Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B49/00—Arrangement or mounting of control or safety devices
  • F25B49/02—Arrangement or mounting of control or safety devices for compression type machines, plants or systems
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
  • F25B2341/06—Details of flow restrictors or expansion valves
  • F25B2341/063—Feed forward expansion valves
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B2600/00—Control issues
  • F25B2600/25—Control of valves
  • F25B2600/2513—Expansion valves
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
  • F25B2700/19—Pressures
  • F25B2700/193—Pressures of the compressor
  • F25B2700/1931—Discharge pressures
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
  • F25B2700/19—Pressures
  • F25B2700/193—Pressures of the compressor
  • F25B2700/1933—Suction pressures
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
  • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
  • F25B2700/00—Sensing or detecting of parameters; Sensors therefor
  • F25B2700/21—Temperatures
  • F25B2700/2115—Temperatures of a compressor or the drive means therefor
  • F25B2700/21151—Temperatures of a compressor or the drive means therefor at the suction side of the compressor

Definitions

• the invention relates to a heat pump system as well as a method for operating a heat pump system.
• Heat pump systems with a simple vapor compression cycle with a compressor, condenser, expansion valve and evaporator and where the refrigerant used is a hydrofluorocarbon (HFC) like R134A or R410A or azeotropes like R513A are known in the art.
• HFC hydrofluorocarbon
• DE 102020115270 A1 discloses a method and a device for controlling a compression refrigeration machine with a refrigerant, an evaporator, a pressure increasing unit, a condenser and a throttle element with the steps: determining a first control value for the throttle element depending on the deviation of an actual superheat of the refrigerant from a target superheat; determining condenser pressure; determining a speed of the pressure increasing unit; measuring evaporator pressure; forming a model comparing the refrigerant mass flow at the evaporator inlet with the refrigerant mass flow at the evaporator outlet; calculating a second control value for the throttle element based on the model; determining a third control value for the throttle element by linking the first control value with the second control value and setting the throttle element to the third control value.
• the second control value for the throttle element is calculated based on the model from the evaporator pressure, the condenser pressure and the speed of the pressure increasing unit.
• US 11137164 B2 discloses a heat pump system including a first unit having a first unit heat exchanger, a compressor, an accumulator, and a first unit expansion valve, a second unit fluidly connected to the first unit by piping, the second unit having a second unit heat exchanger, and a system controller.
• the system controller has a PID control element receiving as inputs gain scheduling, an error signal, and feedback relating to an opening command of the first unit expansion valve, and a feedforward control element generating a feedforward term that is combined with an output of the PID control element to generate the opening command of the first unit expansion valve.
• the system controller controls an opening of the first unit expansion valve using the opening command of the first unit expansion valve.
• the object of the invention is to provide a robust heat pump system.
• a further object is to specify a method for operating such a robust heat pump system.
• a heat pump system providing a cooling mode and a heating mode
• a refrigeration cycle comprising a refrigeration cycle, the refrigeration cycle in succession at least comprising: a first heat exchanger, a compressor, a second heat exchanger, and a first expansion valve.
• the compressor is configured to compress a refrigerant circulating in the refrigeration cycle.
• the second heat exchanger is configured to transfer thermal energy from the compressed refrigerant to a fluid circuit to be heated.
• the first expansion valve is configured to reduce a pressure of the compressed refrigerant passing through the first expansion valve, and the first heat exchanger is configured to evaporate the refrigerant by transferring thermal energy from a heat transfer medium to the refrigerant.
• the first expansion valve is connected to a controller module at least comprising a feedback controller and a feedforward controller.
• the proposed heat pump system may be used with propane as the refrigerant.
• propane When compared with hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs), propane has a lower system pressure drop and a higher heat transfer performance and a very low global warming potential. But propane is an inflammable medium. Because of these factors, the refrigerant charge for propane is normally 40-60% less than other common refrigerants which in turn poses a tremendous challenge for the control of the superheat by regulating the mass flow rate of propane to the evaporator, especially in the absence of a liquid receiver.
• the heat pump system comprises a PI (proportional-integral) controller as a gain-scheduled feedback controller plus a feedforward superheat controller that can adapt to a wide range of operating conditions to provide tight superheat control at a good performance and speed of response during steady and transient operation comprising system load changes of a variable frequency heat pump system.
• PI proportional-integral
• the control parameters can be automatically adjusted in real-time in accordance with the change of operating condition.
• Three scheduling variables, the evaporation pressure, the condensation pressure and the compressor frequency may be used to vary the parameters of the PI feedback controller because a variable frequency air-to-water heat pump has a very wide operating envelope especially in heating mode, for example, ambient temperature as air intake temperature ranging from -22°C to 45°C, water outlet temperatures ranging from 25°C to 70°C and compressor frequencies ranging from 20 Hz to 100Hz.
