Heat source device

Abstract

To provide a heat source device that reduces loss of a screw compressor.SOLUTION: A heat source device 1 includes: a screw compressor 10 having a screw rotor 11, a slide valve 13 and an electric motor 12; an inverter 21; a condensation part 23; an evaporation part 25; an outflow heating medium temperature detector 45 for detecting an outlet temperature of a heating medium C, etc.; a heat source device output detector 93 for detecting output of the heat source device 1, etc.; a correction amount grasp section 95; and a control section 91. The correction amount grasp section 95 grasps an optimal volume ratio, etc. that is a ratio of suction volume when a ratio between a condensation pressure and an evaporation pressure becomes equal to a compression ratio of a rotor 11 to maximum suction volume. The control section 91 moves the slide valve 13 to a direction for returning to a target value when a detection value obtained by the outflow heating medium temperature detector 45 is deviated from the target value, and controls the inverter 21 so that a rotating speed of the rotor 11 changes in proportion to the detection value obtained by the heat source device output detector 93 and in inverse proportion to a value grasped by the correction amount grasp section.SELECTED DRAWING: Figure 1

Description

Technical Field

[0001] The present disclosure relates to a heat source device, and more particularly to a heat source device including a screw compressor having a slide valve.

Background Art

[0002] Some screw compressors installed in screw refrigerators can change their internal volume by moving a slide valve. And the following are screw compressors that can always operate at maximum efficiency according to the load. Based on the suction-side and discharge-side pressures of the screw compression section and the rotational frequency of the electric motor that rotationally drives the screw compression section, the opening degree of the slide valve is controlled so as to achieve the highest compression efficiency corresponding to the rotational frequency (see, for example, Patent Document 1).

Prior Art Documents

Patent Documents

[0003]

Patent Document 1

Summary of the Invention

Problems to be Solved by the Invention

[0004] Specifically, the screw refrigerator described in Patent Document 1 controls the opening degree of the slide valve based on the optimal internal volume ratio at the current rotational frequency, which is calculated based on the suction pressure, discharge pressure, and rotational frequency detected by sensors. However, particularly in a screw compressor that changes the suction volume by a slide valve, the compression ratio may change when the slide valve is moved, resulting in a decrease in efficiency.

[0005] In view of the above problems, the present disclosure relates to providing a heat source device that reduces the losses of a screw compressor having a slide valve that makes the suction volume variable.

Means for Solving the Problems

[0006] A heat source device according to a first aspect of the present disclosure is a heat source device that cools or heats a heat medium supplied to a heat demand facility, and includes a screw rotor that compresses vapor of a refrigerant that exchanges heat with the heat medium, a slide valve that variably adjusts a suction volume of the vapor of the refrigerant into the screw rotor, an electric motor that rotates the screw rotor, a screw compressor having the electric motor, an inverter that changes a rotation speed of the screw rotor by the electric motor, a condensing unit that condenses the vapor of the refrigerant, an evaporating unit that evaporates the liquid of the refrigerant, an outflow heat medium temperature detector that detects a temperature of the heat medium flowing out of the heat source device or a physical quantity correlated therewith, a heat source device output detector that detects an output of the heat source device or a physical quantity correlated therewith, a correction amount acquisition unit that acquires an optimum volume ratio defined by a ratio of the suction volume when the ratio of a condensing pressure when the vapor of the refrigerant condenses in the condensing unit to an evaporation pressure when the liquid of the refrigerant evaporates in the evaporating unit and a compression ratio of the vapor of the refrigerant in the screw rotor becomes equal, and is defined by a ratio to the maximum suction volume, or a physical quantity correlated therewith, and a control unit that controls the slide valve and the inverter. The control unit moves the slide valve in a direction to return to the target value when a value detected by the outflow heat medium temperature detector deviates from a preset target value, and controls the inverter such that the rotation speed of the screw rotor changes in proportion to a value detected by the heat source device output detector and in inverse proportion to a value acquired by the correction amount acquisition unit. Here, changing in inverse proportion to a value acquired by the correction amount acquisition unit includes changing in proportion to the reciprocal of a value acquired by the correction amount acquisition unit.

[0007] With such a configuration, it is possible to bring the compression ratio of the vapor of the refrigerant in the screw compressor having the slide valve that variably adjusts the suction volume closer to the ratio of the condensing pressure and the evaporation pressure of the refrigerant during operation, and it is possible to reduce the loss of the screw compressor.

