An optimization control method for refrigerant dynamic flow distribution

Abstract

The application discloses a kind of optimization control methods of refrigerant dynamic flow distribution, realize the dynamic coupling of air side and refrigerant side, so as to be able to according to wind speed high and low distribution match different refrigerant flow size, improve the utilization of evaporator effective heat exchange area, improve system energy efficiency.It includes the following steps: establishing wind speed-refrigerant flow structure model;Establishing multi-flow path fin heat exchanger distribution parameter model:Along the refrigerant flow direction, establish two-dimensional microelement model, that is: each refrigerant flow path is regarded as equal length flow path and extends along two-dimensional direction, and the enthalpy value of heat exchange tube is divided to calculate microelement;According to refrigerant flow order, microelement model of single refrigerant flow path is solved in sequence, according to the solution result, wind speed module and refrigerant flow are matched.The method solves the contradiction between overheating degree and liquid strike risk in system operation by flow dynamic regulation, realizes the efficient and stable operation of system under variable working condition, and improves system energy efficiency.

Description

TECHNICAL FIELD

[0001] The present application relates to the field of regulation technology, and more particularly, to an optimization control method for refrigerant dynamic flow distribution. BACKGROUND

[0002] In an air source heat pump system, the evaporator as the core heat exchange component, its performance directly determines the system energy efficiency, however, due to the influence of axial flow fan layout and fin structure, the wind speed on the windward surface of the evaporator is significantly non-uniformly distributed, and the traditional flow distributor uniformly distributes the refrigerant, which will cause incomplete evaporation of the refrigerant in the high wind speed area, and the heat transfer coefficient decreases, and in the low wind speed area, the excessive charging causes the risk of liquid knock and heat exchange dead zone, resulting in insufficient utilization of the effective heat exchange area of the evaporator, and the system COP and heating capacity decrease. The patent document with publication number CN104457046A and invention name "Rectifying nozzle type equal flow distributor and refrigeration system" discloses a distributor, the two-phase refrigerant flow pattern can be actively regulated to a stable annular flow, which is beneficial to uniformly supplying liquid to each flow path of the evaporator, and the supersonic nozzle achieves a critical state through the throat to suppress the influence of downstream pressure wave oscillation on uniform flow distribution, and the supersonic nozzle adopts a unified fixed aperture design to ensure uniform liquid supply to each flow path of the evaporator. However, when applied to an actual air source heat pump system, the design of a unified fixed aperture has the defect that it cannot match the flow to supply liquid according to different wind speed evaporation areas, which hinders the full utilization of the effective heat exchange area of the evaporator. The patent document with publication number CN117847739A and invention name "Dual-evaporation-temperature air conditioner and air conditioner inlet air control method" discloses an optimization scheme for a dual-evaporation-temperature air conditioner, which divides the indoor heat exchanger into front and rear rows of heat exchange flow paths, each row is composed of two independent heat exchange units in parallel, detects the refrigerant temperature of each branch through a temperature sensing bag, and adjusts the inlet air angle of the inlet air component according to the temperature difference to improve the outlet air temperature difference and the uniformity of the refrigerant outlet temperature. However, when applied to an actual air source heat pump system or a multi-flow path fin heat exchanger, this scheme has obvious limitations: first, it only passively adjusts the air inlet angle, does not touch the core of refrigerant flow distribution, cannot adapt to the non-uniform distribution of wind speed on the windward surface of the evaporator, and cannot solve the problem of insufficient refrigerant supply in the high wind speed area and excessive refrigerant supply in the low wind speed area. It does not establish a "wind speed-flow-heat transfer" coupling mechanism, which cannot improve the utilization rate of the effective heat exchange area of the heat exchanger, and also cannot optimize the core performance such as system COP and heating capacity; second, the adjustment of the inlet air angle is restricted by the hardware structure, and the adaptability is poor, which cannot cope with the wind speed fluctuation under different working conditions, and it is difficult to realize efficient heat exchange under all working conditions, which hinders the full play of the overall performance of the heat exchanger.

[0003] At present, the research on the performance of the evaporator is isolated optimization of the air side or the refrigerant side, and lacks systematic research on the dynamic coupling mechanism of "air volume-heat transfer-flow rate", and it is difficult to realize multi-factor collaborative optimization. Therefore, the dynamic coupling of the air side and the refrigerant side must be realized, so as to match different refrigerant flow rates according to different air velocities, so as to realize the full use of the effective heat exchange area of the evaporator and improve the performance of the evaporator. SUMMARY

[0004] The purpose of the present application is to overcome the shortcomings of the prior art, provide an optimization control method for dynamic flow distribution of refrigerant, realize the dynamic coupling of the air side and the refrigerant side, so as to match different refrigerant flow rates according to the distribution of air velocity, improve the performance of the evaporator and improve the system energy efficiency.

