Heat and humidity balance type two-stage temperature control and humidity regulation air conditioning system and control method

Abstract

The invention discloses a heat and humidity balance type two-stage temperature control and humidity adjustment air conditioning system and a control method, and belongs to the technical field of air constant temperature and humidity treatment, the heat and humidity balance type two-stage temperature control and humidity adjustment air conditioning system comprises a refrigeration air conditioning system, an air treatment unit, a heat and humidity balance system and a temperature fine adjustment system based on cold and heat recovery, the air handling unit is used for conducting temperature and humidity treatment on the treated air. The air conditioning system switches operation modes, namely a pure refrigeration regulation mode, a low-temperature and constant-humidity regulation mode, a medium-temperature and constant-humidity regulation mode, a high-temperature and constant-humidity regulation mode and a pure heating regulation mode. According to the heat and humidity balancer, the problems that when a traditional air conditioner conducts temperature and humidity adjustment, temperature control and humidity control interfere with each other, and ultra-precise independent control is difficult to achieve are solved, and the novel air conditioning system has the wider temperature and humidity application range and the high temperature and humidity adjustment and control precision.

Description

Technical Field

[0001] The present invention belongs to the technical field of air constant temperature and humidity processing, and in particular relates to a heat and moisture balance type two-stage temperature and humidity control air conditioning system and a control method. Background Art

[0002] With the rapid development of industrial upgrading and the information age, many high-end equipment, biomedicine, precision manufacturing and other places have increasingly stringent requirements for the precision control of indoor environment. According to the national standard "GB / T 25858 Precision Air Conditioning Unit Performance Test Method", ordinary precision air conditioners can simultaneously or independently maintain the temperature fluctuation of ±1℃~±0.5℃ and the relative humidity fluctuation of ±5%RH~±2%RH in the regulated closed area; while high-precision constant temperature and humidity air conditioners can simultaneously or independently maintain the temperature of ±0.5℃~±0.3℃ and the relative humidity of ±2%RH.

[0003] Precision air conditioning is often used in process that has strict requirements on temperature, humidity fluctuation and regional deviation control, so precision air conditioning is a typical process air conditioning. Although, ordinary process air conditioning can provide relatively accurate temperature control under normal circumstances, its temperature control stability is significantly better than conventional household air conditioning. However, in some occasions with higher requirements for ambient air quality, such as semiconductor processing, special drug production process, etc., the temperature and humidity control level of ordinary process air conditioning is still insufficient. For this reason, the patentee of this invention focuses on the research and development of ultra-high precision process air conditioning, which can achieve the control of outlet temperature at ≤±0.1℃ and relative humidity fluctuation of ≤1~2%RH.

[0004] Furthermore, how to effectively reduce energy consumption and emissions while meeting the precise control of the environment is a major challenge that energy equipment faces in achieving leapfrog development. Compared with ordinary household air conditioners, firstly, the process air conditioning system needs to simultaneously regulate multiple parameters of indoor air, and is often more complex in structure, involving a more detailed energy conversion process, and thus consumes more energy during operation. Secondly, unlike traditional air conditioners that focus on adjusting the comfort of personnel, the design focus of process air conditioners is to ensure the stability and reliability of system operation, so the fluctuation of operating conditions and the accuracy of parameter settings are more stringent, resulting in the need for air conditioning units to be frequently regulated at a higher frequency to meet the corresponding high-precision requirements, which increases power consumption. Thirdly, a significant difference from ordinary household air conditioners is that process air conditioners often need to run continuously for a long time without stopping. If the energy efficiency level of the unit is not improved, the above-mentioned high energy consumption problem will be further amplified over the years.

[0005] Furthermore, when actually applied in production scenarios, the air conditioning system must not only be able to maintain the stability of the ambient temperature and humidity in the controlled space, but also often needs to be able to handle different air conditions under a variety of working conditions, and has a wide range of applications. In a variety of application scenarios, the process air conditioning system can show good adaptability in terms of daytime and nighttime environmental changes, process flow changes, usage scenarios, and user personalized needs. Market feedback shows that the new generation of process air conditioning systems should have the ability to flexibly adjust and strong adaptability to complex occasions.

[0006] In the prior art, CN115638486B discloses a low-temperature heat compensation process air conditioning reheating system and method, which can coordinate the control of the whole system, use the switching and adjustment of valves, and parallelize the cold source system and the plate heat exchange system, and can reduce the energy consumption in the hot water heating process on the basis of meeting the heat demand of the terminal reheating section of the high-precision process industrial air conditioning system, so as to achieve the purpose of energy saving. However, there are the following problems: there is a thermal short circuit or heat reflux between the water supply pipe and the return pipe, which has a serious obstacle effect on the heat exchange efficiency, the temperature difference between the supply and return water, and the development of the buried pipe heat exchanger. Summary of the invention

[0007] The key technical routes for developing a new generation of technological air-conditioning systems with precise temperature and humidity control capabilities are: first, upgrading the control system (including algorithms); second, developing new refrigeration / heating systems (including structures). The present invention is based on the second key technical route. In order to solve the above-mentioned problems, the present invention proposes: a heat and moisture balance type two-stage temperature and humidity control air-conditioning system and control method. The setting of the heat and moisture balancer overcomes the problems of mutual interference between temperature and humidity control and difficulty in achieving ultra-precision independent control in traditional air conditioners when performing temperature and humidity adjustment, so that the new air conditioning system has a wider temperature and humidity application range and higher temperature and humidity control accuracy. The design of two-stage temperature and humidity control enables the treated air to reach the set target temperature value more quickly and accurately. The optimized design of the energy flow of the entire air conditioning system combined with the matching cold and heat recovery scheme has effectively improved the overall energy efficiency of the system.

[0008] The invention discloses a heat-moisture balance type two-stage temperature-control and humidity-regulating air-conditioning system, which mainly comprises the following subsystems: a refrigeration and air-conditioning system, an air handling unit, a heat-moisture balance system and a temperature fine-tuning system based on cold and heat recovery.

[0009] The refrigeration and air-conditioning system mainly includes a compressor, a main evaporator, a main condenser, an auxiliary evaporator, an auxiliary condenser, a heat balancer, a wet balancer, a throttling expansion device and a thermostat. The refrigeration and air-conditioning system provides the required cooling or heat for the dehumidifier, the temperature increaser, the recooler and the reheater of the air handling unit. The dehumidifier is used to cool and dehumidify the treated air; the temperature increaser is used to quickly adjust the temperature of the treated air to near the set value; the recooler is used to recool the air whose outlet temperature is higher than the set value after passing through the temperature increaser, and the reheater is used to reheat the air whose outlet temperature is lower than the set value after passing through the recooler. In addition, a plurality of throttle valves, three-way valves, four-way reversing valves and one-way stop valves are reasonably arranged on the system pipeline to control the flow direction of the refrigerant and the migration and heat exchange characteristics of the cold and hot streams.

[0010] A control method for a heat and moisture balance two-stage temperature and humidity control air conditioning system can realize five control requirements for the treated air, namely: pure cooling control mode, low temperature constant humidity control mode, medium temperature constant humidity control mode, high temperature constant humidity control mode and pure heating control mode.

[0011] First, in the pure refrigeration control mode, the air treatment process only has refrigeration requirements, but no constant humidity requirements. The high-temperature and high-pressure gaseous refrigerant at the compressor outlet passes through the oil separator and the first four-way reversing valve, and then enters the heat balancer, where it is converted from gas to liquid after releasing heat. The condensed refrigerant returns to the first four-way reversing valve, and then flows to the first expansion valve through the first electric three-way valve. Under the throttling effect of the first expansion valve, the pressure and temperature of the refrigerant are greatly reduced. The low-temperature and low-pressure gas-liquid two-phase (mainly liquid phase) refrigerant flows into the branches where the main evaporator and the auxiliary evaporator are located in a certain proportion through the second electric three-way valve. For the stream flowing into the branch where the main evaporator is located, the low-temperature and low-pressure refrigerant enters the main evaporator, absorbs the heat in the air to be treated and vaporizes (the air to be treated is subjected to primary refrigeration). The refrigerant leaving the main evaporator is in a superheated vapor state and merges with the refrigerant diversion stream from the branch where the auxiliary evaporator is located. This is the confluence point of the two gaseous refrigerants. The flow stream flowing into the branch where the auxiliary evaporator is located is adjusted by the second four-way reversing valve and the third electric three-way valve, and the low-temperature and low-pressure refrigerant enters the auxiliary evaporator, where it further absorbs the heat in the air to be treated (secondary refrigeration of the air to be treated). The refrigerant converted into superheated vapor after absorbing heat merges with the refrigerant vapor stream from the branch where the main evaporator is located. The merged refrigerant vapor passes through the gas-liquid separator and enters the compressor in a completely gaseous state and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core function of the system in the pure refrigeration control mode can be realized.

