16°C, respectively, and the evaporation temperatures were -34°C, -32°C, -30°C, -28°C, and -26°C, respectively. The finned evaporator equipped with the existing rectifying nozzle-type flow divider of Example 1 achieved cooling capacity increases of 22.2%, 20.8%, 12.1%, 16.8%, and 21.8%, respectively, compared to the Venturi-type flow divider. The cooling capacity increased by 13.7%, 11.1%, 3.0%, 5.3%, and 13.1%, respectively, compared to the original rectifying nozzle-type liquid divider (patent publication number CN104457046A). The current rectifying nozzle-type flow splitter designed in Example 1 of the present invention can effectively improve the flow splitting performance and thus improve the heat exchange effect of the evaporator, and exhibits better performance at lower evaporation temperatures.

[0109] Figure 6 、 Figure 7 The figures are respectively the diverter pressure drop and the proportion of diverter pressure drop to the total pressure drop after throttling when different types of diverters are installed in the system at different test room temperatures.

[0110] exist Figure 6In the present application, as the temperature of the test chamber rises, the valve opening of the expansion valve becomes larger, the total throttling pressure drop decreases, the evaporation pressure rises, and the evaporation temperature rises accordingly. The larger valve opening of the expansion valve increases the flow of liquid refrigerant in the system, increases the flow rate of the refrigerant in the flow divider, and increases the flow divider pressure drop. The flow divider pressure drop of the present flow divider of the original straightening nozzle type flow divider designed in Embodiment 1 of the present application is much higher than the flow divider pressure drop of the Venturi type flow divider and the original straightening nozzle type flow divider disclosed in Patent No. CN104457046A.

[0111] As shown in Figure 7 , at a test chamber temperature of -24 ℃ to -16 ℃, the proportion of the flow divider pressure drop of the Venturi type flow divider to the total pressure drop after throttling is low, less than 0.2, and has a gradually decreasing trend as the temperature rises; the proportion of the flow divider pressure drop of the straightening nozzle type flow divider to the total pressure drop after throttling is high, more than 0.6. The straightening nozzle type flow divider bears most of the pressure drop after throttling, so that when the evaporation temperature is adjusted by the expansion valve opening, the expansion valve opening does not have to be too small, and more liquid refrigerant can be supplied to the evaporator, and the evaporation area of the evaporator is fully utilized. At a test chamber temperature of -24 ℃ to -16 ℃, the pressure drop after throttling borne by the present straightening nozzle type flow divider designed in Embodiment 1 of the present application is greater than that of the straightening nozzle type flow divider disclosed in Patent No. CN104457046A, indicating that the straightening nozzle type flow divider designed in Embodiment 1 of the present application can bear more pressure drop after throttling.

[0112] Figure 8 reflects the relationship between the proportion of the flow divider pressure drop to the total pressure drop after throttling. In Figure 8 , the proportion of the flow divider pressure drop of the present straightening nozzle type flow divider of Embodiment 1 to the total pressure drop after throttling rises from 13.4% to 24.9% as the temperature of the test chamber rises, which is greater than the rising amplitude of the original straightening nozzle type flow divider disclosed in Patent No. CN104457046A. The proportion of the flow divider pressure drop to the total pressure drop after throttling as the temperature of the test chamber rises indicates that the larger the valve opening of the expansion valve, the smaller the pressure drop of the expansion valve, the greater the flow of liquid refrigerant supplied, and the greater the degree to which the straightening nozzle type flow divider bears the pressure drop after throttling. The comparison between the two straightening nozzle type flow dividers shows that the present straightening nozzle type flow divider of Embodiment 1 can bear a greater degree of pressure drop after throttling under the same expansion valve opening, so that the expansion valve opening does not have to be too small to achieve the same total pressure drop after throttling, increases the flow of liquid refrigerant through the evaporator under the same total pressure drop after throttling, and improves the working range of the expansion valve.

[0113] In summary, the present rectifying nozzle type flow divider designed according to embodiment 1 can effectively reduce the unevenness of refrigerant distribution, effectively improve the flow dividing performance, thereby improving the heat exchange effect of the evaporator and improving the refrigerating capacity. The flow divider pressure drop of the flow divider is much higher than that of other flow dividers, which can better bear the throttling pressure drop of the expansion valve, and the higher the test temperature and the larger the valve opening of the expansion valve, the more obvious the effect of bearing the throttling pressure drop. Under the same total throttling pressure drop, the liquid refrigerant flow is larger than that of the Venturi flow divider and the original rectifying nozzle type flow divider with the patent number CN104457046A. When the refrigeration system using the rectifying nozzle type flow divider designed in the present application is optimized, a smaller electronic expansion valve can be considered, which can save equipment cost and reduce cost.