[0005] The technical scheme adopted to achieve the purpose of the present application is:

[0006] An optimization control method for dynamic flow distribution of refrigerant, comprising the following steps:

[0007] Step 1: Establish a wind speed-refrigerant flow structure model: in the air side, the fin plane is divided into Y wind speed modules according to the different air supply amounts, and each wind speed module corresponds to Z refrigerant flow paths;

[0008] Step 2: Establish a multi-flow path fin heat exchanger distribution parameter model: a two-dimensional microelement model is established along the refrigerant flow direction;

[0009] Step 3: Solve the microelement model of a single refrigerant flow path in sequence according to the refrigerant flow order, and realize the matching of the wind speed module and the refrigerant flow rate according to the solving result; comprising the following steps:

[0010] Step 3-1: respectively establish a single flow path air side heat transfer model, a single flow path refrigerant side heat transfer model and a single flow path refrigerant side pressure drop model;

[0011] Step 3-2: based on the known single flow path refrigerant and air inlet parameters, set the initial mass flow rate m0 of the single flow path refrigerant according to the distribution mode of the wind speed module, assume the inlet and outlet enthalpy values and pressure values of the refrigerant according to the test working condition, and solve the microelement model of the single refrigerant flow path;

[0012] Step 3-3: calculate the single microelement length, and accumulate to obtain the total length of the single heat exchange pipe; and compare the calculated total length of the single heat exchange pipe with the actual total length of the single flow path;

[0013] Step 3-4: iterative calculation is carried out with the convergence of the single heat exchange pipe total length calculation value to the actual length as the criterion, and the outlet enthalpy value is adjusted:

[0014] Step 3-5: The relative error between the calculated value and the actual length of the single heat exchange tube is not more than 1%, and the initial mass flow rate of the single flow path refrigerant set in this state is taken as the actual single flow path refrigerant mass flow rate, and the refrigerant outlet parameters of the single flow path are output;

[0015] Step 3-6: The refrigerant mass flow rates of each flow path under the same air speed module are added to obtain the total mass flow rate of the refrigerant under the air speed module.

[0016] Step 3-7: The corresponding refrigerant flow rates of the Y air speed modules are added to obtain the calculated total mass flow rate of the heat exchanger.

[0017] Step 3-8: According to the predicted heat exchange amount and the calculated estimated tube length, the refrigerant side pressure drop model established in step 3-1 is used to calculate the pressure drop; according to the outlet calculated pressure drop and temperature, the outlet enthalpy value is obtained, and according to the outlet refrigerant temperature and pressure of the heat exchanger, the saturation temperature is obtained, and the superheat is calculated.

[0018] Step 3-9: Iterative adjustment is performed according to the criterion that the calculated total mass flow rate of the heat exchanger converges to the true value.

[0019] Step 3-10: The relative error between the calculated total mass flow rate of the heat exchanger and the actual value is not more than 1%, and the corresponding refrigerant total mass flow rate of each air speed module is determined to determine the distribution ratio of the air speed and the refrigerant mass flow rate of each air speed module.

[0020] In step 2, a two-dimensional microelement model is established along the refrigerant flow direction: each refrigerant flow path is regarded as an equal length flow path and is extended along the two-dimensional direction, and the microelements are calculated according to the enthalpy value of the heat exchange tube.

[0021] Step 3-2 includes the following steps: the inlet parameters of the first microelement are directly determined by the design parameters of the test working condition; based on the mass conservation principle, the energy conservation principle, the momentum conservation principle and the heat transfer principle, the air side heat transfer amount is calculated by using the single flow path air side heat transfer model established in step 3-1, the refrigerant side heat transfer amount is calculated by using the single flow path refrigerant side heat transfer model, the refrigerant side pressure loss is calculated by using the single flow path refrigerant side pressure drop model, and the refrigerant state parameters at the end of the first microelement are obtained through the flow enthalpy difference relationship that the air side heat transfer amount is equal to the refrigerant side heat transfer amount. The refrigerant state parameters at the end of the first microelement are taken as the refrigerant state parameters at the beginning of the next microelement for iterative calculation, and the refrigerant state parameters at the beginning and end of each microelement are obtained, which provides boundary conditions for microelement length calculation.

[0022] Step 3-3, the method for calculating the length of a single microelement is: according to the set single flow path refrigerant initial mass flow rate m0 and the refrigerant state at the beginning and end of each microelement, the predicted heat exchange amount of the refrigerant and the air side wind speed module in the microelement section is calculated by using the single flow path air side heat transfer model and the single flow path refrigerant side heat transfer model established in step 3-1 according to the mass conservation principle, the energy conservation principle, the momentum conservation principle and the heat transfer principle, and the maximum heat exchange amount is taken as the optimization objective, and the flow distribution ratio is taken as the optimization variable, so that the length of a single microelement is calculated along the path by using the difference equation set.

[0023] Step 3-4, adjusting the outlet enthalpy value includes: if the calculated value of the total length of a single heat exchange pipe is greater than the actual length, it indicates that the set value of the refrigerant outlet enthalpy is too small, and the set value needs to be increased; on the contrary, if the calculated value of the total length of a single heat exchange pipe is less than the actual length, it indicates that the set value of the refrigerant outlet enthalpy is too large, and the set value needs to be reduced; then, the calculation along the path is repeated until the relative error between the calculated value of the total length of a single heat exchange pipe and the actual length is not more than 1%.

[0024] In step 3-8, if the superheat degree is greater than 8 DEG C, the single flow path refrigerant initial mass flow rate m0 is increased, if the superheat degree is less than 3 DEG C, the single flow path refrigerant initial mass flow rate m0 is reduced, and step 3-1 is repeated until the superheat degree is 3 DEG C-8 DEG C.