[0008] Further, the heat source device according to the second aspect of the present disclosure is the heat source device according to the first aspect of the present disclosure, and includes a condensation pressure detector that detects the condensation pressure and an evaporation pressure detector that detects the evaporation pressure.

[0009] With this configuration, the condensation pressure and the evaporation pressure can be grasped.

[0010] Further, the heat source device according to the third aspect of the present disclosure is the heat source device according to the second aspect of the present disclosure, and the correction amount grasping unit is configured to grasp, as a physical quantity correlated with the optimum volume ratio, a value proportional to a pressure ratio that is a ratio of the condensation pressure detected by the condensation pressure detector to the evaporation pressure detected by the evaporation pressure detector.

[0011] With this configuration, a physical quantity correlated with the optimum volume ratio can be grasped relatively simply.

[0012] Further, the heat source device according to the fourth aspect of the present disclosure is the heat source device according to the second aspect of the present disclosure, and the correction amount grasping unit is configured to grasp, as a physical quantity correlated with the optimum volume ratio, a value proportional to the n-th root of a pressure ratio that is a ratio of the condensation pressure detected by the condensation pressure detector to the evaporation pressure detected by the evaporation pressure detector when the polytropic index is n.

[0013] With this configuration, control with improved accuracy can be performed, and the reduction of the loss of the screw compressor can be performed more effectively.

[0014] Further, in the heat source device according to the fifth aspect of the present disclosure, in the heat source device according to the second aspect of the present disclosure, the correction amount grasping unit is configured to grasp, as a physical quantity correlated with the optimum volume ratio, an experimentally obtained value corresponding to the pressure ratio which is the ratio of the condensation pressure detected by the condensation pressure detector and the evaporation pressure detected by the evaporation pressure detector. The experimentally obtained value is the ratio of the output of the heat source device when the efficiency of the screw compressor is maximized at an arbitrary position of the slide valve to the output of the heat source device at the position of the slide valve where the suction volume of the refrigerant vapor is maximized.

[0015] With this configuration, it is possible to grasp a physical quantity correlated with the optimum volume ratio relatively simply and accurately.

[0016] Further, in the heat source device according to the sixth aspect of the present disclosure, in the heat source device according to the second aspect of the present disclosure, when the pressure ratio which is the ratio of the condensation pressure detected by the condensation pressure detector and the evaporation pressure detected by the evaporation pressure detector exceeds the design pressure ratio, the control unit resets the control of the slide valve and the inverter, moves the slide valve to the position where the suction volume of the refrigerant vapor is maximized, and controls the inverter so as to make the value detected by the outflow heat medium temperature detector the target value by adjusting the rotational speed of the screw rotor.

[0017] With this configuration, it is possible to make the compression ratio of the refrigerant vapor in the screw compressor closer to the pressure ratio, and it is possible to suppress an increase in the loss of the screw compressor.

[0018] Further, in the heat source device according to the seventh aspect of the present disclosure, in the heat source device according to any one of the first to sixth aspects of the present disclosure, when the heat source device operates to heat the heat medium, the control unit resets the control of the slide valve and the inverter, moves the slide valve to a position where the suction volume of the refrigerant vapor is maximized, and controls the inverter so that the value detected by the outflow heat medium temperature detector becomes the target value by adjusting the rotational speed of the screw rotor.

[0019] With this configuration, in the operation of heating the heat medium, which is likely to increase the output of the heat source device, the loss of the screw compressor can be easily reduced.

[0020] Further, in the heat source device according to the eighth aspect of the present disclosure, in the heat source device according to any one of the first to seventh aspects of the present disclosure, a heat medium passage through which the heat medium passes, a heat source fluid passage through which a heat source fluid that exchanges heat with the refrigerant passes, and a first switching valve that switches between guiding the refrigerant vapor compressed by the screw compressor to the heat medium passage and guiding it to the heat source fluid passage are provided. When the heat source device operates to cool the heat medium, the control unit guides the refrigerant vapor compressed by the screw compressor to the heat source fluid passage, and when the heat source device operates to heat the heat medium, the control unit guides the refrigerant vapor compressed by the screw compressor to the heat medium passage, and controls the first switching valve. When the heat source device operates to cool the heat medium, the heat medium passage functions as the evaporation section and the heat source fluid passage functions as the condensation section, and when the heat source device operates to heat the heat medium, the heat medium passage functions as the condensation section and the heat source fluid passage functions as the evaporation section.