[0012] Secondly, in the low temperature and constant humidity control mode, the temperature and humidity are required to be controlled together, and a lower temperature is required. The high temperature and high pressure gaseous refrigerant at the outlet of the compressor is divided into two streams after passing through the oil separator, and flows into the main condenser and the branch where the heat balancer is located respectively. For the stream flowing into the branch where the heat balancer is located, the high temperature and high pressure gaseous refrigerant is turned back by the first four-way reversing valve and enters the heat balancer, changing from gas to liquid. The cooled refrigerant liquid returns to the first four-way reversing valve and flows to the first expansion valve through the first electric three-way valve. Under the throttling effect of the first expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and then merged with the gas-liquid two-phase refrigerant stream from the branch where the main condenser is located. For the stream flowing into the branch where the main condenser is located, the high temperature and high pressure gaseous refrigerant first directly enters the main condenser to release the condensation heat, changes from gas to liquid, and then flows to the second expansion valve. Under the throttling effect of the second expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the gas-liquid two-phase refrigerant split stream from the branch where the heat balancer is located. The mixed refrigerant enters the main evaporator through the second electric three-way valve, and absorbs the heat in the treated air and gasifies (primary treatment of the treated air: refrigeration), converting it into a refrigerant vapor with a certain degree of superheat, and then passes through the gas-liquid separator, enters the compressor in a completely gaseous state and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core function of the system under the low-temperature constant humidity control mode can be realized. The difference from the above-mentioned pure refrigeration control mode is that the precision air conditioning system under the low-temperature constant humidity control mode uses the main condenser to recover part of the refrigerant condensation heat, and uses it to increase the temperature and adjust the relative humidity of the treated air after it is cooled and humidified (absolute moisture content is reduced and relative humidity is increased) when passing through the main evaporator (secondary treatment of the treated air: heating).

[0013] Third, in the medium temperature constant humidity control mode, it is required to control the air temperature and humidity together, and the difference is not much relative to the room temperature. The high temperature and high pressure gaseous refrigerant at the outlet of the compressor is divided into two streams after passing through the oil separator, and flows into the main condenser and the branch where the heat balancer is located respectively. For the stream flowing into the branch where the heat balancer is located, the high temperature and high pressure gaseous refrigerant is turned back by the first four-way reversing valve, enters the heat balancer, and releases part of the condensation heat to the external environment. After completing the heat unloading, the refrigerant returns to the first four-way reversing valve, enters the auxiliary condenser through the first electric three-way valve, releases the condensation heat into the air duct, and converts it into a liquid state. After the liquid refrigerant passes through the drying filter, it flows to the third expansion valve. Under the throttling effect of the third expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and then merges with the refrigerant diversion stream from the branch where the main condenser is located. For the stream flowing into the branch where the main condenser is located, the high temperature and high pressure gaseous refrigerant first enters the main condenser to release the condensation heat, changes from gas to liquid, and then flows to the second expansion valve. Under the throttling effect of the second expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the low-temperature and low-pressure refrigerant split stream from the branch where the heat balancer is located. The mixed refrigerant stream enters the main evaporator through the second electric three-way valve, and absorbs the heat in the air to be treated to achieve the first-level cooling and debugging treatment (first-level refrigeration treatment of the air to be treated). It is gasified into a refrigerant vapor with a certain degree of superheat, passes through the gas-liquid separator, enters the compressor in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. In this way, the basic core function of the system under the medium-temperature constant humidity control mode can be realized. The difference from the low-temperature constant humidity control mode is that the air-conditioning system under the medium-temperature constant humidity control mode uses the condensation heat released by the main condenser (first-level heating treatment of the air to be treated), and then introduces the heat after the heat unloading of the heat balancer into the air duct through the auxiliary condenser (secondary heating treatment of the air to be treated). At the same time, the heat unloading amount can be controlled by adjusting the heat exchange conditions on the heat balancer side and the opening of the first electric three-way valve.

[0014] Fourth, in the high-temperature constant-humidity control mode, the air temperature and humidity need to be controlled jointly, and the outlet temperature is significantly higher than the inlet. The high-temperature and high-pressure gaseous refrigerant at the outlet of the compressor is divided into two streams after passing through the oil separator, and flows into the main condenser and the branch where the auxiliary condenser is located respectively. For the stream flowing into the branch where the main condenser is located, the high-temperature and high-pressure gaseous refrigerant condenses and releases heat therein (performs a primary heating treatment on the air to be treated), and flows to the second expansion valve after changing from gas to liquid. After throttling by the second expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and then merged with the refrigerant diversion stream from the branch where the auxiliary condenser is located. For the stream flowing into the branch where the auxiliary condenser is located, the high-temperature and high-pressure gaseous refrigerant enters the auxiliary condenser through the first four-way reversing valve and the first electric three-way valve to condense and release heat (performs a secondary heating treatment on the air to be treated), and changes from gas to liquid. After passing through the drying filter, the liquid refrigerant flows to the third expansion valve. After throttling by the third expansion valve, the pressure and temperature of the refrigerant are greatly reduced and transformed into a gas-liquid two-phase state (mainly the liquid phase). Then it merges with the refrigerant diversion stream from the branch where the main condenser is located. The mixed gas-liquid two-phase fluid flows into the main evaporator and the branch where the wet balancer is located in a certain proportion through the second electric three-way valve. For the stream flowing into the branch where the main evaporator is located, the low-temperature and low-pressure refrigerant enters the main evaporator, and absorbs the heat in the air to be treated to achieve a first-level cooling and humidity control treatment (a first-level refrigeration treatment of the air to be treated). The vapor leaving the main evaporator merges with the refrigerant vapor diversion stream from the branch where the wet balancer is located. This is the confluence point of the two refrigerant vapors. The flow stream flowing into the branch where the wet balancer is located is diverted by the second four-way reversing valve, and the low-temperature and low-pressure refrigerant enters the wet balancer, absorbs heat from the external environment and vaporizes into superheated steam, and then passes through the second four-way reversing valve and the third electric three-way valve in turn, and merges with the refrigerant vapor diversion stream from the branch where the main evaporator is located. The merged refrigerant vapor passes through the gas-liquid separator and enters the compressor in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core function of the system in the high-temperature constant humidity control mode can be realized. The difference from the medium-temperature constant humidity control mode lies in the role played by the wet balancer: from the perspective of humidity control, the air-conditioning system in the high-temperature constant humidity control mode releases part of the cold air to the outside through the wet balancer, solving the problem of low absolute humidity content of the air caused by the large-scale cooling treatment at the front end, which needs to be solved when the system subsequently increases the outlet temperature of the treated air; from the perspective of temperature control, the air-conditioning system in the high-temperature constant humidity control mode absorbs heat from the external environment through the wet balancer, and the enthalpy value of the refrigerant increases, which is beneficial to improve the heating and temperature-raising capacity of the treated air.

[0015] Fifth, in the pure heating control mode, the air treatment process only has heating requirements, but no constant humidity requirements. The high-temperature and high-pressure gaseous refrigerant at the outlet of the compressor is divided into two streams after passing through the oil separator, and flows into the main condenser and the branch where the auxiliary condenser is located respectively. For the stream flowing into the branch where the main condenser is located, the high-temperature and high-pressure gaseous refrigerant directly enters the main condenser to condense and release heat (perform primary heating treatment on the air to be treated), and flows to the second expansion valve after changing from gas to liquid. After throttling by the second expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and it is converted into a gas-liquid two-phase state (mainly liquid phase). The low-temperature and low-pressure refrigerant in the gas-liquid two-phase state merges with the refrigerant split stream from the branch where the auxiliary condenser is located. For the stream flowing into the branch where the auxiliary condenser is located, the high-temperature and high-pressure gaseous refrigerant enters the auxiliary condenser through the first four-way reversing valve and the first electric three-way valve to release the condensation heat (perform secondary heating treatment on the air to be treated), and changes from gas phase to liquid phase. After passing through the drying filter, the liquid refrigerant flows to the third expansion valve. After throttling by the third expansion valve, the pressure and temperature of the refrigerant are greatly reduced, and it is converted into a gas-liquid two-phase state (mainly the liquid phase), and then merges with the refrigerant diversion stream from the branch where the main condenser is located. The mixed gas-liquid two-phase fluid enters the wet balancer after passing through the second electric three-way valve and the second four-way reversing valve, and absorbs heat from the external environment and vaporizes into superheated steam. Subsequently, the refrigerant vapor returns to the second four-way reversing valve and flows to the gas-liquid separator through the third electric three-way valve. After the action of the gas-liquid separator, it enters the compressor in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core functions of the system in the pure heating control mode can be realized. The difference from the high temperature constant humidity control mode is that in the pure heating control mode, after the refrigerants in the main condenser and the auxiliary condenser are throttled and mixed, all the main refrigerants must flow to the wet balancer through the second electric three-way valve and the second four-way reversing valve, and all the cold loads generated by the system cycle are unloaded to the external environment through the wet balancer, and at the same time, heat is absorbed from the external environment to increase the enthalpy value of the refrigerant. Thus, the entire air conditioning system can exert the maximum heating capacity under its own structure.