[0025] The iterative adjustment of step 3-9 is: if the calculated value of the total mass flow rate of the heat exchanger is greater than the actual value, the single flow path refrigerant initial mass flow rate m0 under each wind speed module is reduced by the same proportion; if the calculated value of the total mass flow rate of the heat exchanger is less than the actual value, the single flow path refrigerant initial mass flow rate m0 under each wind speed module is increased by the same proportion, and the calculation is repeated until the relative error between the calculated value of the total mass flow rate of the heat exchanger and the actual value is not more than 1%.

[0026] According to the distribution ratio of the wind speed module and the refrigerant flow rate, the refrigerant flow rate of each nozzle of the flow divider is determined.

[0027] According to the refrigerant flow rate of each nozzle, the throat size of the distribution section is determined.

[0028] In step 3-4, the outlet enthalpy value is adjusted by using the dichotomy method, and the upper limit is the refrigerant inlet enthalpy, and the lower limit is the supercooled liquid enthalpy at the air inlet temperature.

[0029] The present application has the following technical progress:

[0030] 1. Multi-factor collaborative optimization: the method of the present application realizes efficient heat exchange in the whole region of the evaporator by establishing the correlation model among the air side wind speed, the refrigerant flow rate and the heat transfer coefficient, solves the problem of "under-supply in high wind speed area and over-supply in low wind speed area" caused by isolated optimization in the traditional technology, and is beneficial to improving the utilization rate of the effective heat exchange area of the evaporator.

[0031] 2. Dynamic flow regulation mechanism: The method of the present invention monitors the wind speed distribution on the windward side of the evaporator in real time and dynamically adjusts it based on the "air volume-heat transfer-flow" coupling model, so that more refrigerant is obtained in the high wind speed area and the flow is reduced in the low wind speed area, thereby improving the overall heat exchange efficiency of the heat exchanger. Attached Figure Description

[0032] Figure 1 The diagram shown is a schematic of the wind speed-refrigerant flow structure model with an evaporator as an example.

[0033] Figure 2 The diagram shows the layout of the wind speed measuring points.

[0034] Figure 3 The image shows a wind speed distribution map;

[0035] Figure 4 The diagram shown is a schematic of a rectifier nozzle type distributor.

[0036] Figure 5 The diagram shown is a schematic of the nozzle structure of a rectifier nozzle type distributor.

[0037] Figure 6 The diagram shown is a schematic of the nozzle diameter of a rectifier nozzle type distributor.

[0038] Figure 7 The diagram shown illustrates the correspondence between the nozzle and the wind speed module in a rectifier nozzle-type distributor.

[0039] Figures 8-10 The figure shown is a comparison of the COP of different shunts.

[0040] Figures 11-13 The image shows a comparison of the heating capacity of different splitters. Detailed Implementation

[0041] This invention addresses the problems of low heat exchange efficiency, liquid slugging risk, and reduced system energy efficiency in air-source heat pump systems caused by non-uniform wind speed distribution on the evaporator's front side. It proposes an evaporator refrigerant flow distribution scheme based on a dynamic coupling mechanism of "air volume-heat transfer-flow rate." Specifically, it includes:

[0042] Multi-factor collaborative monitoring: An anemometer is placed on the windward side of the evaporator to collect wind speed distribution data in each area, and the refrigerant temperature and pressure at the inlet and outlet of the evaporator and the compressor operating parameters are monitored simultaneously;

[0043] Dynamic allocation control model: Establish the coupling relationship between wind speed distribution and refrigerant phase change heat transfer, and form an allocation strategy of "matching high flow rate in high wind speed area and low flow rate in low wind speed area". Use the flow allocation ratio as the optimization variable, the optimization objective is the maximum heat exchange, and dynamically calculate the refrigerant flow demand of each flow path.

[0044] The optimization control method for refrigerant dynamic flow distribution of the application comprises the following steps:

[0045] Step 1: Establish a wind speed-refrigerant flow structure model: On the refrigerant side, the multi-flow finned heat exchanger has X refrigerant flow paths from top to bottom, and on the air side, the fin plane is divided into Y wind speed modules from top to bottom according to different air supply amounts, and each wind speed module corresponds to Z refrigerant flow paths. X is determined according to the actual heat exchanger; Y is divided into Y wind speed modules according to the refrigerant flow path perpendicular to the fan coil air outlet direction and the wind speed gradient, and is determined according to the actual situation, and Y is at least 2; Z is the refrigerant flow path corresponding to each wind speed module, and is determined according to the actual situation, and Z is at least 1.

[0046] Step 2: Establish a multi-flow finned heat exchanger distribution parameter model: a two-dimensional microelement model is established along the refrigerant flow direction, that is, each flow path is regarded as an equal-length flow path and is extended along the two-dimensional direction, and the microelement is calculated according to the enthalpy value of the heat exchange pipe.

[0047] In a common heat exchanger, the mathematical model of the air side and the refrigerant side usually adopts a counter-flow heat transfer model, and the application adopts a conventional counter-flow type evaporator steady-state distribution parameter model. According to the model, the microelement is divided according to the enthalpy value of the heat exchange medium. By subdividing the flow path into a plurality of calculation microelements, the state change of the refrigerant in each microelement can be more accurately calculated.