• ambient temperature as air intake temperature ranging from -22°C to 45°C
• water outlet temperatures ranging from 25°C to 70°C
• compressor frequencies ranging from 20 Hz to 100Hz.
• the setpoint can be shifted closer to the operating constraint (0 K), thereby decreasing electrical power consumption of the compressor and hence increasing a coefficient of performance, also known as COP.
• Drastic changes of the compressor frequency during system load changes may cause severe superheat regulation issues including oscillations and even wet compression.
• PI controllers are effective in rejecting disturbances with slow variations, for example external disturbances like ambient air temperature fluctuations.
• external disturbances like ambient air temperature fluctuations.
• internal disturbances such as changes of compressor frequency can have an immediate effect on the superheat and therefore cannot be rejected quickly by feedback controllers.
• PI controllers can take corrective actions only after the superheat as a process output has drifted from a setpoint.
• a compensating control actuation can be applied immediately to the expansion valve to reject the disturbance effect associated with compressor frequency changes under a wide range of operating conditions thereby preventing the superheat being drifted too far from its setpoint.
• the feedforward gain is calculated from measurable disturbance (ambient air temperature) on the fly.
• the performance gains include fast post-disturbance recovery of the superheat with shortened settling time, complete avoidance of wet compression and enhanced transient energy efficiency.
• the proposed heat pump system as a variable frequency air-to-water heat pump system comprises a robust control system that regulates the superheat at the evaporator outlet of the first air heat exchanger effectively to the setpoint in the heating mode over the entire operating envelope, both during steady operation and system load changes.
• a temperature sensor and a pressure sensor may be provided at an outlet of the first heat exchanger for measuring a refrigerant temperature and an evaporation pressure of a superheated vapor refrigerant.
• the refrigerant temperature downstream the first heat exchanger and the evaporation pressure of the superheated vapor refrigerant measured by both sensors may serve as an input to the feedback controller.
• the feedback controller may comprise a gain scheduled linear feedback controller at least based on an input of the refrigerant temperature and the evaporation pressure of the superheated vapor refrigerant at the outlet of the first heat exchanger.
• the feedforward controller may comprise a gain scheduled dynamic feedforward controller at least based on an input of a compressor frequency.
• I/O data may be collected through identification experiments at several properly selected operating points distributed throughout the operating envelope of the heat pump.
• An input may be an opening degree of the first expansion valve and an output may be the superheat at the outlet of the first heat exchanger.
• the recorded I/O data may be utilized to identify linearized mathematical models in the form of a transfer function for the behavior of the superheat in response to changes in the opening degree of the first expansion valve. So there may be one model for each operating point obtained, using the system identification technique.
• control parameters as e.g. a proportional gain and an integral time, for each operating point using e.g. a direct synthesis model-based approach.
• control parameters e.g. a proportional gain and an integral time
• the resulting family of linear controllers may be finally implemented as a single adaptive nonlinear controller whose parameters are changed in real-time by monitoring the scheduling variables evaporation pressure, condensation pressure and compressor frequency, thus providing the gain scheduling.
• I/O data may be collected through identification experiments at several properly selected operating points distributed throughout the operating envelope of the heat pump.
• An input may be the compressor frequency
• an output may be the superheat at the outlet of the first heat exchanger.
• the recorded I/O data may be utilized to identify linearized mathematical models in the form of a transfer function for the behavior of the superheat in response to changes in the compressor frequency. Thus there may be one model for each operating point.
• the feedforward gain may be changed in the real-time using the air intake temperature as the scheduling variable, thus providing the gain scheduling.
• the controller module may be configured to combine outputs of the feedback controller and the feedforward controller.
• the controller module may be configured to add and/or multiply and/or weight outputs of the feedback controller and the feedforward controller. Favourably, any unwanted high pressure or low-pressure alarms/faults, thus leading to the shutdown of the heat pump, during starting of the heat pump or during transition of the operating modes of the heat pump may be avoided.
• the controller module may be configured to use the feedback controller as a leading controller. Favourably, any unwanted high pressure or low-pressure alarms/faults, thus leading to the shutdown of the heat pump, during starting of the heat pump or during transition of the operating modes of the heat pump may be avoided.
• the refrigeration cycle further may comprise an internal heat exchanger, having a first conduit and a second conduit, the first conduit being in heat exchanging contact with the second conduit.
• the first conduit is part of the fluid line between the second heat exchanger and the first expansion valve and the second conduit is part of the fluid line between the accumulator and the compressor.
• the heat pump system further may comprise a fluid bypass line parallel to the first conduit of the internal heat exchanger.
• the internal heat exchanger may favourably be used for further superheating the refrigerant before entering the compressor.
• a control valve may be provided in the refrigeration cycle in series with the internal heat exchanger.
• the control valve may be provided in the bypass line.
• the controlled variable is the suction superheat, i.e. the superheat at the inlet of the compressor.
• the control valve may advantageously be a ball valve.
• the control valve may be configured to be controlled by a multi-step control with discrete openings, or by a feedback controller.