[0021] With this configuration, it is possible to switch between the operation of cooling the heat medium and the operation of heating the heat medium without switching the flow path of the heat medium.

[0022] Further, the heat source device according to the ninth aspect of the present disclosure is the heat source device according to any one of the first to seventh aspects of the present disclosure, and includes a second switching valve that switches between the heat medium passing through the evaporation section and the heat medium passing through the condensation section. When the heat source device operates to cool the heat medium, the control section causes the heat medium to pass through the evaporation section, and when the heat source device operates to heat the heat medium, the control section controls the second switching valve so that the heat medium passes through the condensation section.

[0023] With this configuration, it is possible to switch between the operation of cooling the heat medium and the operation of heating the heat medium without changing the refrigerant flow path.

[0024] Further, the heat source device according to the tenth aspect of the present disclosure is the heat source device according to any one of the first to ninth aspects of the present disclosure, and includes an inflow heat medium temperature detector that detects the temperature of the heat medium flowing into the heat source device or a physical quantity correlated therewith, and a heat medium flow rate detector that detects the flow rate of the heat medium passing through the heat source device or a physical quantity correlated therewith. The heat source device output detector is configured to obtain the output of the heat source device or a physical quantity correlated therewith by multiplying the difference between the value detected by the inflow heat medium temperature detector and the value detected by the outflow heat medium temperature detector by the value detected by the heat medium flow rate detector.

[0025] With this configuration, it is possible to grasp the output of the heat source device or a physical quantity correlated therewith without providing a flow meter for the heat medium.

Advantages of the Invention

[0026] According to the present disclosure, the compression ratio of the refrigerant vapor in the screw compressor having a slide valve that variably changes the suction volume can be made close to the ratio of the condensation pressure and the evaporation pressure of the refrigerant during operation, and the loss of the screw compressor can be reduced.

Brief Description of the Drawings

[0027]

Figure 1

Figure 2

Figure 3

Figure 4

Mode for Carrying Out the Invention

[0028] Hereinafter, embodiments will be described with reference to the drawings. In each figure, the same or corresponding members are denoted by the same or similar reference numerals, and redundant descriptions are omitted.

[0029] First, referring to FIG. 1, a screw chiller 1 according to an embodiment will be described. FIG. 1 is a schematic system diagram showing the schematic configuration of the screw chiller 1. The screw chiller 1 is a device that cools chilled water C that processes the heat load of a heat demand facility (not shown), and is a form of a heat source device. Examples of the heat demand facility (not shown) include air conditioning equipment such as an air handling unit and a fan coil unit. The chilled water C is a medium that conveys cold heat to the heat demand facility (not shown), and corresponds to a heat medium. The screw chiller 1 is configured such that a refrigerant R that performs a refrigeration cycle takes heat from the chilled water C and transfers the taken heat to the cooling water CD, thereby cooling the chilled water C. The screw chiller 1 includes, as main elements, a screw compressor 10 (hereinafter simply referred to as "compressor 10"), an inverter 21, a condenser 23, an evaporator 25, and a control device 90.

[0030] The compressor 10 compresses the vapor of the refrigerant R by the rotation of the screw rotor 11 (hereinafter simply referred to as "rotor 11"). In addition to the rotor 11, the compressor 10 has an electric motor 12 for rotating the rotor 11 and a slide valve 13. The rotor 11 is composed of a pair of a male rotor and a female rotor. The male rotor and the female rotor each have a generally cylindrical appearance, and spiral grooves are formed on the outer surface of the cylinder. The rotor 11 is configured such that spiral compression chambers are formed between the contact lines where the male rotor and the female rotor contact each other and between the cylindrical end faces where the rotor 11 rotates. By rotating, the rotor 11 is configured to compress the vapor of the refrigerant R by changing the volume while the compression chamber moves in the axial direction of the cylinder. The rotor 11 is configured to suck the vapor of the refrigerant R from the suction port provided on the cylindrical end face and discharge the vapor of the refrigerant R compressed by rotation from the discharge port provided on the opposite end face of the cylinder. For the rotor 11, the suction port and the discharge port communicate with or are partitioned from the compression chamber by the rotation of the rotor 11. The rotor 11 is configured such that the compression ratio, which is the ratio of the suction volume to the discharge volume (suction volume / discharge volume), becomes the designed compression ratio. Lubricating oil L circulates inside the compressor 10. The lubricating oil L is pressurized together with the vapor of the refrigerant R, and the gas and liquid are separated on the discharge side. The lubricating oil L returns to the suction side of the rotor 11, and the compressed vapor of the refrigerant R is discharged from the discharge port of the compressor 10. That is, the driving source of the lubricating oil L is the discharge pressure itself.