[0048] Step 3: The microelement model of a single flow path is solved in sequence according to the refrigerant flow order, and the matching of the wind speed module and the refrigerant flow is realized. It comprises the following steps:

[0049] The solution of the application is based on the following assumptions:

[0050] 1) The heat transfer process in the evaporator is one-dimensional steady-state flow and in counter-flow form;

[0051] 2) The axial heat conduction and the elbow heat exchange of the heat exchange pipe are ignored;

[0052] 3) The refrigerant flow in the pipe is one-dimensional flow along the pipe axis;

[0053] 4) The air flow direction is parallel to the fin;

[0054] 5) The axial heat conduction is ignored.

[0055] Step 3-1: The air side heat transfer model of a single flow path, the refrigerant side heat transfer model of a single flow path and the refrigerant side pressure drop model of a single flow path are respectively established by using a conventional method.

[0056] Step 3-2: Based on the known single flow path refrigerant and air side inlet parameters, the initial mass flow rate m0 of the single flow path refrigerant is set according to the distribution mode of the wind speed module. The inlet and outlet states of the refrigerant under the test working condition are assumed, including the enthalpy and pressure values of the refrigerant side inlet and outlet, to solve the micro-element model of the single refrigerant flow path. Among them, the distribution mode of the wind speed module is: the greater the wind speed of the wind speed module, the greater the initial mass flow rate m0 of the single flow path refrigerant set.

[0057] The inlet parameters of the first micro-element refrigerant and air are directly determined by the design parameters of the test working condition, including the inlet temperature, inlet pressure of the refrigerant side, and the inlet temperature, inlet pressure of the air side. Based on the mass conservation principle, energy conservation principle, momentum conservation principle and heat transfer principle, the air side heat transfer amount is calculated by using the single flow path air side heat transfer model established in step 3-1, the refrigerant side heat transfer amount is calculated by using the single flow path refrigerant side heat transfer model, the refrigerant side pressure loss is calculated by using the single flow path refrigerant side pressure drop model, and the terminal refrigerant state parameters of the first micro-element are obtained through the flow enthalpy difference relationship that the air side and the refrigerant side heat transfer amount is equal. The terminal refrigerant state parameters of the first micro-element are taken as the initial refrigerant state parameters of the next micro-element for iterative calculation, and so on. The initial and terminal refrigerant state parameters of each micro-element, including temperature, pressure, dryness and enthalpy, are obtained, which provide boundary conditions for micro-element length calculation.

[0058] Step 3-3: According to the set initial mass flow rate m0 of the single flow path refrigerant and the initial and terminal refrigerant state parameters of each micro-element, according to the mass conservation principle, energy conservation principle, momentum conservation principle and heat transfer principle, the predicted heat exchange amount of the refrigerant and the air side wind speed module in the micro-element section is calculated by using the single flow path air side heat transfer model and the single flow path refrigerant side heat transfer model established in step 3-1. Taking the maximum heat exchange amount as the optimization objective and the flow distribution ratio as the optimization variable, the single micro-element length is calculated along the path through the difference equation set, and the total length of the single heat exchange pipe is obtained by accumulating the length of each micro-element. Then, the calculated total length of the single heat exchange pipe is compared with the actual total length of the single flow path.

[0059] Step 3-4: The outlet enthalpy setting value is iteratively adjusted with the convergence of the single heat exchange pipe total length calculation value to the actual length as the criterion. If the single heat exchange pipe total length calculation value is greater than the actual length, it indicates that the refrigerant outlet enthalpy setting value is too small and needs to be increased. Conversely, if the single heat exchange pipe total length calculation value is less than the actual length, the outlet enthalpy setting value is too large and needs to be reduced. Then, the calculation along the path is repeated until the relative error between the single heat exchange pipe total length calculation value and the actual length is not more than 1%. The calculated single heat exchange pipe total length calculation value is taken as the estimated pipe length.

[0060] The application adjusts the outlet enthalpy value by using the conventional dichotomy, and the upper limit is the refrigerant inlet enthalpy, and the lower limit is the supercooled liquid enthalpy at the air inlet temperature.

[0061] Step 3-5: The single flow path refrigerant mass flow rate set in the state that the relative error of the total length of the single heat exchange pipe calculated value and the actual length is not more than 1% is taken as the actual single flow path refrigerant mass flow rate, and the refrigerant outlet parameters of the single flow path are output, including temperature, pressure and enthalpy value.

[0062] Step 3-6: The refrigerant mass flow rates and refrigerant outlet parameters of the flow paths belonging to the same air speed module are the same, the refrigerant mass flow rates of the flow paths belonging to the same air speed module are added to obtain the total refrigerant mass flow rate under the air speed module. Each air speed module is solved to obtain the total refrigerant mass flow rate corresponding to each of the Y air speed modules.

[0063] Step 3-7: The total refrigerant mass flow rates corresponding to the Y air speed modules are added to obtain the total mass flow rate calculation value of the evaporator.

[0064] Step 3-8: The predicted heat exchange amount obtained in step 3-3 and the estimated pipe length calculated in step 3-4 are used to perform pressure drop calculation by using the refrigerant side pressure drop model established in step 3-1, the enthalpy value is obtained according to the outlet calculated pressure drop and temperature, the superheat is calculated by calling Refprop according to the temperature and pressure of the evaporator outlet refrigerant, if the superheat is greater than 8℃, the single flow path refrigerant initial mass flow rate m0 is increased, if the superheat is less than 3℃, the single flow path refrigerant initial mass flow rate m0 is reduced; repeat step 3-2 until the superheat is 3℃-8℃.