• a coupling or an interaction between the control valve and the first expansion valve may be reduced.
• the control valve control loop may determine an optimized (multi-step discrete) opening of 10%, 32% ... upfront to keep the suction superheat i.e. the superheat at the inlet of the compressor, within the range which ensures the safety and efficient operation of the compressor. This range can be obtained from the manufacturer's datasheets. For the particular advantageous embodiment the range is 5 K to 30 K.
• the superheat control loop of the first expansion valve then regulates the first expansion valve according to the actual opening of the control valve to enable an optimized system setup for the current operating point.
• a discharge pressure sensor may be provided in the fluid line between the compressor and the second heat exchanger for measuring a discharge pressure.
• a current value of the condensation pressure may be obtained from the measured discharge pressure.
• a second expansion valve may be provided in the fluid line between the second heat exchanger and the internal heat exchanger.
• the direction of refrigerant flow is reversed for proper operation of the heat pump system.
• a method for operating a heat pump system comprising a refrigeration cycle, the refrigeration cycle in succession at least comprising: a first heat exchanger, a compressor, a second heat exchanger, and a first expansion valve.
• a heating mode the compressor compresses a refrigerant circulating in the refrigeration cycle.
• the second heat exchanger transfers thermal energy from the compressed refrigerant to a fluid circuit to be heated.
• the first expansion valve is configured to reduce a pressure of the compressed refrigerant passing through the first expansion valve, and the first heat exchanger evaporates the refrigerant by transferring thermal energy from a heat transfer medium to the refrigerant.
• An opening degree of the first expansion valve is controlled by a controller module at least comprising a feedback controller and a feedforward controller.
• the proposed method may be used for operating a heat pump system with propane as the refrigerant.
• propane When compared with hydrochlorofluorocarbons (HCFCs) and hydrofluorocarbons (HFCs), propane has a lower system pressure drop and a higher heat transfer performance and a very low global warming potential. But propane is an inflammable medium. Because of these factors, the refrigerant charge for propane is normally 40-60% less than other common refrigerants which in turn poses a tremendous challenge for the control of the superheat by regulating the mass flow rate of propane to the evaporator, especially in the absence of a liquid receiver.
• the heat pump system comprises a PI (proportional-integral) controller as a gain-scheduled feedback controller plus a feedforward superheat controller that can adapt to a wide range of operating conditions to provide tight superheat control at a good performance and speed of response during steady and transient operation comprising system load changes of a variable frequency heat pump system.
• PI proportional-integral
• control parameters can be automatically adjusted in real-time in accordance with the change of operating condition.
• Three scheduling variables may be used to vary the parameters of the PI feedback controller because a variable frequency air-to-water heat pump has a very wide operating envelope especially in heating mode, for example, air intake temperature ranging from, e.g., -22°C to 45°C, water outlet temperatures ranging from 25°C to 70°C and compressor frequencies ranging from 20 Hz to 100Hz.
• air intake temperature ranging from, e.g., -22°C to 45°C
• water outlet temperatures ranging from 25°C to 70°C
• compressor frequencies ranging from 20 Hz to 100Hz.
• Drastic changes of the compressor frequency during system load changes may cause severe superheat regulation issues including oscillations and even wet compression.
• PI controllers are effective in rejecting disturbances with slow variations, for example external disturbances like ambient air temperature fluctuations.
• external disturbances like ambient air temperature fluctuations.
• internal disturbances such as changes of compressor frequency can have an immediate effect on the superheat and therefore cannot be rejected quickly by feedback controllers.
• PI controllers can take corrective actions only after the superheat as a process output has drifted from a setpoint.
• a compensating control actuation can be applied immediately to the expansion valve to reject the disturbance effect associated with compressor frequency changes under a wide range of operating conditions. This allows for preventing the superheat being drifted too far from its setpoint.
• the first expansion valve may use the real-time compressor frequency as input.
• the feedforward gain is calculated from measurable disturbance (ambient air temperature) on the fly.
• the performance gains include fast post-disturbance recovery of the superheat with shortened settling time, complete avoidance of wet compression and enhanced transient energy efficiency.
• the proposed method for operating a heat pump system as a variable frequency air-to-water heat pump system offers a robust control method that regulates the superheat at the evaporator outlet of the first air heat exchanger effectively to the setpoint in the heating mode over the entire operating envelope, both during steady operation and system load changes.
• a gain scheduled linear feedback controller may be used as the feedback controller.
• a gain scheduled dynamic feedforward controller may be used as the feedforward controller.
• I/O data may be collected through identification experiments at several properly selected operating points distributed throughout the operating envelope of the heat pump.
• An input may be an opening degree of the first expansion valve and an output may be the superheat.
• the recorded I/O data may be utilized to identify linearized mathematical models in the form of a transfer function for the behavior of the superheat in response to changes in the opening degree of the first expansion valve. So there may be one model for each operating point, using the system identification technique.