[0031] The rotational speed of the electric motor 12 changes when the frequency of the input power changes. In the present embodiment, the rotational speed of the electric motor 12 is changed by an inverter 21. In the present embodiment, a general-purpose induction motor is used for the electric motor 12, but an inverter - dedicated electric motor may be used. The inverter 21 is typically configured to appropriately adjust the frequency of the power input from the commercial power supply and supply power to the electric motor 12.

[0032] In the present embodiment, the slide valve 13 is a member that changes the suction volume of the refrigerant R vapor into the rotor 11. The slide valve 13 is provided so as to be in contact with the rotor 11 and to be part of a member that cooperates with the cylindrical portions of the female rotor and the male rotor to form a compression chamber. When the slide valve 13 moves, the compressor 10 can change the suction volume of the rotor 11 because the closed point (hereinafter referred to as the "trapping point") where the suction side of the compression chamber is closed changes. Due to such a change in volume, the screw compressor 1 can change the compression ratio of the refrigerant R and the flow rate of the vapor relatively easily and steplessly with little loss. Since the flow rate of the refrigerant R vapor is obtained by multiplying the suction volume of the rotor 11 by the rotational speed and the number of rotor 11, the flow rate of the refrigerant R vapor is proportional to the refrigeration output of the screw refrigerator 1 approximately. Note that there is a slide valve for a general screw compressor that changes the discharge volume by changing the end position of compression, but the slide valve 13 in the present embodiment changes the suction volume by changing the start position of compression. That is, in the compressor 10 in the present embodiment, while the suction volume changes according to the position of the slide valve 13, the discharge volume is configured to be constant without changing. Therefore, in the compressor 10 in the present embodiment, the compression ratio increases or decreases in proportion to the suction volume. Also, in the compressor 10 in the present embodiment, the suction volume is configured to change substantially linearly according to the position of the slide valve 13. The relationship among the position of the slide valve 13, the suction volume, and the compression ratio of the compressor 10 in the present embodiment is illustrated in FIG. 2. In the following description, the moving direction of the slide valve 13 in which the suction volume increases may be expressed as the "increasing force side", and the moving direction of the slide valve 13 in which the suction volume decreases may be expressed as the "decreasing force side". Also, the position of the slide valve 13 where the suction volume is maximum may be referred to as the "maximum position".

[0033] The slide valve 13 is connected to the piston 14 inside the compressor 10 and can move to the force-increasing side or the force-decreasing side by the reciprocating movement of the piston 14. In the present embodiment, it is configured such that the piston 14 can be reciprocated by the lubricating oil L circulating inside the compressor 10 described above. A specific example of this configuration in the present embodiment is as follows. One end of a load line 15 for injecting the lubricating oil L and one end of an unload line 17 for extracting the lubricating oil L are connected to the cylinder in which the piston 14 is accommodated. The other end of the load line 15 is connected to the discharge side of the rotor 11. A load valve 16 is provided in the load line 15. The other end of the unload line 17 is connected to the suction side of the rotor 11. An unload valve 18 is provided in the unload line 17. When the slide valve 13 opens the load valve 16 and injects the lubricating oil L into the cylinder of the piston 14 via the load line 15, it moves to the force-increasing side. Further, when the slide valve 13 opens the unload valve 18 and extracts the lubricating oil L from the cylinder of the piston 14 via the unload line 17, it moves to the force-decreasing side. The load line 15 upstream of the load valve 16 and the unload line 17 downstream of the unload valve 18 are connected by a bypass line 19. Thereby, the lubricating oil L discharged from the rotor 11 can bypass the piston 14 and enter the suction side of the rotor 11.