[0016] Preferably, the air handling unit may include an air volume regulating section, an air filtering section, a dehumidifying section, a heating section, a recooling section and a reheating section. The air volume regulating section controls the amount of air that needs to be treated by the size of the opening of the fresh air inlet; the filtering section first performs a primary filtration on the input fresh air to prevent the impurities and dust therein from affecting the accuracy of the subsequent air treatment process, while ensuring the cleanliness of the air handling unit; then performs a medium-efficiency filtration to further filter the impurities therein to meet the requirements of air cleanliness in the production and life of personnel. The functions realized by the dehumidifying section, the heating section, the recooling section and the reheating section are also different under different control requirements. The dehumidifying section, when a low-temperature and low-pressure refrigerant flows through the main evaporator, can cool the air to be treated flowing through it. If the air temperature drops below the dew point temperature, the water molecules in the air condense into liquid water on the surface of the main evaporator, resulting in a decrease in the water vapor content, thereby realizing the dehumidification function; the dehumidifying section does not function in the pure heating control mode. The temperature rise section can heat and raise the temperature of the treated air when high-temperature and high-pressure refrigerant flows through the main condenser 7 and condenses and releases heat therein; the temperature rise section does not take effect under the control requirements of pure refrigeration. The temperature fine-tuning system based on cold and heat recovery includes a recooler and a reheater, which are arranged at the end of the air handling unit to achieve precise adjustment of the outlet temperature of the treated air. The reheat section can heat and raise the temperature of the treated air for the second time when high-temperature and high-pressure refrigerant flows through the auxiliary condenser and condenses and releases heat therein. In the recooling section, in the pure refrigeration control mode, low-temperature and low-pressure refrigerant flows through the auxiliary evaporator, which can perform secondary cooling on the air passing through.

[0017] Preferably, the heat and moisture balance system includes a heat balancer and a moisture balancer. In order to improve the air conditioning system's ability to control the outlet temperature and humidity of the processed air, in pure refrigeration, low temperature constant humidity and medium temperature constant humidity control modes, the balancer is used to release part or all of the condensation heat to the outside world; in high temperature constant humidity and pure heating control, the moisture balancer is used to unload part or all of the cold load to the outside world. It should be noted that due to the unloading of part of the cold capacity, the system's ability to achieve absolute moisture content dehumidification by cooling is reduced. At this time, it is necessary to increase the circulation flow of the refrigerant accordingly, enhance the heat exchange capacity of the main condenser and the auxiliary condenser, so that the temperature of the gas to be treated can be further increased to meet the requirement of relative humidity stability.

[0018] The beneficial effects of the present invention are as follows: The present invention proposes: a heat and moisture balance type two-stage temperature and humidity control air conditioning system and control method. The setting of the heat and moisture balancer overcomes the problem of mutual interference between temperature and humidity control and difficulty in achieving ultra-precision independent control in traditional air conditioners when performing temperature and humidity adjustment, so that the new air conditioning system has a wider temperature and humidity application range and higher temperature and humidity control accuracy. The two-stage temperature and humidity control design enables the treated air to reach the set target temperature value more quickly and accurately. The optimized design of the energy flow of the entire air conditioning system combined with the matching cold and heat recovery solution effectively improves the overall energy efficiency of the system. BRIEF DESCRIPTION OF THE DRAWINGS

[0019] Figure 1 This is a flow chart of the constant temperature and humidity air conditioning system of the present invention during pure refrigeration control;

[0020] Figure 2 This is a flow chart of the constant temperature and humidity air conditioning system of the present invention during pure heating regulation;

[0021] Figure 3 This is a flow chart of the constant temperature and humidity air conditioning system of the present invention during low temperature and constant humidity control;

[0022] Figure 4 This is a flow chart of the constant temperature and humidity air conditioning system of the present invention during medium temperature and humidity control;

[0023] Figure 5 This is a flow chart of the constant temperature and humidity air conditioning system of the present invention during high temperature and humidity control;

[0024] Figure 6 It is a structural schematic diagram of the air handling unit of the present invention.

[0025] As shown in the figure: 1. compressor; 2. oil separator; 3. first four-way reversing valve; 4. heat balancer; 5. first electric three-way valve; 6. first expansion valve; 7. main condenser; 8. second expansion valve; 9. auxiliary condenser; 10. drying filter; 11. third expansion valve; 12. second electric three-way valve; 13. main evaporator; 14. gas-liquid separator; 15. second four-way reversing valve; 16. wet balancer; 17. third electric three-way valve; 18. auxiliary evaporator; 19. switch valve; 20. one-way stop valve; 21. air supply regulator; 22. primary air filter; 23. medium-efficiency air filter; 24. fan. DETAILED DESCRIPTION

[0026] The technical solution of the present invention will be described below in detail with reference to the accompanying drawings in specific embodiments of the present invention, which is only used to explain the present invention and should not be construed as limiting the present invention.

[0027] A heat and moisture balance type two-stage temperature and humidity control air conditioning system with condensation heat recovery capability comprises a refrigeration and air conditioning system, an air handling unit, a heat and moisture balance system and a temperature fine-tuning system based on cold and heat recovery. The refrigeration and air conditioning system provides the required cooling or heat to the air handling unit, the air handling unit performs temperature and humidity treatment on the treated air, the heat and moisture balance system regulates the outlet temperature and humidity of the treated air, and the temperature fine-tuning system based on cold and heat recovery is arranged at the end of the air handling unit to regulate the outlet temperature of the treated air.

[0028] Among them, Figure 1 , Figure 6As shown, the refrigeration and air-conditioning system includes a compressor 1, an oil separator 2, a first four-way reversing valve 3, a heat balancer 4, a first electric three-way valve 5, a first expansion valve 6, a main condenser 7, a second expansion valve 8, an auxiliary condenser 9, a drying filter 10, a third expansion valve 11, a second electric three-way valve 12, a main evaporator 13, a gas-liquid separator 14, a second four-way reversing valve 15, a wet balancer 16, a third electric three-way valve 17, an auxiliary evaporator 18, a switch valve 19, an air supply regulator 21, a primary air filter 22, a medium-efficiency air filter 23, and a fan 24; wherein the output end of the compressor 1 is connected to the input end of the oil separator 2 through a pipeline, and the output end of the oil separator 2 is divided into two branches, one of which is a branch. A branch pipe is connected to the main condenser 7, and another branch pipe is connected to the first four-way reversing valve 3, which is the diversion point 1; for the pipeline where the first four-way reversing valve 3 is located, the output end of the first four-way reversing valve 3 is connected to the input end of the heat balancer 4, the output end of the heat balancer 4 and the first electric three-way valve 5; by adjusting the first four-way reversing valve 3, the system can control whether to unload the heat in the high-temperature and high-pressure refrigerant vapor; then, the refrigerant flows into two branches in a certain proportion through the first electric three-way valve 5, one branch is connected to the first expansion valve 6, and the other branch is connected to the auxiliary condenser 9, which is the diversion point 2; for the pipeline where the auxiliary condenser 9 is located, the output end of the auxiliary condenser 9 is connected to the drying filter 10, the drying filter The output end of the filter 10 is then connected to the third expansion valve 11; the output end of the third expansion valve 11 is connected to the fourth one-way stop valve, and then merges with the pipeline where the first expansion valve 6 is located, which is the confluence point 2; for the pipeline where the first expansion valve 6 is located, the other end of the first expansion valve 6 is connected to the third one-way stop valve, and then merges with the pipeline where the auxiliary condenser 9 is located, which is the confluence point 2; after the two pipelines are connected, they merge with the pipeline where the main condenser 7 is located through the second one-way stop valve, which is the confluence point 1; for the pipeline where the main condenser 7 is located, when the switch valve 19 is in the open state, the high-temperature and high-pressure refrigerant enters the main condenser 7 and releases heat to the air in the air duct, thereby increasing the temperature of the treated air; the other end of the main condenser 7 The end is connected to the second expansion valve 8, and then merges with the pipeline where the second one-way stop valve is located through the first one-way stop valve, which is the confluence point 1; the two-phase refrigerant in each section of the pipeline merges and flows to the second electric three-way valve 12; the output end of the second electric three-way valve 12 is divided into two branches, one branch is connected to the main evaporator 13, and the other branch is connected to the second four-way reversing valve 15, which is the diversion point 3; for the pipeline where the second four-way reversing valve 15 is located, the output end of the second four-way reversing valve 15 is connected to the input end of the wet balancer 16, the output end of the wet balancer 16 and the third electric three-way valve 17; by adjusting the second four-way reversing valve 15, the system can control whether to unload the cold capacity in the low-temperature and low-pressure refrigerant;Subsequently, the refrigerant flows into two branches in a certain proportion through the third electric three-way valve 17, one branch is connected to the sixth one-way stop valve, and the other branch is connected to the auxiliary evaporator 18, which is the diversion point 4; for the pipeline where the auxiliary evaporator 18 is located, the output end of the auxiliary evaporator 18 is connected to the seventh one-way stop valve, and then merges with the pipeline where the sixth one-way stop valve is located, which is the confluence point 4; for the pipeline where the sixth one-way stop valve is located, the other end of the six one-way stop valves is merged with the pipeline where the seventh one-way stop valve is located, which is the confluence point 4; after the two pipelines are connected, they are connected to the pipeline where the main evaporator 13 is located. The pipelines of the main evaporator 13 are merged, which is the confluence point 3; for the pipeline where the main evaporator 13 is located, the other end of the main evaporator 13 is merged through the common pipeline of the pipeline where the fifth one-way stop valve is located and the pipeline where the sixth one-way stop valve is located and the pipeline where the seventh one-way stop valve is located, which is the confluence point 3; after the superheated refrigerant in each section of the pipeline is merged, it passes through the gas-liquid separator 14 and returns to the compressor 1 in a completely gaseous state; the air supply regulator 21, the primary air filter 22, the medium-efficiency air filter 23, the main evaporator 13, the main condenser 7, the auxiliary evaporator 18, the auxiliary condenser 9, and the fan 24 are arranged on the air duct. ;

[0029] The quantitative analysis of the following specific embodiments is based on a certain evaporator and condenser model. After the specifications and dimensions of the heat exchanger are determined according to relevant rules, the air outlet state after the processed air and the refrigerant in the corresponding inlet state are simulated and calculated. The main assumptions and specific control equations based on it are as follows:

[0030] The following main assumptions are used when modeling the evaporator: (1) The refrigerant and air are in countercurrent; (2) The radial temperature of the tube wall is consistent. For each microelement along the tube length, the physical properties of the refrigerant side, the air side, and the tube wall are considered to be consistent, and the thermal resistance of the tube wall is not considered; (3) Considering that the superheated zone is short and the acceleration pressure drop is small, the pressure drop in the superheated zone is ignored.