• control parameters may then be used to obtain the control parameters, as e.g. a proportional gain and an integral time, for each operating point using e.g. a direct synthesis model-based approach.
• control parameters may then be used to obtain the control parameters, as e.g. a proportional gain and an integral time, for each operating point using e.g. a direct synthesis model-based approach.
• models based on artificial intelligence may beneficially be used for determining control parameters.
• Neural networks based systems as well as other machine learning algorithms may be implemented for optimizing processes of obtaining control parameters.
• the resulting family of linear controllers may be finally implemented as a single adaptive nonlinear controller whose parameters are changed in real-time by monitoring the scheduling variables evaporation pressure, condensation pressure and compressor frequency, thus providing the gain scheduling.
• I/O data may be collected through identification experiments at several properly selected operating points distributed throughout the operating envelope of the heat pump.
• An input may be the compressor frequency
• an output may be the superheat.
• the recorded I/O data may be utilized to identify linearized mathematical models in the form of a transfer function for the behavior of the superheat in response to changes in the compressor frequency. Thus there may be one model for each operating point.
• the feedforward gain may be changed in the real-time using the air intake temperature as the scheduling variable, thus providing the gain scheduling.
• an output of the feedback controller may be determined comprising determining the refrigerant temperature and the evaporation pressure of the superheated vapor refrigerant by a temperature sensor and a pressure sensor at the outlet of the first heat exchanger; determining a compressor frequency by an inverter driving the compressor; determining a saturation temperature using a refrigerant property table; determining a superheat by subtracting the saturation temperature from the temperature of the superheated vapor refrigerant; determining a control error by comparing the calculated superheat with the superheat setpoint as an input to the feedback controller; determining a condensation pressure from a discharge pressure by a pressure sensor; determining a proportional gain and an integral time by a parameter adjustment mechanism using the evaporation pressure, the condensation pressure and the compressor frequency, the proportional gain and the integral time being used as an input to the feedback controller.
• the gain scheduling functions may be implemented as ten-coefficient quadratic polynomials of the evaporation pressure, the condensation pressure and the compressor frequency.
• the saturation temperature corresponds to the evaporation temperature.
• the refrigerant temperature may be different from an evaporation temperature or saturation temperature that is a tabular value.
• the pressure sensor measures an evaporation pressure of a superheated vapor refrigerant. Calculation of the evaporation temperature using a refrigerant property table thus may be based on the evaporation pressure.
• the proportional gain and the integral time may be limited.
• high speed fluctuations of the proportional gain and the integral time may be dampened by a saturation function block as input to the feedback controller.
• a saturation function block as input to the feedback controller.
• an output of wrong values may be avoided.
• High speed fluctuations may be suppressed, thus avoiding unwanted wear in an actuator of the first expansion valve.
• control error may be determined in a control error calculation block as an input to the feedback controller.
• control error caused by a deviation of the calculated superheat from the superheat setpoint may be corrected.
• an output of the feedforward controller may be determined comprising determining an air intake temperature by an air intake temperature sensor; determining a feedforward gain by a parameter adjustment mechanism, the feedforward gain being used as an input to the feedback controller.
• the parameter adjustment mechanism may be implemented as a linear polynomial of the air intake temperature.
• control parameters of the feedforward controller may be determined.
• a value of the feedforward gain may be limited.
• high speed fluctuations of the feedforward gain may be dampened by a saturation function block as input to the feedforward controller.
• a saturation function block as input to the feedforward controller.
• the compressor frequency minus a base frequency may be used as an input to the feedforward controller.
• the determined value may be used for calculating the output of the transfer function of the feedforward controller.
• the outputs of the feedback controller and the feedforward controller may be combined in a combine function.
• the outputs of the feedback controller and the feedforward controller may be combined in a combine function by adding and/or multiplying and/or weighting both outputs.
• the opening degree of the first expansion valve may be determined by adapting an output of the combine function in a saturation function to control range limits for the opening degree. Favourably, any unwanted high pressure or low-pressure alarms/faults, thus leading to the shutdown of the heat pump, during starting of the heat pump or during transition of the operating modes of the heat pump may be avoided.
• Figure 1 depicts a refrigeration cycle of a heat pump system 100 according to an exemplary embodiment of the invention.
• the heat pump system 100 providing a cooling mode and a heating mode, comprises a refrigeration cycle.
• the refrigeration cycle in succession at least comprises a first heat exchanger 10, e.g. an air-to-refrigerant heat exchanger, an accumulator 18, a compressor 22, a second heat exchanger 28, and a first expansion valve 38.
• the compressor 22 is configured to compress a refrigerant, e.g. propane, circulating in the refrigeration cycle.
• a refrigerant e.g. propane
• the second heat exchanger 28 is configured to transfer thermal energy from the compressed refrigerant to a fluid circuit 50 to be heated via a secondary fluid flow line 52 and a secondary fluid flow return line 54.