[0034] The condenser 23 introduces and condenses the vapor of the refrigerant R compressed by the compressor 10, and corresponds to the condensation section. The condenser 23 introduces, in addition to the vapor of the refrigerant R compressed by the compressor 10, cooling water CD as a heat source fluid, and cools the vapor of the refrigerant R with the cooling water CD to turn it into a liquid of the refrigerant R. A condensation pressure gauge 33 for detecting the condensation pressure of the refrigerant R inside is provided in the condenser 23. The condensation pressure gauge 33 corresponds to the condensation pressure detector. The cooling water CD that has taken the condensation heat from the vapor of the refrigerant R in the condenser 23 is sent to a cooling tower (not shown) to be cooled and then introduced into the condenser 23 again. The liquid of the refrigerant R condensed in the condenser 23 passes through the expansion valve 24, has its pressure reduced, and then is introduced into the evaporator 25. The evaporator 25 introduces the liquid of the refrigerant R that has flowed out of the condenser 23 and passed through the expansion valve 24 and evaporates it, and corresponds to the evaporation section. The evaporator 25 introduces, in addition to the liquid of the refrigerant R, cold water C, and cools the cold water C by taking the heat (latent heat of evaporation) required when the liquid of the refrigerant R evaporates from the cold water C. An evaporation pressure gauge 35 for detecting the evaporation pressure of the refrigerant R inside is provided in the evaporator 25. The evaporation pressure gauge 35 corresponds to the evaporation pressure detector. In the screw chiller 1, the compressor 10, the condenser 23, the expansion valve 24, and the evaporator 25 are provided in this order in the refrigerant flow path 28 formed in a ring shape. The refrigerant R circulates through the refrigerant flow path 28 to the compressor 10, the condenser 23, the expansion valve 24, the evaporator 25, and then again to the compressor 10, and performs a refrigeration cycle while undergoing a phase change between vapor and liquid.

[0035] The evaporator 25 is provided with a cold water outflow pipe 41 and a cold water inflow pipe 42. The cold water outflow pipe 41 is a flow path through which the cooled cold water C flows out toward a heat demand facility (not shown). The cold water inflow pipe 42 is a flow path through which the cold water C with an increased temperature that has returned from the heat demand facility (not shown) flows in. An outlet thermometer 45 for detecting the temperature of the cold water C flowing out of the evaporator 25 is provided in the cold water outflow pipe 41. The outlet thermometer 45 corresponds to an outflow heat medium temperature detector. In the cold water inflow pipe 42, an inlet thermometer 46 for detecting the temperature of the cold water C flowing into the evaporator 25 and a flow meter 48 for detecting the flow rate of the cold water C passing through the evaporator 25 are provided. The inlet thermometer 46 corresponds to an inflow heat medium temperature detector, and the flow meter 48 corresponds to a heat medium flow rate detector. As the flow meter 48, an electromagnetic flow meter, a positive displacement flow meter, or the like can be used. Note that the flow meter 48 may be provided in the cold water outflow pipe 41 instead of the cold water inflow pipe 42. Further, the condenser 23 is provided with a cooling water outflow pipe 51 through which the cooling water CD with an increased temperature flows out toward a cooling tower (not shown) and a cooling water inflow pipe 52 through which the cooling water CD from the cooling tower (not shown) flows in.

[0036] The control device 90 is a device that controls the operation of the screw chiller 1. The control device 90 has a control unit 91 that controls the operation of the devices constituting the screw chiller 1. The control unit 91 is connected to the load valve 16 and the unload valve 18 by a communication line, and is configured to be able to control the position of the slide valve 13 by controlling the opening and closing of the load valve 16 and the unload valve 18. Further, the control unit 91 is connected to the inverter 21 by a communication line, and is configured to be able to control the rotational speed of the electric motor 12, and thus the rotational speed of the rotor 11, by giving a frequency command to the inverter 21. Further, the control unit 91 is configured to be able to grasp the condensation pressure and the evaporation pressure of the refrigerant R by receiving pressure information signals from the condensation pressure gauge 33 and the evaporation pressure gauge 35, respectively. Further, the control unit 91 is configured to be able to grasp the outlet temperature and the inlet temperature of the chilled water C by receiving temperature information signals from the outlet thermometer 45 and the inlet thermometer 46, respectively. Further, the control unit 91 is configured to be able to grasp the flow rate of the chilled water C passing through the evaporator 25 by receiving a signal of the flow rate information of the chilled water C from the flow meter 48. Further, in the present embodiment, the control device 90 has an output calculation unit 93. The output calculation unit 93 is a part that calculates the output (refrigeration output) of the screw chiller 1 based on the outlet temperature, the inlet temperature, and the flow rate of the chilled water C received by the control unit 91 from the outlet thermometer 45, the inlet thermometer 46, and the flow meter 48. The output of the screw chiller 1 is obtained by multiplying the difference between the inlet temperature and the outlet temperature of the chilled water C by the flow rate ((inlet temperature of chilled water C - outlet temperature of chilled water C) × flow rate of chilled water C). Thus, the outlet thermometer 45, the inlet thermometer 46, the flow meter 48, and the output calculation unit 93 cooperate to constitute a heat source device output detector.