[0031] In the model, the evaporator includes two phase zones: the two-phase zone and the superheat zone, and each root zone is divided into several micro-elements. For the two-phase zone, the temperature change depends on the pressure drop, and the refrigerant enthalpy value changes greatly. Therefore, the micro-element division is achieved by dividing the enthalpy difference equally; for the superheat zone, under the assumption that the pressure remains unchanged, the refrigerant temperature changes greatly, and the micro-element is divided by dividing the refrigerant temperature equally.

[0032] The following specific governing equations are used when modeling the evaporator:

[0033] Air side energy equation:

[0034] Q a =m a (h a,in -h a,out ) (1)

[0035] Moist air humidity equation:

[0036] -m a dW a =h d (W a -W w )dA a (2)

[0037] Where W w W is the saturated humidity content of moist air corresponding to the temperature of the outer wall of the tube; a To calculate the average moisture content on the air side of the microelement.

[0038] Refrigerant side energy equation:

[0039] Q r =m r (h r1 -h r2 )=α i A i (T w -T rm ) (3)

[0040] In the formula, α i is the heat transfer coefficient on the refrigerant side, A i is the inner surface area of ​​the tube, T w is the tube wall temperature, T rm is the average refrigerant temperature.

[0041]

[0042] For the superheated area, the refrigerant side heat transfer coefficient α i,SH Calculated by the Dittus-Boeler heat transfer correlation:

[0043] Nu i =0.023Re 0.8 Pr 0.4 (5)

[0044] In the formula, G r is the refrigerant mass flow density, d i is the inner diameter of the tube.

[0045] When the refrigerant is in a two-phase state, there are two flow patterns in the process of boiling flow in the tube: mist flow and annular flow. The transition of the flow pattern is determined by the critical dryness:

[0046] x cr =7.94[Re vs ×(2.03×10 4 R vs-0.81 ΔT-1)] -0.161 (6)

[0047] In the formula, Re vs is the Reynolds number when the gaseous refrigerant flows alone in the tube. cr When , the two-phase refrigerant flow pattern is annular flow, otherwise it is mist flow. The heat transfer coefficient of annular flow in the two-phase region of the evaporator is;

[0048]

[0049] Where Nu is the Nusselt number of the refrigerant; P rl is the Prandtl number of the refrigerant liquid phase; μ v With μ l are the dynamic viscosity coefficients of the refrigerant saturated vapor phase and saturated liquid phase respectively; ρ v and ρ l are the densities of the saturated vapor phase and saturated liquid phase of the refrigerant respectively; C pl is the isobaric specific heat capacity of the saturated refrigerant liquid; ΔT is the temperature difference between the refrigerant and the pipe wall.

[0050] The refrigerant gas and liquid in the mist flow are evenly mixed and can be approximated as homogeneous flow. At this time, the refrigerant heat transfer coefficient is:

[0051]

[0052] In the formula, λ v is the thermal conductivity of saturated refrigerant vapor; G m is the mass flow rate of the homogeneous flow of refrigerant; d is the inner diameter of the heat exchange tube.

[0053] Under the actual working conditions of the evaporator, the refrigerant not only absorbs the heat from the air flowing through it, but also absorbs some heat conducted through the pipe wall and other components. Therefore, the heat exchange between the refrigerant side and the air side is not completely equal. In addition, the refrigerant will inevitably contain lubricating oil. When measuring the refrigerant flow rate, the influence of the lubricating oil factor must be considered, resulting in the heat exchange obtained by multiplying the flow rate by the enthalpy difference being larger than the actual heat exchange rate. Considering the above factors and based on a large amount of experimental data, the following relationship is concluded between the two:

[0054] Q a =γQ r (11)

[0055] In the formula, the value of γ depends on the specific situation. It can generally be taken as 0.9, and 0.85 is taken in this simulation model.

[0056] Micro-element heat transfer equation:

[0057]

[0058] Wall length:

[0059]

[0060] Refrigerant side pressure drop:

[0061]

[0062] Air side pressure drop:

[0063]

[0064] Where Q, h, T and m are heat transfer, enthalpy, temperature and mass flow rate respectively; A i is the heat exchange area on the refrigerant side; A o is the total surface area on the air side; A c is the cross-sectional area of ​​air flow; σ is the ratio of the minimum flow area to the windward area; G c For the definition in A c The air mass flux on the tube; the subscript a represents the air side, r represents the refrigerant side, i represents the inside of the tube (refrigerant side), and in and out represent the inlet and outlet, respectively.

[0065] Mass transfer coefficient h d , calculated as follows:

[0066]

[0067] Where Le is the Lewis number, R a and c p,a are the air side thermal resistance and air specific heat capacity respectively.

[0068] By integrating equation (4), the humidity content of the air at the outlet of the microelement can be derived:

[0069]

[0070] U is the total heat transfer coefficient based on the heat exchange area on the refrigerant side, and the calculation formula is:

[0071]

[0072] In the formula, α i is the heat transfer coefficient of the refrigerant side surface; α o is the heat transfer coefficient of the air side surface; R w is the thermal resistance of the tube wall (including fins); A i / A o is the ratio of the heat exchange area inside and outside the evaporator tube. For each infinitesimal element, it can be considered that A i / A o They are equal to the ratio of the total heat exchange area inside and outside the evaporator tube.

[0073] The following main assumptions are used when modeling the condenser: (1) The condenser is a countercurrent heat exchanger; (2) The thermal resistance of the tube wall is negligible; (3) The actual refrigerant flow in the tube is a complex phase-separated flow, which is simplified to a one-dimensional homogeneous flow model to increase the calculation speed; (4) The flow of air outside the tube is also considered as a one-dimensional flow.

[0074] In the process of condenser modeling, it is divided into three phase zones (superheat zone, two zones, and subcooling zone) and treated separately. For the single-phase zone, i.e., the superheat zone and the subcooling zone, the micro-element division is evenly divided according to the temperature drop on the refrigerant side; for the two-phase zone, since the temperature remains unchanged, the heat exchange is manifested as a change in enthalpy value, and the micro-element division can be evenly divided according to the enthalpy difference of the refrigerant in the two-phase zone.

[0075] The following specific governing equations are used when modeling the condenser:

[0076] Air side energy equation:

[0077] Q a =m a (h a,out -h a,in )(twenty two)

[0078] Refrigerant side energy equation:

[0079] Q r =m r (h r,in -h r,out )(twenty three)

[0080] The energy balance equation of the refrigerant in the condenser and the correlation equation of the heat transfer coefficient in the single-phase zone (including the superheat zone and the subcooling zone) refer to equations (5) and (7) respectively. Unlike the evaporator, the refrigerant flow pattern in the horizontal tube of the condenser can be divided into three forms according to its mass flow rate: mist flow, annular flow and wavy flow. The transition criteria between different flow patterns are usually described by the Froude number Fr and the Weber number We.

[0081] The heat transfer coefficients of annular flow and mist flow in the two-phase region of the condenser are obtained by equations (9) and (10) respectively. For wavy flow, the heat transfer coefficient correlation equation can be expressed as:

[0082]

[0083] Energy balance equation on both sides:

[0084] Q a =ξQ r (25)

[0085] Micro-element heat transfer equation:

[0086]

[0087] Wall length:

[0088]

[0089] Where Q, h, T and m are heat transfer, enthalpy, temperature and mass flow rate respectively; the heat leakage coefficient ξ is taken as 0.85.

[0090] U is the total heat transfer coefficient based on the heat exchange area on the refrigerant side, which satisfies the following relationship:

[0091]

[0092] In the formula, α i is the heat transfer coefficient of the refrigerant side surface; α o is the heat transfer coefficient of the air side surface; R w is the thermal resistance of the tube wall (including fins); A i / A o is the ratio of the heat exchange area inside and outside the evaporator tube. For each infinitesimal element, A i / A o They are equal to the ratio of the total heat exchange area inside and outside the evaporator tube.