• the first expansion valve 38 is configured to reduce a pressure of the compressed refrigerant passing through the first expansion valve 38.
• the first heat exchanger 10 is configured to evaporate the refrigerant by transferring thermal energy from a heat transfer medium, e.g. air, to the refrigerant.
• a heat transfer medium e.g. air
• the first expansion valve 38 is connected to a controller module 60 ( Figures 2 and 3 ) at least comprising a feedback controller 62 and a feedforward controller 64.
• a temperature sensor 14 and a pressure sensor 16 are provided at an outlet 11 of the first heat exchanger 10.
• the temperature sensor 14 measures a temperature 15 of the superheated vapor refrigerant at this position.
• the refrigerant temperature 15 may be different from an evaporation temperature or saturation temperature 67 that is a tabular value.
• the pressure sensor 16 measures an evaporation pressure 17 of a superheated vapor refrigerant. Calculation of the evaporation temperature using a refrigerant property table is based on the evaporation pressure 17.
• the refrigeration cycle further comprises an internal heat exchanger 20 having a first conduit 40 and a second conduit 42.
• the first conduit 40 is in heat exchanging contact with the second conduit 42.
• the first conduit 40 is part of the fluid line between the second heat exchanger 28 and the first expansion valve 38 and the second conduit 42 is part of the fluid line between the accumulator 18 and the compressor 22.
• the refrigeration cycle further comprises a fluid bypass line 36 parallel to the first conduit 40 of the internal heat exchanger 20.
• the bypass line 36 to the IHX is in the liquid line of the refrigeration cycle.
• the aim of introducing a bypass to the internal heat exchanger 20 is to vary and/or influence the rate of heat transfer in the internal heat exchanger 20 by varying the amount of fluid flowing through the internal heat exchanger 20 with the objective of increasing the cooling capacity and coefficient of performance.
• a control valve 34 is provided in the refrigeration cycle in series with the internal heat exchanger 20.
• the control valve 34 may also be located in the bypass line 36.
• the control valve 34 may favourably be a ball valve.
• the mass flow rate and hence the rate of heat transfer through the internal heat exchanger 20 can be controlled by the control valve 34.
• the control valve 34 is actuated independent of the first expansion valve 38 with the aim for achieving any one of the following objectives: regulating the suction superheat or the suction temperature to a fixed setpoint and/or maintaining the suction superheat or the suction temperature within a predetermined range.
• the control valve 34 is configured to be controlled at least based on the suction superheat, i.e. the superheat at the inlet of the compressor 22, e.g. by a multi-step control with discrete openings, or by a feedback controller.
• a control scheme of the control valve 34 is depicted later in Figure 6 .
• a discharge pressure sensor 26 is provided in the fluid line between the compressor 22 and the second heat exchanger 28 for measuring a discharge pressure 27.
• a second expansion valve 30 is provided in the fluid line between the second heat exchanger 28 and the internal heat exchanger 20.
• an opening degree 68 of the first expansion valve 38 is controlled by a controller module 60 at least comprising a feedback controller 62 and a feedforward controller 64.
• Figure 2 depicts the controller module 60 of the heat pump system 100 according to Figure 1 .
• Part of the refrigerant cycle is shown depicting the first heat exchanger 10 with the first expansion valve 38.
• the liquid refrigerant 56 enters the first expansion valve 38 streaming to the heat exchanger 10.
• the refrigerant leaves the first heat exchanger 10 at the outlet 11 in a flow 58 to the accumulator 18 ( Figure 1 ). After the outlet 11 the temperature sensor 14 and pressure sensor 16 are located, measuring the refrigerant temperature 15 and the evaporation pressure 17.
• control module 60 Further inputs to the control module 60 are the compressor frequency 23, the air intake temperature 13, the discharge pressure 27 as the condensation pressure and the superheat setpoint 66.
• An output of the control module 60 is controlling the opening degree 68 of the first expansion valve 38.
• controller module 60 Further details of the controller module 60 are depicted in Figure 3 , where a functional behaviour is shown.
• the controller module 60 comprises the feedback controller 62 and the feedforward controller 64.
• the feedback controller 62 comprises a gain scheduled linear feedback controller at least based on an input of the refrigerant temperature 15 and the evaporation pressure 17 of the superheated vapor refrigerant at the outlet 11 of the first heat exchanger 10.
• the feedforward controller 64 comprises a gain scheduled dynamic feedforward controller at least based on an input of a compressor frequency 23.
• Functioning of the feedback controller 62 is based on an input of the temperature sensor 14, which is used to measure the refrigerant temperature 15 of the superheated vapor refrigerant at the outlet 11 of the evaporator of the first heat exchanger 10 and is further based on an input of the pressor sensor 16, which is used to measure the evaporation pressure 17 of the superheated vapor refrigerant at the outlet 11 of the first heat exchanger 10.
• the saturation temperature 67 corresponding to the evaporation pressure 17 is obtained.