[0065] Step 3-9: Iterative adjustment is performed according to the criterion that the total mass flow rate calculation value of the evaporator converges to the actual value, if the total mass flow rate calculation value of the evaporator is greater than the actual value, the single flow path refrigerant initial mass flow rate m0 under each air speed module is reduced by the same proportion, if the total mass flow rate calculation value of the evaporator is less than the actual value, the single flow path refrigerant initial mass flow rate m0 under each air speed module is increased by the same proportion, and the calculation is repeated until convergence, so that the relative error of the total mass flow rate calculation value of the evaporator and the actual value is not more than 1%.

[0066] Step 3-10: The refrigerant total mass flow rates corresponding to each air speed module in the state that the relative error of the total mass flow rate calculation value of the evaporator and the actual value is not more than 1% are used to determine the distribution proportion of the air speed and the refrigerant refrigeration flow rate of each air speed module.

[0067] Step 4: Determine the refrigerant flow rate of each nozzle of the flow divider according to the distribution ratio of each wind speed module and the refrigerant flow rate. Determine the throat size of the distribution section according to the refrigerant flow rate of each nozzle.

[0068] The design of the throat size of the distribution section of the straightening nozzle type flow divider is based on the Wood sound speed model to establish the critical flow condition. The sound speed condition is assumed to be isothermal and phase equilibrium with equal velocity of two-phase flow, and the sound speed formula is simplified as:

[0069] (1);

[0070] wherein, is the sound speed of the gas-liquid two-phase flow mixture, unit: m / s; is the specific heat ratio of the two-phase flow mixture, is the gas constant, unit: J / (kg·K); is the gas phase temperature, unit: K; is the mass fraction, is the volume fraction.

[0071] (2);

[0072] wherein, is the throat cross-sectional area of the sonic nozzle, is the refrigerant mass flow rate of a single circuit, is the density of the gas-liquid mixed refrigerant, is the throat radius of the sonic nozzle.

[0073] The following describes the embodiment by taking the traditional straightening nozzle type flow divider as an equal flow distribution flow divider, and taking the variable flow straightening nozzle type flow divider as shown in Figure 4 as a dynamic flow distribution flow divider.

[0074] In this embodiment, the wind speed-refrigerant flow structure model is as shown in Figure 1As shown, including multi-flow finned heat exchanger 2 and fan 1, multi-flow finned heat exchanger 2 geometric structure includes fin row spacing, fin spacing, fin thickness, tube row number, tube column number, tube row length, single flow tube row length, fin height, fin width, tube wall thickness and tube outer diameter, etc. In this embodiment, the multi-flow finned heat exchanger 2 has a total of 29 refrigerant flow paths from top to bottom, and on the air side, the fin plane is divided into 4 wind speed modules from top to bottom, which are low wind speed area 11-1, first medium wind speed area 11-2, second medium wind speed area 11-3 and high wind speed area 11-4. Different wind speed modules correspond to different air volume; the heat exchange tube and the fin are arranged vertically, and the refrigerant and the air are respectively in the inner and outer walls of the tube to generate convective heat exchange and the tube wall and the fin to conduct heat. Among them, the low wind speed area 11-1 corresponds to 8 refrigerant flow paths, the first medium wind speed area 11-2, the second medium wind speed area 11-3 and the high wind speed area correspond to 7 refrigerant flow paths respectively.

[0075] Affected by the wind pressure of the axial flow fan and the coil placement position, the wind speed on the windward surface of the U-shaped fin is not uniformly distributed. The closer to the air inlet position, the higher the wind speed; the farther from the air inlet, the suction rate of the fan decreases, and the air flow through the farthest end coil is very limited. In order to test the relationship between wind speed and fin height, an anemometer is arranged on the front surface of the evaporator to collect wind speed distribution data in each area. In this paper, a thermosensitive anemometer is used to test the wind speed on the fin surface, a total of 112 test points are designed, and the axial flow fan is set to an air volume of 23500m3 / h, Figure 2 As shown in the figure, since the side of the heat exchanger is a symmetrical structure, the wind speed measuring point O is arranged on the P surface and half of the Q surface to reduce the calculation amount.

[0076] The above measuring points are longitudinally averaged into 4 parts, and the wind speed of each part is determined as the average value of 2 rows of wind speed. The wind speed distribution of each test interval is as Figure 3 shown. Under the condition of the same windward area, the greater the wind speed, the more air passes through per unit time, and the stronger the heat exchange capacity. After calculation, the wind speed of the 4 parts is as shown in Table 1.

[0077]

[0078] In this embodiment, the 29 refrigerant flow paths are considered as equal length flow paths and are extended in two-dimensional direction, and the microelement is calculated according to the enthalpy value of the heat exchange tube.

[0079] Based on the known inlet state of refrigerant and air, as shown in Table 2, the outlet state parameters of each microelement are obtained.