[0093] In Specific Embodiment 1 and Specific Embodiment 2, according to the state parameters of the refrigerant R410A at the inlet and outlet of each heat exchanger and the parameters of the air to be treated on the inlet side, based on the evaporator and condenser model established by the aforementioned algorithm, the specification parameters of the heat exchanger and the parameters of the treated outlet air that meet the relevant working conditions are obtained by dividing the microelement and iterating multiple times. In the specific implementation, the evaporator and the condenser both use finned tube heat exchangers, the fin type is flat fins, the fin spacing is 0.18cm, and the fin thickness is 0.012cm. In Specific Embodiment 3, Specific Embodiment 4 and Specific Embodiment 5, according to the corresponding specification heat exchangers obtained in Specific Embodiment 1 and Specific Embodiment 2, through the thermodynamic parameters, circulation flow rate and inlet air parameters of the refrigerant R410A at the inlet of the heat exchanger, based on the evaporator and condenser model established by the aforementioned algorithm, the thermodynamic parameters of the refrigerant at the outlet of each heat exchanger and the outlet air parameters are obtained by dividing the microelement and iterating multiple times.

[0094] Specific implementation method 1, as Figure 1As shown, the embodiment of the present invention provides an operation mode of a heat and moisture balance type two-stage temperature and humidity control system under the demand of pure refrigeration control. In the specific implementation, the compressor 1 uses a vertical scroll compressor. After the high-temperature and high-pressure gaseous refrigerant compressed by the compressor 1 leaves the exhaust port of the compressor 1, it passes through the oil separator 2 and the first four-way reversing valve 3, and enters the heat balancer 4 to release the condensation heat and convert from gas to liquid. In this specific implementation, the heat balancer 4 adopts a fin-tube heat exchanger, and the dry mass flow rate of the refrigerant is 0.01389 kg / s. The condensed refrigerant returns to the first four-way reversing valve 3, and then enters the first expansion valve 6 through the first electric three-way valve 5. In this specific implementation, the first expansion valve 6 adopts an electronic expansion valve. The low-temperature and low-pressure refrigerant after throttling by the first expansion valve 6 passes through the second electric three-way valve 12 and flows into the branch where the main evaporator 13 and the auxiliary evaporator 18 are located at a ratio of 50%:50%. In this specific implementation, the main evaporator 13 and the auxiliary evaporator 18 both adopt fin-tube heat exchangers. For the branch where the main evaporator 13 is located, the low-temperature and low-pressure refrigerant enters the main evaporator 13, and absorbs the heat in the air to be treated (in this specific embodiment, the temperature of the inlet of the air to be treated is 30°C, the relative humidity is 20% RH, and the flow rate is 2m / s) and vaporizes (the air to be treated is subjected to primary refrigeration), and the refrigerant after absorbing heat is converted into a refrigerant vapor with a superheat of 5K, and then merges with the refrigerant vapor of the branch where the auxiliary evaporator 18 is located. For the branch where the auxiliary evaporator 18 is located, the low-temperature and low-pressure refrigerant enters the auxiliary evaporator 18 through the second four-way reversing valve 15 and the third electric three-way valve 17, and further absorbs the heat in the air and vaporizes (the air to be treated is subjected to secondary refrigeration), and is converted into a refrigerant vapor with a superheat of 5K, and then merges with the refrigerant vapor of the branch where the auxiliary evaporator 13 is located. After the merged refrigerant vapor is processed by the gas-liquid separator 14, it returns to the compressor 1 in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If the cycle works like this, the basic core functions of the system in pure refrigeration control mode can be realized.

[0095] Furthermore, a performance evaluation is carried out under the following operating conditions of the air-conditioning system: the evaporation pressure of the system is 0.7 MPa and the condensation pressure is 2.2 MPa; the vapor superheat of the refrigerant at the outlet of the main evaporator 13 and the auxiliary evaporator 18 is 5K; and the liquid subcooling of the refrigerant at the outlet of the heat balancer 4 is 2K. According to the mass flow rate, evaporation pressure, inlet enthalpy, outlet superheat of the refrigerant in the system and the inlet parameters of the air side, the single tube length of the main evaporator 13 is solved to be 0.0887m, 2 branch flow channels, 5 rows of tubes are arranged in total, 6 single tubes are arranged in each row of tubes, the inner diameter of the tube is 0.009m, the outer diameter of the tube is 0.01m, the tube spacing is 0.027m, and the row spacing is 0.022m; the single tube length of the auxiliary evaporator 18 is 0.1568m, 2 branch flow channels, 5 rows of tubes are arranged in total, 6 single tubes are arranged in each row of tubes, the inner diameter of the tube is 0.009m, the outer diameter of the tube is 0.01m, the tube spacing is 0.027m, and the row spacing is 0.022m. After the aforementioned refrigerant mass flow, condensing pressure, inlet enthalpy, outlet subcooling and inlet parameters of the air side, the single tube length of the heat balancer 4 is 1.0864m, 3 branch flow channels, 3 rows of tubes are arranged in total, each row of tubes has 6 single tubes, the inner diameter of the tube is 0.009m, the outer diameter of the tube is 0.0095m, the tube spacing is 0.020m, and the row spacing is 0.0217m.

[0096] In the specific implementation, the air to be treated first flows through the main evaporator 13, and then flows through the auxiliary evaporator 18, and is treated in two stages in sequence. Therefore, the outlet state condition of the air treated by the main evaporator 13 can be used as the inlet state condition of the auxiliary evaporator 18. According to the system simulation modeling analysis, under the aforementioned heat exchanger specifications, the outlet temperature of the air to be treated after being treated by the main evaporator 13 is 15.02°C and the relative humidity is 49.7%RH; and then after the secondary cooling by the auxiliary evaporator 18, its outlet temperature is 7.28°C and the relative humidity is 81.3%RH. From the perspective of the energy efficiency of the air conditioning system, at this time, the refrigerant heat exchange power on the main evaporator 13 side and the auxiliary evaporator 18 side is 1180.67W, and the system's refrigeration performance coefficient (COP) is 5.22.

[0097] Specific implementation method 2, as Figure 2As shown, the embodiment of the present invention provides an operation mode of a heat and moisture balance two-stage temperature and humidity control system under the demand of pure heating control. In the specific implementation, the high-temperature and high-pressure gaseous refrigerant at the outlet of the compressor 1 is divided into two streams after passing through the oil separator 2, and flows into the branches where the main condenser 7 and the auxiliary condenser 9 are located respectively. In this specific embodiment, the fin-tube heat exchanger used to enter the main condenser 7 and the auxiliary condenser 9 is 0.01389 kg / s. The refrigerant flow ratios of the main condenser 7 branch and the auxiliary condenser 9 branch are 50% and 50% respectively (i.e. 0.5:0.5). For the stream flowing into the branch where the main condenser 7 is located, the high-temperature and high-pressure gaseous refrigerant directly enters the main condenser 7 to condense and release heat (the air to be treated is subjected to a first-level heating treatment), and flows to the second expansion valve 8 after changing from gas to liquid. After throttling by the second expansion valve 8, the pressure and temperature of the refrigerant are greatly reduced, and it is converted into a gas-liquid two-phase state. The low-temperature and low-pressure refrigerant in the gas-liquid two-phase state merges with the refrigerant flow stream from the branch where the auxiliary condenser 9 is located. For the flow stream flowing into the branch where the auxiliary condenser 9 is located, the high-temperature and high-pressure gaseous refrigerant enters the auxiliary condenser 9 through the first four-way reversing valve 3 and the first electric three-way valve 5 to release the condensation heat (secondary heating treatment of the air to be treated), and changes from the gas phase to the liquid phase. After the liquid refrigerant passes through the drying filter 10, it flows to the third expansion valve 11. After throttling by the third expansion valve 11, the pressure and temperature of the refrigerant are greatly reduced, and it is converted into a gas-liquid two-phase state, and then merges with the refrigerant flow stream from the branch where the main condenser 7 is located. The mixed gas-liquid two-phase fluid enters the wet balancer 16 after passing through the second electric three-way valve 12 and the second four-way reversing valve 15, and absorbs heat from the external environment and vaporizes into superheated steam. Subsequently, the refrigerant vapor returns to the second four-way reversing valve 15 and flows to the gas-liquid separator 14 through the third electric three-way valve 17. After the gas-liquid separator 14 acts, it enters the compressor 1 in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. In this way, the basic core function of the system in the pure heating control mode can be realized.

[0098] Furthermore, a performance evaluation is carried out under the following operating conditions of the air-conditioning system: the evaporation pressure of the system is 0.7 MPa and the condensation pressure is 2.2 MPa; the vapor superheat of the refrigerant at the outlet of the wet balancer 16 is 5K; the liquid subcooling of the refrigerant at the outlet of the main condenser 7 and the auxiliary condenser 9 is 2K. According to the mass flow rate, condensing pressure, inlet enthalpy, outlet subcooling and inlet parameters of the air side of the refrigerant in the system, the single tube length of the main condenser 7 is solved to be 0.2039m, 3 branch flow channels, 3 rows of tubes are arranged, 6 single tubes are arranged in each row of tubes, the inner diameter of the tube is 0.009m, the outer diameter of the tube is 0.0095m, the tube spacing is 0.020m, and the row spacing is 0.0217m; the single tube length of the auxiliary condenser 9 is 0.3102m, 3 branch flow channels, 3 rows of tubes are arranged, 6 single tubes are arranged in each row of tubes, the inner diameter of the tube is 0.009m, the outer diameter of the tube is 0.0095m, the tube spacing is 0.020m, and the row spacing is 0.0217m. According to the mass flow rate, evaporation pressure, inlet enthalpy, outlet superheat and inlet parameters of the refrigerant on the air side, it can be solved that the single tube length of the wet balancer is 0.2792m, there are 2 branch flow channels, 5 rows of tubes are set up, and each row of tubes has 6 single tubes, with an inner diameter of 0.009m, an outer diameter of 0.01m, a tube spacing of 0.027m, and a row spacing of 0.022m.