• the superheat at the outlet 11 of the first heat exchanger 10 is given by a difference of the refrigerant temperature 15 and the saturation temperature 67.
• the superheat is then compared with the superheat setpoint 66 to generate the control error, which in turn is used as input to the feedback controller 62.
• the parameters of the feedback controller 62, the proportional gain 70 and the integral time 72, are obtained from a parameter adjustment block 76.
• This parameter adjustment block 76 is implemented in the form of so-called gain scheduling functions, one each for the proportional gain and the integral time.
• the gain scheduling functions are implemented as ten-coefficient quadratic polynomials of the evaporation pressure 17, the condensation pressure and the compressor frequency 23.
• the current value of the condensation pressure is obtained from the discharge pressure 27 measured by the discharge pressure sensor 26.
• the current value of the compressor frequency 23 is obtained from the data point of an inverter 24 driving the compressor 22 via bus communication, e.g. MODbus or CANbus.
• the gain scheduled dynamic feedforward controller 64 is used for rejection of disturbances from compressor frequency 23.
• a feedforward gain 74 of the feedforward controller 64 is obtained from a further parameter adjustment block 78.
• the parameter adjustment block 78 is implemented in the form of a gain scheduling function as a linear polynomial of the air intake temperature 13.
• the feedforward gain 74 is used as an input to the feedback controller 62.
• a difference of the compressor frequency 23 minus a base frequency is determined as an input to the feedforward controller 64.
• the current value of the air intake temperature 13 is measured using the air intake temperature sensor 12 ( Figure 1 ), positioned at the first heat exchanger 10.
• the transfer function as an output 99 of the feedforward controller 64 is then given by: G ff s ⁇ K ff 1 + sT 1 1 + sT 2 where K ff is the feedforward gain 74.
• the other two parameters of the feedforward controller 64, the time constants of the pole T 2 and zero T 1 are fixed values. Determining of the control parameter 74 of the feedforward controller 64 is described below.
• the controller module 60 is configured to combine outputs 98, 99 of the feedback controller 62 and the feedforward controller 64.
• the controller module 60 is configured to add and/or multiply and/or weight outputs 98, 99 of the feedback controller 62 and the feedforward controller 64.
• the controller module 60 may be configured to use the feedback controller 62 as a leading controller.
• Figure 4 depicts a controller algorithm of the controller module 60 for regulating a superheat at the evaporator outlet 11 via the first expansion valve 38 of the heat pump system 100 in more detail.
• system identification techniques may be used to obtain the control parameters of the feedback controller 62 for each operating point within the operating envelope 84 of the heat pump system 100, instead of manually obtaining the parameters on a test bench and/or prototype which is extremely time consuming.
• control parameters are the proportional gain 70 ( K p ), integral time 72 ( T i ) and derivative time ( T d ).
• Figure 5 depicts an operating compressor envelope 84.
• Corners of the envelope 84, where air intake temperature 13 and/or water outlet temperature 25 is changing, are marked by corresponding temperature pairs.
• an adaptive controller in the form of gain scheduling may be developed to regulate the degree of superheating at the outlet 11 of the first heat exchanger 10 as evaporator.
• Four steps for the design of a gain scheduled feedback controller 62 may favourably be: system identification to obtain linear models for small deviations about several operating/equilibrium points, design of the linear feedback controller at each of the above linearized operating points, implementation of the gain scheduling functions, and implementing the final non-linear gain scheduled controller on the actual system i.e. experimental verification.
• identification experiments may be performed on the real prototype of the heat pump system 100 with the purpose of obtaining the linear models for the following two input-output responses: An effect of the first expansion valve opening degree 68 on the superheat at the outlet 11 of the first heat exchanger 10 and an effect of the compressor frequency 23 on the superheat at the outlet 11 of the first heat exchanger 10.
• the first expansion valve 38 In total, there are four actuators in the refrigeration cycle of the heat pump system 100 in heating mode: the first expansion valve 38, the first control valve 34, the compressor 22 and a fan of the first heat exchanger 10. In cooling mode the second expansion valve 30 is used.
• step inputs in the form of staircases may be applied to the first expansion valve 38 while holding the other actuators' inputs (compressor frequency 23, fan speed, opening of the first control valve 34) constant. Additionally, the following ambient conditions may be held as stationary as possible: the water outlet and inlet temperature, the water volume flow, the air intake temperature 13, and the air intake humidity.
• Step inputs in the form of staircases may be selected as the excitation signal for the identification experiment because they are easy to apply to the first expansion valve 38 in the real test set-up, and staircase inputs have a large spectrum i.e. the spectral contents of a staircase input are distributed over a wide range of frequencies. As a result, this input signal can excite many modes of the system.
• the superheat at the outlet 11 of the first heat exchanger 10 may be brought nearly to 10°C by controlling the opening degree 68 of the first expansion valve 38 manually.
• the opening of the first expansion valve 38 may be changed in a step manner.