[0037] In this embodiment, the control device 90 has a correction amount grasping unit 95. The correction amount grasping unit 95 is a part that grasps the amount for correcting the rotational speed of the rotor 11. In this embodiment, the correction amount grasping unit 95 is configured to grasp the optimum volume ratio or a physical quantity correlated therewith. Here, the optimum volume ratio is defined as the ratio of the suction volume when the pressure ratio and the compression ratio of the rotor 11 are equal, to the maximum suction volume ((suction volume when "pressure ratio = compression ratio") / maximum suction volume). Note that the pressure ratio is the ratio of the condensation pressure of the refrigerant R to the evaporation pressure (condensation pressure of the refrigerant R / evaporation pressure of the refrigerant R). In this embodiment, the correction amount grasping unit 95 is configured to grasp, as a physical quantity correlated with the optimum volume ratio, a value proportional to the pressure ratio. That is, the ratio of the pressure ratio between the evaporator and the condenser under the rated conditions (the position of the slide valve 13 is at the position where the suction volume is maximum, that is, the compression ratio is maximum) to the pressure ratio under the operating state is used as the physical quantity correlated with the optimum volume ratio. This is based on the fact that the suction volume and the compression ratio are proportional (see Fig. 2). Strictly speaking, it is the internal compression ratio that is proportional to the suction volume, which is different from the true compression ratio considering the temperature rise. However, if the difference is small and can be ignored, the physical quantity correlated with the optimum volume ratio can be regarded as a value proportional to the pressure ratio. This can be expressed by the following formula. First, let the discharge volume be represented by V2. As described above, in this embodiment, the discharge volume V2 is constant. And when the suction volume of the compressor 10 under the rated conditions is V0 and the compression ratio is R0, the compression ratio R0 ≒ V0 / V2. Also, when the suction volume of the compressor 10 at an arbitrary position of the slide valve 13 is V1 and the compression ratio is Rx, the compression ratio Rx ≒ V1 / V2. From these and the definitions, the optimum volume ratio = V1 / V0 = (V1 / V2) / (V0 / V2) ≒ Rx / R0. Here, since the compression ratio R0 under the rated conditions is a constant of the compressor 10, the physical quantity correlated with the optimum volume ratio can be regarded as proportional to the ideal compression ratio Rx under the operating state. And since the ideal compression ratio Rx under the operating state is equal to the pressure ratio, the ratio of the pressure detected by the condensation pressure gauge 33 to the pressure detected by the evaporation pressure gauge 35 can be regarded as the physical quantity correlated with the optimum volume ratio.