[0099] In specific implementation, the air to be treated first flows through the main condenser 7 and then flows through the auxiliary condenser 9, and is heated in two stages in sequence. The outlet state of the air treated by the main condenser 7 can be used as the inlet condition of the auxiliary condenser 9.

[0100] Through system simulation modeling analysis, under the aforementioned heat exchanger specifications, the air to be treated (inlet parameter temperature 15°C, relative humidity 60%RH, flow rate 2m / s) is treated by the main condenser, and its outlet temperature is 23.25°C and relative humidity 35.9%RH; and then after secondary heating by the auxiliary condenser 9, its outlet temperature rises to 28.64°C and the relative humidity drops to 26.1%RH. From the perspective of the energy efficiency of the air conditioning system, at this time, the refrigerant heat exchange power on the main condenser 7 side and the auxiliary condenser 9 side is 1406.78W, and the system's heating performance coefficient is 6.22.

[0101] In the following specific embodiments 3, 4, and 5, in order to highlight the contrast between the data, the main evaporator 13, auxiliary evaporator 18 and heat balancer 4 that meet the corresponding specifications and sizes in specific embodiment 1 and the main condenser 7, auxiliary condenser 9 and wet balancer 16 that meet the corresponding specifications and sizes in specific embodiment 2 will continue to be used.

[0102] Specific implementation method 3, as Figure 3As shown, the embodiment of the present invention provides an operation mode of a heat and moisture balance type two-stage temperature and humidity control system under the requirement of low temperature and constant humidity control. In the specific implementation, the high-temperature and high-pressure gaseous refrigerant at the outlet of the compressor 1 is divided into two streams after passing through the oil separator 2, and flows into the main condenser 7 and the branch where the heat balancer 4 is located respectively. In this specific implementation, the heat balancer 4 adopts a finned tube heat exchanger. In this specific implementation, the dry mass flow rate of the refrigerant is 0.01389 kg / s, and the refrigerant flow rate of the main condenser 7 branch and the branch where the first expansion valve 6 is located accounts for 70% and 30% of the total refrigeration flow rate of the dry road respectively (i.e. 0.7:0.3). For the stream flowing into the branch where the heat balancer 4 is located, the high-temperature and high-pressure gaseous refrigerant is adjusted by the first four-way reversing valve 3 and enters the heat balancer 4, changing from gas to liquid. The cooled refrigerant liquid returns to the first four-way reversing valve 3 and flows to the first expansion valve 6 through the first electric three-way valve 5. In this specific implementation, the first expansion valve 6 adopts a thermal expansion valve. Under the throttling effect of the first expansion valve 6, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the gas-liquid two-phase refrigerant flow stream from the branch where the main condenser 7 is located. For the flow stream flowing into the branch where the main condenser 7 is located, the high-temperature and high-pressure gaseous refrigerant first releases the condensation heat in the main condenser 7, changes from gas to liquid, and then flows to the second expansion valve 8. In this specific embodiment, the first expansion valve 6 adopts an electronic expansion valve. Under the throttling effect of the second expansion valve 8, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the gas-liquid two-phase refrigerant flow stream from the branch where the heat balancer 4 is located. The mixed gas-liquid two-phase fluid enters the main evaporator 13 through the second electric three-way valve 12, and absorbs the heat in the treated air therein to realize the functions of refrigeration and humidity control (primary treatment of the treated air: refrigeration). The heat-absorbing and gasified refrigerant vapor enters the compressor 1 after passing through the gas-liquid separator 14, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If the cycle works in this way, the basic core function of the system in the low temperature constant humidity control mode can be realized. The main condenser 7 is used to recover part of the refrigerant condensation heat, and is used to increase the temperature and reduce the relative humidity of the air to be treated after it has been cooled and humidified (absolute moisture content is reduced and relative humidity is increased) when passing through the main evaporator 13 (secondary treatment of the air to be treated: heating).

[0103] Furthermore, the performance evaluation is carried out under the following air conditioning system operating conditions: the evaporation pressure of the system is 0.7MPa and the condensation pressure is 2.2MPa. The subcooling degree of the refrigerant stream flowing through the first expansion valve 6 and the refrigerant stream flowing through the second expansion valve 8 before throttling is 2K; the suction superheat degree at the inlet side of the compressor 1 is 5K. It can be obtained that the refrigerant enthalpy value of the refrigerant at the inlet side of the main evaporator 13 is 255.368kJ / kg; the refrigerant enthalpy value of the refrigerant at the inlet side of the main condenser 7 and the heat balancer 4 is 457.944kJ / kg.

[0104] In the specific implementation, the inlet temperature and relative humidity of the air to be treated are 25°C and 40%RH respectively; the air flow rate is 2m / s. The air to be treated first flows through the main evaporator 13, and then flows through the main condenser 7, and is treated in two stages in sequence. Therefore, the air outlet state condition after being treated by the main evaporator 13 can be used as the inlet state condition of the main condenser 7. According to the system simulation modeling analysis, under the aforementioned heat exchanger specifications, after the air to be treated flows through the main evaporator 13, its outlet temperature is 12.63°C and the relative humidity is 77.6%RH; and then after flowing through the main condenser 7 for heat recovery, its outlet temperature can be cooled to 22.28°C and the relative humidity is 42%RH. From the perspective of the energy efficiency of the air conditioning system, at this time, the refrigerant heat exchange on the main evaporator 13 side and the main condenser 7 side is 1184.20W and 1822.50W respectively. The system's refrigeration performance coefficient (COP) is 2.62; the system's heating performance coefficient reaches 4.03.

[0105] Specific implementation method 4, as Figure 4As shown, the embodiment of the present invention provides an operation mode of a heat and moisture balance type two-stage temperature and humidity control system under the demand of medium temperature constant humidity control. In the specific implementation, the high-temperature and high-pressure gaseous refrigerant at the outlet of the compressor 1 is separated by the oil separator 2 and then divided into two streams, which flow into the main condenser 7 (for primary heating treatment of the air to be treated) and the branch where the heat balancer 4 is located. In this specific implementation, the main condenser 7 and the heat balancer 4 both use fin-tube heat exchangers. In this specific implementation, the dry mass flow rate of the refrigerant is 0.01389 kg / s, and the refrigerant flow rates of the main condenser 7 branch and the branch where the first expansion valve 6 is located account for 70% and 30% of the total dry refrigeration flow rate, respectively (i.e., 0.7:0.3). For the stream flowing into the branch where the heat balancer 4 is located, the high-temperature and high-pressure gaseous refrigerant is adjusted backward by the first four-way reversing valve 3, enters the heat balancer 4, and releases part of the condensation heat to the external environment. After completing the heat unloading, the refrigerant returns to the first four-way reversing valve 3, enters the auxiliary condenser 9 through the first electric three-way valve 5, and releases the condensation heat into the air duct (secondary heating treatment of the air to be treated). The refrigerant condensed into a liquid state (subcooling degree 2K) flows to the third expansion valve 11 after passing through the drying filter 10. After throttling by the third expansion valve 11, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the gas-liquid two-phase refrigerant split stream from the branch where the main condenser 7 is located. For the stream flowing into the branch where the main condenser 7 is located, the high-temperature and high-pressure gaseous refrigerant first enters the main condenser 7 to release the condensation heat, and changes from gas to liquid (subcooling degree 2K), and then flows to the second expansion valve 8. After throttling by the second expansion valve 8, the pressure and temperature of the refrigerant are greatly reduced, and then merge with the gas-liquid two-phase refrigerant split stream from the branch where the heat balancer 4 is located. The mixed gas-liquid two-phase fluid enters the main evaporator 13 through the second electric three-way valve 12, and absorbs the heat in the air to be treated to achieve cooling and debugging (primary refrigeration treatment of the air to be treated). After that, it passes through the gas-liquid separator 14 and enters the compressor 1 in a completely gas phase state (superheat 5K), and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core function of the system in the medium-temperature constant humidity control mode can be realized.

[0106] Furthermore, the performance evaluation is carried out under the following operating conditions of the air-conditioning system: the evaporation pressure of the system is 0.7MPa and the condensation pressure is 2.2MPa. The subcooling degree of the refrigerant stream flowing through the second expansion valve 8 and the refrigerant stream flowing through the third expansion valve 11 before throttling is 2K; the suction superheat degree at the inlet side of the compressor 1 is 5K. It can be obtained that the refrigerant enthalpy value of the refrigerant at the inlet side of the main evaporator 13 is 255.368kJ / kg; the refrigerant enthalpy value of the refrigerant at the inlet side of the main condenser 7 and the heat balancer 4 is 457.944kJ / kg. At this time, the refrigerant enthalpy value at the inlet side of the auxiliary condenser 9 is affected by the heat exchange intensity on the heat balancer 4 side.