• the opening degree 68 of the first expansion valve 38 may be changed in a step manner and so on for the entire duration of the test.
• Both increasing and decreasing step inputs may be applied to the first expansion valve 38 to obtain data for both the direction of change of superheat.
• the following three scheduling variables may be used to create the gain scheduling functions for the above control parameters: evaporation pressure 17 ( p e ) or evaporation temperature 15 ( T e ), condensation pressure ( p c ) or condensation temperature ( T c ), compressor frequency 23 ( f comp [Hz ]) or speed ( n comp [ RPM ]) .
• the above three variables can be measured in a refrigeration cycle with the help of pressure sensors 16, 26 and temperature sensor 14 or read as datapoint from the inverter 24 of the compressor 22.
• 3D gain scheduling functions may be used for determining the control parameters, the proportional gain 70 ( K p ), integral time 72 ( T i ) and derivative time ( T d ), in the parameter adjustment block 76.
• K p f 1 p e p c f comp
• T i f 2 p e p c f comp
• T d f 3 p e p c f comp
• K p a 1 + a 2 p e + a 3 p c + a 4 f comp + a 5 p e p c + b 6 p
• 3D gain scheduling functions are possible like twenty-coefficients polynomials or any other non-linear combination of the above three scheduling variables.
• a feedback control law may be implemented in the feedback controller 62 which receives the following inputs: the above described gain scheduled control parameters, the proportional gain 70 ( K p ), integral time 72 ( T i ) and derivative time ( T d ), as well as the control error, i.e. the deviation of the calculated superheat from the superheat setpoint 66 determined in the control error calculation block 88.
• the superheat is compared with the superheat setpoint 66 to generate the control error, which in turn is used as input to the feedback controller 62.
• the control error is determined in a control error calculation block 88 as an input to the feedback controller 62.
• the feedback control law is a proportional-integral (PI) controller.
• a dynamic feedforward controller in the form of a variable gain or a lead or a lag or a lead-lag compensator may be implemented.
• EEV is equivalent to the first expansion valve 38
• s is the usual Laplace operator
• ⁇ x EEV ( s ) denotes the changes in opening degree 68 of the first expansion valve 38 required to compensate for the disturbances caused by the changes ⁇ f comp FF in the compressor frequency 23.
• the four parameters of the dynamic feedforward controller i.e. K 1 , K 2 , T 1 , T 2 depend on the operating point. Therefore, there is a need to schedule these four parameters based on the three scheduling variables, p e , p c and f comp . But doing so will create a direct feedthrough/coupling of evaporation pressure at the output of the first heat exchanger 10 to the feedforward controller i.e. any fluctuations in p e due to varying ambient conditions or by the feedback controller will be propagated also to the feedforward controller.
• the compressor frequency is already an input to the feedforward controller. Also these 4 parameters do not depend strongly on p c . Therefore, fixed values of K 1 , T 1 , T 2 can be used which can be obtained by taking the mean values of the parameters of the linearized mathematical models obtained previously from system identification.
• the control signal (output) of the feedforward controller 64 will take both positive and negative values, with zero at the base frequency 82. This ensures that the feedforward controller 64 acts as lagging controller to provide a trimming effect around the control signal of the leading feedback controller 62, especially when there is a change in the compressor frequency 23.
• the base frequency 82 is 50 Hz. In further embodiments, other base frequencies 82 can be used like a mean of the minimum and maximum compressor frequencies 23.
• the air intake temperature 13 may be used as the scheduling variable for the feedforward gain 74 K ff .
• Combining the control signals of the feedback controller 62 and the feedforward controller 64 is implemented in the combine function block 94.
• the outputs 98, 99 of the feedback controller 62 and the feedforward controller 64 are combined in a combine function 94.
• the outputs 98, 99 of the feedback controller 62 and the feedforward controller 64 may be combined in a combine function 94 by adding and/or multiplying and/or weighting both outputs 98, 99.
• the two control signals 98, 99 are simply added together with the gain scheduled feedback controller 62 being the leading controller around which the feedforward controller 64 provides a so-called trimming effect.
• the feedforward controller can also be made the leading one.
• the addition way of combination and making the feedback controller 62 the leading may advantageously help to avoid any unwanted high pressure or low-pressure alarms/faults, thus leading to the shutdown of the heat pump system 100, during starting of the heat pump system 100 or during transition of the operating modes of the heat pump system 100, e.g. cooling to heating or defrost to heating or heating to cooling.
• the combined control signal, the output 97 of the combine block 94 is adapted to the control range limits of the opening degree 68 of the first expansion valve 38, e.g. 0% to 100% before being applied to the actuator, e.g. a stepper motor of the first expansion valve 38 as the opening command.
• saturation is used to limit the outputs of the gain-scheduling functions, the control parameters 70, 72 of the feedback controller 62 and the feedforward gain 74 to some meaningful minimum and maximum values. This is required because it may be possible for the gain scheduling functions to output some wrong values, either in magnitude or in sign.