[0038] In this embodiment, the control unit 91 is configured to control the position of the slide valve 13 through the control of the opening and closing of the load valve 16 and the unloading valve 18, and to control the rotational speed of the rotor 11 through the inverter 21. The control unit 91 controls the position of the slide valve 13 so that the temperature of the chilled water C detected by the outlet thermometer 45 becomes the target value (for example, 7°C). Specifically, when the temperature of the chilled water C detected by the outlet thermometer 45 exceeds the target value, the control unit 91 is configured to move the slide valve 13 to the boosting side, and when it is lower than the target value, to move the slide valve 13 to the reducing side. On the other hand, the control unit 91 controls the rotational speed of the rotor 11 to be proportional to the output of the screw chiller 1 calculated by the output calculation unit 93 and inversely proportional to the pressure ratio (optimal volume ratio or a physical quantity correlated thereto) grasped by the correction amount grasping unit 95. Specifically, when the output calculated by the output calculation unit 93 increases, the control unit 91 is configured to increase the rotational speed of the rotor 11, and when it decreases, to decrease the rotational speed of the rotor 11. Further, when the pressure ratio grasped by the correction amount grasping unit 95 increases, the control unit 91 is configured to decrease the rotational speed of the rotor 11, and when it decreases, to increase the rotational speed of the rotor 11. In a conventional screw chiller with a variable rotational speed of the rotor, the rotational speed of the rotor is changed based on the temperature of the supplied chilled water, and the slide valve is moved based on the pressure ratio. In contrast, in the screw chiller 1 according to this embodiment, the physical quantities that are the basis for the rotational speed of the rotor 11 and the movement of the slide valve 13 are in a swapped form compared to the conventional variable rotational speed screw chiller.

[0039] Generally, in a positive displacement compressor including a screw compressor, the compression flow rate of the suction gas (e.g., refrigerant vapor) per unit time changes according to the rotational speed of the compressor. In a screw compressor with a variable suction volume, the compression ratio of the compressor changes according to the position of the slide valve. At this time, since the suction volume also changes, the compression flow rate per rotational speed of the compressor also changes. At this time, even if the actual compression ratio of the compressor is unknown because the actual position of the slide valve is unknown, the compression flow rate per rotational speed of the compressor at the compression ratio (ideal compression ratio) corresponding to the pressure ratio at that time can be obtained from the characteristics of the compressor. As described above, since the flow rate of the refrigerant vapor is proportional to the refrigeration output of the screw chiller, dividing the refrigeration output by the compression flow rate per rotational speed of the compressor at the ideal compression ratio can obtain the "optimal rotational speed when assuming that the slide valve is at the optimal position". When the compressor is actually operated at the optimal rotational speed obtained in this way, if the slide valve is closer to the increasing side than the optimal position, the refrigeration output will be greater than the assumption, so the chilled water temperature will decrease. If it is closer to the decreasing side, the chilled water temperature will increase. Therefore, while controlling the rotational speed of the compressor rotor as described above, when the chilled water temperature decreases, move the slide valve to the decreasing side, and when it increases, move it to the increasing side. This allows the position of the slide valve of the compressor to move to the optimal position, and the chilled water temperature will stabilize at the target value when it reaches the optimal position. By doing so, it becomes possible to stabilize both the position of the slide valve and the rotational speed of the compressor rotor at the optimal positions. The screw chiller 1 according to the present embodiment utilizes this principle to control the position of the slide valve 13 and the rotational speed of the rotor 11.

[0040] Continuing to refer to FIG. 1, the operation of the screw chiller 1 will be described. For the time being, the operation under rated conditions will be described. Under rated conditions, the slide valve 13 is positioned at the position where the suction volume of the rotor 11 is maximized, and the rotor 11 rotates at a rotational speed of 100%. Prior to the start of the screw chiller 1, a chilled water pump (not shown) for flowing chilled water C and a cooling water pump (not shown) for flowing cooling water CD are started. When the chilled water pump (not shown) operates, the chilled water C comes to pass through the evaporator 25. When the cooling water pump (not shown) operates, the cooling water CD comes to pass through the condenser 23. After the operations of the chilled water pump (not shown) and the cooling water pump (not shown) are started, the control unit 91 starts the screw chiller 1.

[0041] When the screw chiller 1 starts, the control unit 91 operates the electric motor 12 via the inverter 21 to rotate the rotor 11. Due to the rotation of the rotor 11, the refrigerant R circulating in the refrigerant flow path 28 enters the compression chamber formed in the rotor 11 in a vapor state. The vapor of the refrigerant R that has flowed into the compression chamber of the rotor 11 is compressed as the compression chamber moves in the axial direction of the rotor 11 after the suction side of the compression chamber is closed, and is discharged from the rotor 11 in a state compressed to the rated compression ratio. The compressed vapor of the refrigerant R that has exited the rotor 11 flows into the condenser 23 and is cooled by the cooling water CD passing through the inside of the condenser 23 to become a liquid of the refrigerant R. When the liquid of the refrigerant R that has exited the condenser 23 passes through the expansion valve 24, the pressure decreases to become a low-pressure liquid of the refrigerant R. The low-pressure liquid of the refrigerant R flows into the evaporator 25 and is heated by taking the latent heat of evaporation from the chilled water C passing through the inside of the evaporator 25 and evaporates to become a vapor of the refrigerant R. Note that the chilled water C from which heat has been taken by the low-pressure liquid of the refrigerant R has its temperature decreased and flows out of the evaporator 25 and flows through the chilled water outlet pipe 41. The vapor of the refrigerant R that has exited the evaporator 25 enters the rotor 11 again, is compressed in the above-described manner, and thereafter, the above-described cycle is repeated.