[0107] In the specific implementation, the inlet temperature and relative humidity of the air to be treated are 25°C and 40%RH respectively; the air flow rate is 2m / s. The air to be treated first flows through the main evaporator 13, and then flows through the main condenser 7 and the auxiliary condenser 9, and is treated in two stages of cooling and heating in sequence. Therefore, the outlet state of the air treated by the main evaporator 13 can be used as the inlet condition of the main condenser 7, and the outlet state of the air treated by the main condenser 7 can be used as the inlet condition of the auxiliary condenser 9. According to the system simulation modeling analysis, under the aforementioned heat exchanger specifications, after the air to be treated flows through the main evaporator 13, its outlet temperature is 12.63°C and the relative humidity is 77.6%RH; after flowing through the main condenser 7 for heat recovery, its outlet temperature is 22.28°C and the relative humidity is 42%RH. Then, the specific enthalpy of the refrigerant at the inlet of the auxiliary condenser 9 is 420.0 kJ / kg (the heat balancer 4 balances 37.944 kJ / kg of heat at this time). After the refrigerant flows out of the auxiliary condenser 9, the outlet temperature of the treated air can be restored to 25.04°C, and the relative humidity can be reduced to 35.7% RH. From the perspective of the energy efficiency of the air conditioning system, at this time, the refrigerant heat exchange on the main evaporator 13 side, the main condenser 7 side and the auxiliary condenser 9 side is 1184.20W, 1822.50W, and 763.85W respectively, and the system refrigeration performance coefficient is 2.62; the system heating performance coefficient is 5.72. If it is necessary to further increase the temperature of the output air and reduce the relative humidity, it can be quickly achieved by reducing the heat release intensity of the heat balancer 4. At this time, the inlet specific enthalpy of the refrigerant on the auxiliary condenser 9 side can be increased to 457.944 kJ / kg. In this specific implementation, after reheating by the auxiliary condenser 9, the outlet temperature of the processed air is 25.60°C and the relative humidity is 34.5%RH. At this time, compared with the specific implementation 3, the outlet temperature of the processed air is increased by 3.32°C. From the perspective of the energy efficiency of the air conditioning system, the refrigerant heat exchange on the main evaporator 13 side, the main condenser 7 side and the auxiliary condenser 9 side is 1184.2W, 1822.5W, and 918.3W respectively, the system refrigeration performance coefficient is 2.62, and the system heating performance coefficient is 6.06.

[0108] Specific implementation method 5, as Figure 5As shown, the embodiment of the present invention provides an operation mode of a heat and moisture balance type two-stage temperature and humidity control system under the requirement of high temperature and constant humidity control. In the specific implementation, the high temperature and high pressure gaseous refrigerant at the outlet of the compressor 1 is divided into two streams after passing through the oil separator 2, and flows into the branch where the main condenser 7 and the auxiliary condenser 9 are located respectively. In this specific implementation, the main condenser 7 and the auxiliary condenser 9 are both finned tube heat exchangers. The air to be treated is heated at the first level after flowing through the main condenser 7, and is heated at the second level after flowing through the auxiliary condenser 9. In this specific implementation, the refrigerant flow of the main condenser 7 branch and the auxiliary condenser 9 branch account for 70% and 30% of the total refrigeration flow of the trunk line respectively (i.e. 0.7:0.3). For the stream flowing into the branch where the main condenser 7 is located, the high temperature and high pressure gaseous refrigerant condenses and releases heat therein, and flows to the second expansion valve 8 after changing from gas to liquid. After throttling by the second expansion valve 8, the pressure and temperature of the refrigerant are greatly reduced, and then merged with the refrigerant branch stream from the branch where the auxiliary condenser 9 is located. For the flow stream flowing into the branch where the auxiliary condenser 9 is located, the high-temperature and high-pressure gaseous refrigerant enters the auxiliary condenser 9 through the first four-way reversing valve 3 and the first electric three-way valve 5 to condense and release heat, and changes from gas to liquid. After the liquid refrigerant passes through the drying filter 10, it flows to the third expansion valve 11. After throttling by the third expansion valve 11, the pressure and temperature of the refrigerant are greatly reduced, and it is transformed into a gas-liquid two-phase state (mainly liquid phase). Then it merges with the refrigerant diversion stream from the branch where the main condenser 7 is located. In this specific embodiment, the dry mass flow rate of the refrigerant is 0.019446kg / s. The mixed gas-liquid two-phase fluid flows into the branches where the main evaporator 13 and the wet balancer 16 are located respectively through the second electric three-way valve 12 in proportion (71.4%: 28.6%). For the stream flowing into the branch where the main evaporator 13 is located, the low-temperature and low-pressure refrigerant enters the main evaporator 13, and absorbs the heat in the air to be treated therein to achieve a first-level cooling and debugging treatment (a first-level refrigeration treatment of the air to be treated). The vapor leaving the main evaporator 13 merges with the refrigerant vapor branch from the branch where the wet balancer 16 is located. This is the confluence point of the two refrigerant vapors. After the stream flowing into the branch where the wet balancer 16 is located is adjusted by the second four-way reversing valve 15, the low-temperature and low-pressure refrigerant enters the wet balancer 16, absorbs heat from the external environment and vaporizes into superheated vapor, and then passes through the second four-way reversing valve 15 and the third electric three-way valve 17 in turn, and merges with the refrigerant vapor branch from the branch where the main evaporator 13 is located. The merged refrigerant vapor passes through the gas-liquid separator 14 and enters the compressor 1 in a completely gaseous state, and is then compressed again into a high-temperature and high-pressure gaseous refrigerant. If it works in a cycle, the basic core function of the system under the high-temperature constant humidity control mode can be realized.

[0109] Furthermore, the performance evaluation is carried out under the following air conditioning system operating conditions: the evaporation pressure of the system is 0.7MPa and the condensation pressure is 2.2MPa. The subcooling degree of the refrigerant stream flowing through the second expansion valve 8 and the refrigerant stream flowing through the third expansion valve 11 before throttling is 2K; the suction superheat degree at the inlet side of the compressor 1 is 5K. It can be obtained that the refrigerant enthalpy values ​​of the refrigerant at the inlet side of the main evaporator 13 and the wet balancer 16 are both 255.368kJ / kg; the refrigerant enthalpy values ​​of the refrigerant at the inlet side of the main condenser 7 and the auxiliary condenser 9 are both 457.944kJ / kg.

[0110] In the specific implementation, the inlet temperature and relative humidity of the air to be treated are 25°C and 40%RH respectively; the air flow rate is 2m / s. The air to be treated first flows through the main evaporator 13, then flows through the main condenser 7 and the auxiliary condenser 9, and is treated in two stages of cooling and heating in sequence. Therefore, the outlet state of the air treated by the main evaporator 13 can be used as the inlet condition of the main condenser 7, and the outlet state of the air treated by the main condenser 7 can be used as the inlet condition of the auxiliary condenser 9. According to the system simulation modeling analysis, under the aforementioned heat exchanger specifications, after the air to be treated flows through the main evaporator 13, its outlet temperature is 12.63°C and the relative humidity is 77.6%RH; after flowing through the main condenser 7 for heat recovery, its outlet temperature is 22.62°C and the relative humidity is 41.3%RH. After reheating in the auxiliary condenser 9, the outlet of the treated air rises to 27.17°C, and the relative humidity is reduced to 31.5%RH. At this time, compared with the specific embodiment 4, the outlet temperature of the processed air is further increased by 1.57°C. From the perspective of the energy efficiency of the air conditioning system, at this time, the cooling capacity of the main evaporator 13 side is 1184.20W, which is the same as that in the specific embodiment 4, while the heat exchange of the main condenser 7 and the auxiliary condenser 9 has increased, reaching 1903.70W and 1258.40W respectively. However, because the refrigeration cycle capacity has increased by 40%, the compressor power has increased by 180.88W, so the system refrigeration performance coefficient has decreased to 1.79; the system heating performance coefficient is 4.99.

[0111] The above description is only a preferred specific implementation manner of the present invention, but the protection scope of the present invention is not limited thereto. Any technician familiar with the technical field can make equivalent replacements or changes according to the technical solutions and concepts of the present invention within the technical scope disclosed by the present invention, which should be covered by the protection scope of the present invention.

Claims

1. A heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability, characterized in that: It includes a refrigeration and air-conditioning system, an air handling unit, a heat and moisture balance system, and a temperature fine-tuning system based on heat and cold recovery. The refrigeration and air-conditioning system provides the required cooling or heat to the air handling unit, the air handling unit performs temperature and humidity treatment on the treated air, the heat and moisture balance system regulates the outlet temperature and humidity of the treated air, and the temperature fine-tuning system based on heat and cold recovery is arranged at the end of the air handling unit to adjust the outlet temperature of the treated air.