[0080]

[0081] The relative error between the total mass flow rate of the evaporator calculated by the method of steps 1-3 and the actual value is not more than 1% at each wind speed module corresponding to the refrigerant total mass flow rate. Since the maximum heat exchange is the target, the refrigerant mass flow rate of the four wind speed modules is obtained when the heat exchange is maximum.

[0082] It is calculated that when the heat exchange is maximum, the mass flow rate of the four wind speed zones from high wind speed zone to low wind speed zone is distributed according to the ratio of 6:5:5:4.

[0083] According to the above ratio, the flow divider is designed, and the key design parameters, design conditions and various conditions are determined according to the actual application requirements and working conditions, as shown in Table 3. Considering various factors in actual operation, the refrigerating capacity is corrected by 15%, and the actual calculated refrigerating capacity after correction is 38.827kW.

[0084]

[0085] The overall structure of the variable flow rectifying nozzle flow divider is shown in Figure 4 , which includes a cyclone section L3, a rectifying section L2 and a distribution section L1. The nozzle structure is shown in Figure 5 . In addition to achieving critical flow distribution, the distribution section supersonic nozzle also needs to achieve on-demand distribution function. Figure 5 , where dcr is the throat diameter.

[0086] The total mass flow rate of the system is known to be 0.2576kg / s, and according to the calculation, the mass flow rate of the four wind speed zones from high wind speed zone to low wind speed zone is distributed according to the ratio of 6:5:5:4. In designing the nozzle section, combined with the 6:5:5:4 flow distribution ratio of the 29 branch pipes, it is determined that there are 7 large nozzles, each with a flow rate of 0.0110kg / s, 14 medium nozzles, each with a flow rate of 0.0092kg / s, and 7 small nozzles, each with a flow rate of 0.0064kg / s, to ensure the uniformity of the flow of each circuit of the evaporator and the stability of the gas-liquid volume fraction at 98.42% and 1.58% respectively, and to realize the efficient maintenance of annular flow. The nozzle section adopts the critical flow distribution theory, and the stable distribution of gas-liquid two-phase flow is realized through the Laval nozzle structure. Based on the Wood sound velocity formula, assuming that the gas-liquid two-phase does not change and the flow rate is equal, the mixed sound velocity is calculated to be 136.88m / s by using the above formula (1). The throat cross-sectional area is determined by using the above formula (2).

[0087] The system has a total of 29 branches, and 4 wind speed zones are arranged from high wind speed zone to low wind speed zone according to a flow distribution scheme of 6:5:5:4. After calculation, the throat area of the large branch is 2.38 mm2 (throat diameter 1.742 mm), the throat area of the medium branch is 1.984 mm2 (throat diameter 1.590 mm), and the throat area of the small branch is 1.389 mm2 (throat diameter 1.330 mm). The design of the flow divider in the present application is affected by the machining precision, and the final parameters of the throat diameter are 1.8 mm for the large nozzle, 1.55 mm for the medium nozzle, and 1.35 mm for the small nozzle. The nozzles of different types are arranged in this order.

[0088] When installing the flow divider, the large, medium and small nozzles are connected to the high, medium and low speed wind zone flow paths in this order, and the installation schematic diagram is shown in Figure 6 、 Figure 7 . The large nozzle 10-1 has a total of 7 nozzles, which are connected to the 7 flow paths of the high speed wind zone 11-4; the medium nozzle 10-2 has a total of 14 nozzles, which are connected to the 14 flow paths of the first medium speed wind zone 11-2 and the second medium speed wind zone 11-3; the small nozzle 10-3 has a total of 8 nozzles, which are connected to the 8 flow paths of the low speed wind zone 11-1; a total of 29 nozzle outlets and 29 finned tube heat exchanger flow paths.

[0089] For performance analysis of the flow divider applied to air source heat pump, it is necessary to compare the flow dividing performance of the equal flow rectifying nozzle flow divider and the variable flow rectifying nozzle flow divider under the same experimental conditions, and the influence of the two kinds of flow dividers on the performance of the air source heat pump system. According to the use side end form: floor radiation type, fan coil type and radiator type, three different outlet water temperatures are set. At the same time, five groups of required test environmental temperature conditions are set, which are 7°C, 2°C, -7°C, -12°C and -20°C. While controlling different environmental temperatures, the evaporation pressure is also controlled, so that the same evaporation pressure and evaporation temperature can be obtained when the different flow dividers are measured, so that the final data comparison is more analyzable. After controlling the same evaporation temperature and evaporation pressure, the experiments of the two kinds of flow dividers under the same conditions are calculated, and the comparison of the two kinds of flow dividers is carried out. The specific experimental test conditions are shown in Table 4.

[0090]

[0091] Figures 8-10The system COP of the system installed with the equal-flow rectifying nozzle distributor and the variable-flow rectifying nozzle distributor under different environmental temperature is compared. In the environmental temperature range of -20 DEG C to 7 DEG C, the system COP decreases with the decrease of the environmental temperature, which accords with the positive correlation characteristic of the evaporation temperature and the energy efficiency of the heat pump system. When the variable-flow distribution form is adopted, the COP of the floor radiation type, the fan coil type and the radiator type work conditions is respectively increased by 5.8%, 4.1% and 5.4% at the design working condition (-12 DEG C), which proves the working condition adaptability design advantage of the variable-flow distributor.