• the saturation function blocks 90, 92 act as a protective measure against it.
• a PT1 element which is a low pass filter, is used to dampen the high speed fluctuations of the control parameters 70, 72 and the feedforward gain 74 before they are used in the feedback controller 62 and the feedforward controller 64 to calculate the first and second control signals respectively.
• These high speed fluctuations in the parameters can occur due to an error in measurements of evaporation pressure 17, condensation pressure 27, and compressor frequency 23 due to sensor inaccuracies and resolution and/or due to fluctuations in the source and sink sides of the heat pump system 100.
• control valve 34 for controlling the refrigerant flow through the internal heat exchanger 20 is explained hereafter.
• the superheat at the inlet of the compressor 22, the so-called suction superheat, may favourably be maintained between 5 K and 30 K for the safety and proper operation of the compressor 22.
• the liquid refrigerant After leaving the liquid-receiver 32, the liquid refrigerant has two paths to flow to the first expansion valve 38: through the bypass line 36 and through the internal heat exchanger 20.
• the control valve 34 in series with the internal heat exchanger 20
• the flow rate of the refrigerant through the internal heat exchanger 20 is controlled, which in turn controls the rate of heat transfer occurring inside the internal heat exchanger 20.
• heat is transferred from the liquid refrigerant (hot side) to the superheated vapor refrigerant coming from the accumulator 18 (cold side).
• control objective of the control valve control loop is to maintain the suction superheat between the range 5 K to 30 K. Taking safety margins, the range may be reduced to a range of 7 K to 27 K.
• a strong coupling between the control valve 34 and the first expansion valve 38 may be present.
• control valve 34 The consequence of introducing the control valve 34 is the creation of a strong coupling of the superheat at the outlet of the first heat exchanger 10 controlled by first expansion valve 38 with the suction superheat.
• any changes in the control valve opening to regulate the suction superheat will act as a disturbance to the first expansion valve 38 due to the fluctuations in the temperature T junction of the refrigerant entering the first expansion valve 38.
• T junction is the temperature of the liquid refrigerant just after the intersection of the bypass line 36 and the internal heat exchanger 20 and is equal to the temperature at the inlet of the first expansion valve 38.
• the coupling between the control valve 34 and the first expansion valve 38 is strong during the changes in the compressor frequency 23 as is illustrated in Figure 6 .
• step S100 When in step S100 the compressor frequency 23 f comp rises then the suction pressure p suction at the compressor 22 is falling in step S102, followed by the superheat SH Accu in the accumulator 18 rising in step S104 and the superheat SH suction at the compressor 22 rising in step S106.
• step S108 the opening of the control valve 34, x CV ,is decreased in step S108. Inside the internal heat exchanger 20 there will not be enough heat transfer from the hot side to the cold side, step S110. Then the refrigerant may leave the internal heat exchanger 20 in step S112 as two-phase on the hot side and may enter the first expansion valve 38.
• step S114 on the other hand the opening degree 68 of the first expansion valve 38 is rising. Outputs of the steps S114 and S112 are added and this sum is also added to the output of step S112, returning to step S104 with the superheat SH Accu in the accumulator 18 rising.
• control valve control may be implemented in the form of a multi-step control having five fixed discrete opening degrees of the valve: 10%, 32%, 42%, 50% and 65%. These fixed opening degrees are obtained based on the degree of utilization of the internal heat exchanger 20 with the aim of keeping the suction superheat floating between the range 7 K to 27 K.
• control valve control loop may determine an optimized (multi-step discrete) opening of 10%, 32% ... upfront which assists in keeping the suction superheat within the range which ensures the safety and efficient operation of the compressor 22.
• This range can be obtained from the manufacturer's datasheets. For the particular advantageous embodiment the range is 5 to 30K. But taking some safety margins, the range is reduced to 7 to 27K.
• the superheat control loop of the first expansion valve 38 then regulates the first expansion valve 38 according to the actual opening of the control valve 34 to enable an optimized system setup for the current operating point.
• An alternative to the multi-step controller would be a feedback controller like a PID controller with the aim of regulating the suction superheat to some fixed setpoint value like , e.g.,15 K.
• the control signal of the PID controller (opening degrees of the control valve 34) would take any value between 0 and 100%. In this case, the PIDs of the control valve 34 and the first expansion valve 38 will be interacting with each other a lot.
• the interaction between the control valve 34 and the first expansion valve 38 may be reduced because the control valve 34 will be changing its position less as it is restricted to only five discrete openings. With this it may be ensured that the disturbance to the first expansion valve 38 coming from the junction temperature, T junction is a less frequent one i.e. a minor disturbance which can be rejected by the gain scheduled PI controller 62 of the first expansion valve 38.

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• 2024
  • 2024-04-22 EP EP24171608.3A patent/EP4641118A1/de active Pending

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