2. A heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 1, characterized in that: The refrigeration and air-conditioning system comprises a compressor (1), an oil separator (2), a first four-way reversing valve (3), a heat balancer (4), a first electric three-way valve (5), a first expansion valve (6), a main condenser (7), a second expansion valve (8), an auxiliary condenser (9), a drying filter (10), a third expansion valve (11), a second electric three-way valve (12), a main evaporator (13), a gas-liquid separator (14), a second four-way reversing valve (15), a wet balancer (16), a third electric three-way valve (17), an auxiliary evaporator (18), an on-off valve (19), an air supply regulator (21), a primary air filter (22), a medium-efficiency air filter (23), and a fan (24); wherein the compressor ( The output end of the oil separator (1) is connected to the input end of the oil separator (2) through a pipeline, and the output end of the oil separator (2) is divided into two branches, one of which is connected to the main condenser (7) and the other is connected to the first four-way reversing valve (3), which is a branch point 1; for the pipeline where the first four-way reversing valve (3) is located, the output end of the first four-way reversing valve (3) is connected to the input end of the heat balancer (4), the output end of the heat balancer (4) and the first electric three-way valve (5); by adjusting the first four-way reversing valve (3), the system can control whether to unload the heat in the high-temperature and high-pressure refrigerant vapor; then, the refrigerant flows into the two branches in a certain proportion through the first electric three-way valve (5), and one branch is connected to the input end of the heat balancer (4), the output end of the heat balancer (4) and the first electric three-way valve (5). The first expansion valve (6) is connected to the second expansion valve (6), and the other branch is connected to the auxiliary condenser (9), which is the branch point 2; for the pipeline where the auxiliary condenser (9) is located, the output end of the auxiliary condenser (9) is connected to the drying filter (10), and the output end of the drying filter (10) is connected to the third expansion valve (11); the output end of the third expansion valve (11) is connected to the fourth one-way stop valve, and then merges with the pipeline where the first expansion valve (6) is located, which is the confluence point 2; for the pipeline where the first expansion valve (6) is located, the other end of the first expansion valve (6) is connected to the third one-way stop valve, and then merges with the pipeline where the auxiliary condenser (9) is located, which is the confluence point 2; after the two pipelines are connected, they are connected to the main condenser (7) through the second one-way stop valve. The pipelines where the two phases of the main condenser (7) are located are merged, which is the confluence point 1; for the pipeline where the main condenser (7) is located, when the switch valve (19) is in the open state, the high-temperature and high-pressure refrigerant enters the main condenser (7) and releases heat to the air in the air duct, thereby increasing the temperature of the treated air; the other end of the main condenser (7) is connected to the second expansion valve (8), and then merges with the pipeline where the second one-way stop valve is located through the first one-way stop valve, which is the confluence point 1; the two-phase refrigerant in each section of the pipeline merges and flows to the second electric three-way valve (12); the output end of the second electric three-way valve (12) is divided into two branches, one branch is connected to the main evaporator (13), and the other branch is connected to the second four-way reversing valve (15), which is the diversion point 3;For the pipeline where the second four-way reversing valve (15) is located, the output end of the second four-way reversing valve (15) is connected to the input end of the wet balancer (16), the output end of the wet balancer (16) and the third electric three-way valve (17); by adjusting the second four-way reversing valve (15), the system can control whether to unload the cold capacity in the low-temperature and low-pressure refrigerant; then, the refrigerant flows into two branches in a certain proportion through the third electric three-way valve (17), one branch is connected to the sixth one-way stop valve, and the other branch is connected to the auxiliary evaporator (18), which is the diversion point 4; for the pipeline where the auxiliary evaporator (18) is located, the output end of the auxiliary evaporator (18) is connected to the seventh one-way stop valve, and then merges with the pipeline where the sixth one-way stop valve is located, which is the confluence point 4; for the pipeline where the sixth one-way stop valve is located , the other end of the six one-way stop valves merges with the pipeline where the seventh one-way stop valve is located, which is the confluence point 4; after the two pipelines are connected, they merge with the pipeline where the main evaporator (13) is located, which is the confluence point 3; for the pipeline where the main evaporator (13) is located, the other end of the main evaporator (13) merges with the common pipeline where the pipeline where the sixth one-way stop valve is located and the pipeline where the seventh one-way stop valve is located through the fifth one-way stop valve, which is the confluence point 3; after the superheated refrigerant in each section of the pipeline is merged, it passes through the gas-liquid separator (14) and returns to the compressor (1) in a completely gaseous state; the air supply regulator (21), the primary air filter (22), the medium-efficiency air filter (23), the main evaporator (13), the main condenser (7), the auxiliary evaporator (18), the auxiliary condenser (9), and the fan (24) are arranged on the air duct. ; 3. A heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 2, characterized in that: A dehumidifier, a temperature increaser, a recooler and a reheater constitute an air handling unit of a refrigeration and air conditioning system, wherein the dehumidifier corresponds to a main evaporator (13) in the refrigeration and air conditioning system, the temperature increaser corresponds to a main condenser (7) in the refrigeration and air conditioning system, the recooler corresponds to an auxiliary evaporator (18) in the system refrigeration and air conditioning system, and the reheater corresponds to an auxiliary condenser (9) in the system refrigeration and air conditioning system.

4. A heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 3, characterized in that: When pure cooling is required, the dehumidifier has a low-temperature and low-pressure refrigerant flowing in to cool the air passing through; when constant humidity is required, low-temperature and low-pressure refrigerant flows in to cool and dehumidify the air passing through; when pure heating is required, the dehumidifier does not work.

5. The heat and moisture balance type two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 3, characterized in that: The temperature booster does not function when pure cooling is required; when low-temperature constant humidity is required, high-temperature and high-pressure refrigerant flows in to achieve heating and heating of the processed air; when medium-temperature constant humidity, high-temperature constant humidity and pure heating are required, high-temperature and high-pressure refrigerant flows in to heat the processed air once.

6. A heat and moisture balance type two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 3, characterized in that: When the control requirement of pure refrigeration is met, the subcooler has a low-temperature and low-pressure refrigerant flowing in, which is used to perform secondary cooling on the air flowing through it; when the control requirement of low-temperature constant humidity is met, the low-load and low-flow refrigerant unloaded by the moisture balancer (16) can be introduced according to actual needs to accurately adjust the outlet temperature of the processed air; it does not take effect when the control requirements of medium-temperature constant humidity, high-temperature constant humidity and pure heating are met.

7. A heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 3, characterized in that: When the control requirements are for pure refrigeration and low-temperature constant humidity, the reheater can introduce low-load and low-flow refrigerant after unloading by the heat balancer (4) according to actual needs, and accurately adjust the outlet temperature of the processed air; when the control requirements are for medium-temperature constant humidity, high-temperature constant humidity and pure heating, it is used to perform secondary heating on the processed air.

8. The heat and moisture balance two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 3, characterized in that: The heat balancer (4) and the moisture balancer (16) constitute a heat and moisture balance system of a refrigeration and air conditioning system; the heat balancer (4) unloads the heat of the main condenser (7) and the auxiliary condenser (9) and releases it to the outside air through the action of the first four-way reversing valve (3), and can control the flow ratio of the refrigerant in the heat balancer and its heat exchange intensity by adjusting the opening size of the first expansion valve (6), the second expansion valve (8) and the third expansion valve (11); the moisture balancer (16) unloads the cold of the main evaporator (13) and the auxiliary evaporator (18) and releases it to the outside air through the action of the second four-way reversing valve (15), and can control the flow ratio of the refrigerant in the moisture balancer and its heat exchange intensity by adjusting the opening size of the second electric three-way valve (12) and the third electric three-way valve (17).

9. A heat and moisture balance type two-stage temperature and humidity control air conditioning system with condensation heat recovery capability according to claim 8, characterized in that: The auxiliary condenser (9) and the auxiliary evaporator (18) constitute a temperature fine-tuning system based on heat recovery of a refrigeration and air-conditioning system; the auxiliary condenser (9) is controlled by a first four-way reversing valve (3) to be connected in series with the heat balancer (4) to realize the recovery and utilization of the heat of the refrigerant on the heat balancer (4) side; when it is connected in series with the heat balancer (4), due to the heat unloading effect of the heat balancer (4), a small flow rate and low load high-pressure refrigerant flows in it, and by adjusting the first electric three-way valve (5), it can be more accurately The auxiliary evaporator (18) is connected in series with the wet balancer (16) to control the heat input into the air duct; the auxiliary evaporator (18) is controlled by the second four-way reversing valve (15) to realize the recovery and utilization of the cold amount of the refrigerant on the wet balancer (16) side; when it is connected in series with the wet balancer (16), due to the cold amount unloading effect of the wet balancer (16), a low-pressure refrigerant with a small flow rate and low load flows in it, and the cold amount input into the air duct by the auxiliary evaporator can be controlled more accurately by adjusting the third electric three-way valve (17).

10. A control method for a heat and moisture balance type two-stage temperature and humidity control air conditioning system with condensation heat recovery capability as claimed in any one of claims 1 to 9, characterized in that: The air conditioning system switches operating modes, namely: pure cooling control mode, low temperature constant humidity control mode, medium temperature constant humidity control mode, high temperature constant humidity control mode and pure heating control mode.

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