[0114] The above experimental results show that the flow divider calculated according to the mathematical model of embodiment 1 meets the use requirements and can realize uniform distribution of refrigerant in multiple branch evaporators. In addition, it can also bear the throttling pressure drop of the expansion valve.

[0115] The flow divider of the present application forms annular flow under the action of centrifugal force in the cyclone section, and the flow straightening section converts the fluid into stable annular flow. The refrigerant reaches the critical state at the throat of the flow dividing sonic nozzle. Under the action of the cyclone section and the flow straightening section, the two-phase refrigerant is converted into stable annular flow. The stable annular flow then enters the corresponding branch pipes through each flow dividing sonic nozzle. The two-phase refrigerant reaches the critical state at the throat of each flow dividing sonic nozzle. Each branch pipe supplies refrigerant to each flow path of the evaporator, and the refrigerant is evaporated in the evaporator. The flow divider effectively suppresses the upward transmission of pressure wave oscillation, increases the flow rate of the refrigerant, makes the refrigerant flow symmetrically in the flow divider, and ensures uniform distribution of the flow divider. Thus, the heat exchange capacity of the evaporator is greatly improved, which is beneficial to improving the refrigeration performance of the entire system.

[0116] The above only describes the preferred embodiments of the present application. It should be noted that those skilled in the art can make several improvements and refinements without departing from the principles of the present application, and these improvements and refinements should also be considered within the scope of protection of the present application.

Claims

1. A method of designing a rectifier nozzle-type flow divider, characterized by, The rectifying nozzle type flow divider comprises a swirl section, a rectifying section and a distribution section connected in sequence; the swirl section comprises a swirl tube and swirl vanes arranged inside the swirl tube, the swirl vanes are composed of a center column and a spiral plate mounted on the center column, the spiral plate is fixedly mounted between the center column and the inner wall of the swirl tube to make fluid rotate and flow; the distribution section comprises a distribution cavity and a plurality of branch pipes, the inlet of each branch pipe is located in the distribution cavity, a flow dividing sonic nozzle is arranged in each branch pipe inlet respectively, a plurality of branch pipes and the flow dividing sonic nozzles thereon are distributed along the circumference with the center of the distribution cavity as the center, the center line of each branch pipe and the flow dividing sonic nozzle thereon is parallel to the center line of the distribution cavity and parallel to the direction of the incoming flow; the outlet of the swirl section is connected with the inlet of the rectifying section, the outlet of the rectifying section is connected with the inlet of the distribution cavity; In the design of the swirl section, the length of the swirl section is obtained by the following formulas (1) to (8): （1）； in, H is the pitch of the swirl blade, R is the outer diameter of the swirl tube in the swirl section, θ is the helix angle of the swirl blade; （2）； wherein V is the linear velocity of the rotating airflow, is the refrigerant mass flow rate into the single flow passage of the swirl vane, is the refrigerant density, A is the contact surface area of the single flow passage of the swirl vane with the refrigerant; （3）； wherein, Dmax is the maximum entrained droplet diameter, σ is the refrigerant surface tension, pg is the refrigerant gas phase density, vg is the refrigerant gas phase velocity, g g is the gravitational constant; （4）； wherein, is the settling velocity, is the entrained droplet diameter, is the particle density, is the gas density, is the refrigerant gas phase dynamic viscosity; （5）； wherein, t is the time required for the surface of the central column to reach the outer edge of the spiral plate; d is the critical particle diameter; r is the radius of the outer edge of the spiral plate; R is the radius of the central column; C K is the separation constant; g G is the gravitational constant; （6）； （7）； wherein N is the number of the swirl vanes; （8） wherein L is the length of the cyclone section.

2. The method of designing a rectifier nozzle-type flow divider according to claim 1, characterized in that, The design of the rectifying section is also included; in the design of the rectifying section, the length of the rectifying section is obtained by the following formula (9): （9）； wherein, Re Reynolds number, Vw is the wing-side fluid velocity; Vw is the wing-side fluid kinematic viscosity; Lw is the wing-side fluid, preferably wing, length; Vr is the refrigerant flow velocity; Vr is the refrigerant kinematic viscosity; Lr is the rectifier length.

3. The design method of a rectifier nozzle type flow divider according to claim 1 or 2, characterized in that, The design of the distribution section is also included; in the design of the distribution section, the throat size of the flow dividing sonic nozzle section is obtained by the following formulas (10) to (11): （10）； wherein, is the sound speed of the gas-liquid two-phase flow mixture, is the specific heat ratio of the two-phase flow mixture, is the gas constant, is the gas phase temperature, is the gas mass fraction, is the gas volume fraction; （11）； wherein, S is the throat area of the diverging sonic nozzle, is the refrigerant mass flow rate of the branch pipe, is the density of the gas-liquid mixed refrigerant, is the throat radius of the diverging sonic nozzle.

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