[0092] The COP and the heating capacity of the system adopting the variable-flow rectifying nozzle distributor are higher than those of the system installed with the equal-flow uniform distribution form under the same working condition. When the environmental temperature is -20 DEG C, -12 DEG C, -7 DEG C, 2 DEG C and 7 DEG C respectively and the outlet water temperature is the floor radiation type, the system energy efficiency of the variable-flow distribution form is respectively increased by 3.2%, 5.8%, 1.2%, 1.8% and 2.4% compared with the equal-flow distribution form. When the outlet water temperature is the fan coil type, the system energy efficiency is respectively increased by 3.5%, 4.1%, 0.9%, 0.7% and 2.2%. When the outlet water temperature is the radiator type, the system energy efficiency is respectively increased by 4.5%, 5.4%, 1.0%, 4.5% and 0.7%. The data can more directly show that the variable-flow distribution form is generally improved compared with the equal-flow distribution form, and the overall average is increased by 2.8%.

[0093] The method of the application realizes the increase of the refrigerant flow in the high air speed area, the decrease of the liquid holdup in the low air speed area, achieves the reasonable adaptation of "high air volume - high flow" and "low air volume - low flow", significantly enhances the local heat exchange efficiency and improves the system energy efficiency ratio.

[0094] Figures 11-13 The heating capacity of the system adopting different refrigerant distribution forms under different environmental temperatures is compared. In the environmental temperature range of -20 DEG C to 7 DEG C, the air source heat pump heating capacity decreases with the decrease of the environmental temperature, and the overall negative correlation is shown. From the thermodynamic mechanism, the decrease of the environmental temperature causes the corresponding decrease of the system evaporation temperature, and the decrease of the evaporation temperature as the key thermodynamic parameter directly affecting the heating capacity will change the energy transfer efficiency of the heat pump cycle, and then cause the regularity of the overall system heating capacity.

[0095] Compared with two distribution modes, under the same condition, the heating capacity of the system using variable flow distribution is higher than that of the constant flow distribution except for individual conditions (-7 DEG C / 33 DEG C, 7 DEG C / 44 DEG C). When the enthalpy difference room environment temperature is -20 DEG C, -12 DEG C, -7 DEG C, 2 DEG C and 7 DEG C, the heating capacity of the variable flow distribution is 3.2%, 2.5%, -1.1%, 2.5% and 0.1% higher than that of the constant flow distribution respectively when the outlet water temperature is the floor radiation type; the heating capacity of the variable flow distribution is 4.7%, 3.6%, 1.7%, 1.0% and 1.8% higher than that of the constant flow distribution respectively when the outlet water temperature is the fan coil type; the heating capacity of the variable flow distribution is 2.6%, 7.3%, 0%, 4.4% and -0.1% higher than that of the constant flow distribution respectively when the outlet water temperature is the radiator type; the heating capacity of the variable flow distribution is generally higher than that of the constant flow distribution, and the overall average is increased by 2.29%, which verifies the optimization effect in the low-temperature high-load scene.

[0096] The optimization method of the application divides micro elements by using the distribution parameter method, solves based on the mass, energy and momentum conservation equations and multi-physical field coupling, and realizes global optimization by combining the sequential iteration method and convergence mechanism.

[0097] The control method of the application breaks through the traditional distribution mode, solves the contradiction between the insufficient utilization rate of the effective heat exchange area of the evaporator and the safety of the system operation through the deep cooperation of the real-time sensing of the air side flow field, the multi-physical field coupling and the dynamic flow regulation, and significantly improves the energy efficiency (COP) and the heating capacity of the air source heat pump under variable conditions.

[0098] The above describes the application and its embodiments, which are not limited, and the actual embodiments are not limited thereto. In general, if a person skilled in the art is inspired by it, without departing from the purpose of the application, without creative design, similar structural modes and embodiments of the technical solutions are not creative design, which should belong to the protection scope of the application.

Claims

1. A method of optimizing control of refrigerant dynamic flow distribution, characterized in that, The method comprises the following steps: Step 1: Establishing a wind speed-refrigerant flow structure model: on the air side, the fin plane is divided into Y wind speed modules according to different air supply amounts, and each wind speed module corresponds to Z refrigerant flow paths; Step 2: Establishing a distributed parameter model of the multi-flow-path fin heat exchanger: a two-dimensional microelement model is established along the refrigerant flow direction; Step 3: The microelement model of a single refrigerant flow path is solved in sequence according to the refrigerant flow sequence, and the matching of the wind speed module and the refrigerant flow is realized according to the solving result; the method comprises the following steps: Step 3-1: respectively establishing an air side heat transfer model of a single flow path, a refrigerant side heat transfer model of a single flow path and a refrigerant side pressure drop model of a single flow path; Step 3-2: based on the known inlet parameters of the refrigerant and air of a single flow path, setting the initial mass flow rate m0 of a single flow path according to the distribution mode of the wind speed module, assuming the inlet and outlet enthalpy values and pressure values of the refrigerant according to the test working condition, and solving the microelement model of a single refrigerant flow path; Step 3-3: according to the set initial mass flow rate m0 of a single flow path and the refrigerant states at the beginning and end of each microelement, the predicted heat exchange amount of the refrigerant and the air side wind speed module in the microelement section is calculated by using the air side heat transfer model and the refrigerant side heat transfer model of a single flow path established in step 3-1 according to the mass conservation principle, the energy conservation principle, the momentum conservation principle and the heat transfer principle, taking the maximum heat exchange amount as the optimization objective and taking the flow distribution ratio as the optimization variable, the single microelement length is obtained by calculating along the path, and the total length of the single heat exchange pipe is obtained by accumulation; and the calculated total length of the single heat exchange pipe is compared with the actual total length of the single flow path; Step 3-4: iterative calculation is carried out with the convergence of the calculated total length of the single heat exchange pipe to the actual length as the criterion, and the outlet enthalpy value is adjusted: if the calculated total length of the single heat exchange pipe is greater than the actual length, it indicates that the set outlet enthalpy value of the refrigerant is too small, which needs to be increased; on the contrary, if the calculated total length of the single heat exchange pipe is less than the actual length, it indicates that the set outlet enthalpy value of the refrigerant is too large, which needs to be reduced; then, the calculation along the path is repeated until the relative error between the calculated total length of the single heat exchange pipe and the actual length is less than 1%; the calculated total length of the single heat exchange pipe is taken as the estimated pipe length; Step 3-5: the initial mass flow rate m0 of a single flow path set under the condition that the relative error between the calculated total length of the single heat exchange pipe and the actual length is less than 1% is taken as the actual mass flow rate of a single flow path, and the outlet parameters of a single flow path are output; Step 3-6: the refrigerant mass flow rates of the flow paths belonging to the same wind speed module are added to obtain the total mass flow rate of the refrigerant under the wind speed module; Step 3-7: the corresponding refrigerant flow rates of the Y wind speed modules are added to obtain the calculated total mass flow rate of the heat exchanger; Step 3-8: according to the predicted heat exchange amount and the estimated pipe length calculated, the refrigerant side pressure drop model established in step 3-1 is used to calculate the pressure drop; the outlet enthalpy value is obtained according to the calculated outlet pressure drop and temperature, the saturation temperature is obtained according to the outlet refrigerant temperature and pressure of the heat exchanger, and the superheat is calculated. Step 3-9: iterative adjustment is performed according to the criterion that the calculated value of the total mass flow of the heat exchanger converges to the actual value. Step 3-10: the distribution ratio of the air speed and the refrigerant mass flow of each air speed module is determined according to the total refrigerant mass flow corresponding to each air speed module when the relative error between the calculated value and the actual value of the total mass flow of the heat exchanger is not more than 1%.

2. The method of claim 1, wherein, In step 2, a two-dimensional microelement model is established along the flow direction of the refrigerant: each refrigerant flow path is regarded as an equal-length flow path and is extended along the two-dimensional direction, and the microelement is calculated according to the enthalpy value of the heat exchange tube.

3. The method for optimization control of refrigerant dynamic flow distribution according to claim 1, characterized in that, Step 3-2 includes the following steps: the inlet parameters of the first microelement are directly determined by the design parameters of the test working condition; based on the mass conservation principle, the energy conservation principle, the momentum conservation principle and the heat transfer principle, the air-side heat transfer amount is calculated by using the single flow path air-side heat transfer model established in step 3-1, the refrigerant-side heat transfer amount is calculated by using the single flow path refrigerant-side heat transfer model, the refrigerant-side pressure loss is calculated by using the single flow path refrigerant-side pressure drop model, and the terminal refrigerant state parameters of the first microelement are obtained through the flow enthalpy difference relationship that the air-side heat transfer amount is equal to the refrigerant-side heat transfer amount. The terminal refrigerant state parameters of the first microelement are taken as the initial refrigerant state parameters of the next microelement for iterative calculation, and the initial and terminal refrigerant state parameters of each microelement are obtained in this way, which provides boundary conditions for microelement length calculation.

4. The method for optimization control of refrigerant dynamic flow distribution according to claim 1, characterized in that, In step 3-8, if the superheat is greater than 8℃, the initial mass flow of the single flow path refrigerant m0 is increased, if the superheat is less than 3℃, the initial mass flow of the single flow path refrigerant m0 is decreased, and step 3-1 is repeated until the superheat is 3℃-8℃.

5. The method for optimization control of refrigerant dynamic flow distribution according to claim 1, characterized in that, The iterative adjustment of step 3-9 is: if the calculated value of the total mass flow of the heat exchanger is greater than the actual value, the initial mass flow of the single flow path refrigerant m0 under each air speed module is reduced by the same proportion; if the calculated value of the total mass flow of the heat exchanger is less than the actual value, the initial mass flow of the single flow path refrigerant m0 under each air speed module is increased by the same proportion, and the calculation is repeated until convergence, so that the relative error between the calculated value and the actual value of the total mass flow of the heat exchanger is not more than 1%.

6. The method of claim 1-5, wherein, According to the distribution ratio of the air speed module and the refrigerant flow, the refrigerant flow of each nozzle of the flow divider is determined.

7. The method for optimization control of refrigerant dynamic flow distribution according to claim 2, characterized in that, According to the refrigerant flow of each nozzle, the throat size of the distribution section is determined.

8. The method of claim 1-5, wherein, In step 3-4, the outlet enthalpy value is adjusted by using the dichotomy method, and the upper limit is the refrigerant inlet enthalpy, and the lower limit is the supercooled liquid enthalpy at the air inlet temperature.

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