[0042] The screw chiller 1 according to this embodiment adjusts the position of the slide valve 13 and / or the rotational speed of the rotor 11 to adjust the output when the load of the heat demand facility (the supply destination of the chilled water C) changes during operation as described above. Here, for comparison, a conventional screw chiller is mentioned. As described above, since the flow rate of the refrigerant R vapor is substantially proportional to the refrigeration output, in a conventional fixed-speed screw compressor, the refrigeration output can be adjusted to match the load by adjusting the slide valve to a position that balances the refrigeration load without changing the rotational speed of the rotor. On the other hand, in recent years, controlling the refrigeration output using variable-speed technology with an inverter has also been widely practiced. In the case of a positive displacement compressor not limited to the screw type, when an inverter is used, the flow rate of the refrigerant vapor changes substantially in proportion to the rotational speed of the rotor, and it is possible to easily adjust the refrigeration output to match the load even in a scroll or reciprocating compressor. Even in the case of a screw type, control with an inverter has been widely practiced, but when an inverter is used, in many cases, the slide valve is not used, and the refrigeration output is controlled only by the rotational speed of the rotor. In this case, the compression ratio that changes according to the position of the slide valve does not change even when the load fluctuates. When the compression ratio determined by the position of the slide valve is different from the actual pressure ratio, the following disadvantages occur. When the pressure ratio is smaller than the compression ratio, in the compressor, the refrigerant is compressed to an unnecessary pressure, and then the pressure suddenly decreases at the moment when the port on the outlet side opens. This means that extra energy is consumed to compress to an unnecessary pressure, so the efficiency of the compressor decreases. On the other hand, when the pressure ratio is larger than the compression ratio, the refrigerant vapor in the compressor will communicate with the refrigerant vapor in the condenser before reaching the condensation pressure. In this case, the uncondensed refrigerant vapor in the condenser will flow back into the compressor, and the compressor will compress this again. Therefore, the efficiency of the compressor also decreases. That is, when trying to maximize the efficiency of a screw chiller, it is desirable to appropriately control the flow rate of the refrigerant vapor by variable-speed control of the rotor and to control using a slide valve so that the compression ratio is equal to the pressure ratio. However, in practice, it is difficult to perform such control.First, the screw compressor of a screw chiller is often made hermetic (generally semi-hermetic that can be disassembled and inspected) to suppress refrigerant leakage. However, since the slide valve is housed in a sealed mechanism, it is difficult to know its exact position. Therefore, even if the appropriate position of the slide valve can be calculated from the operating state, it is difficult to actually control the slide valve to that position. Also, although the internal compression ratio can be calculated based on the position of the slide valve etc., in actual compression, since the temperature of the refrigerant vapor rises with adiabatic compression, it is difficult to calculate the compression ratio (the ratio of the discharge pressure to the suction pressure of the compressor). In this regard, although it is possible to perform the calculation by assuming the specific heat ratio etc., since it involves exponential function calculation, it is difficult to perform the calculation with a general chiller controller (sequencer etc.). Based on the above circumstances, in the screw chiller 1 according to the present embodiment, when the load of the heat demand facility (the supply destination of the chilled water C) changes during operation, the following control is to be performed.

[0043] When the screw chiller 1 according to the present embodiment is operating under rated conditions, as described above, the slide valve 13 is positioned at the position where the suction volume of the rotor 11 is maximized, and the rotor 11 is rotating at a rotational speed of 100%. Generally, since the compressor is designed such that the point where the suction volume of the rotor 11 is maximized, that is, the compression ratio is maximized, becomes the rated condition, the efficiency of the compressor is maximized during operation under the rated condition. For the time being, it is assumed that the pressure ratio does not change for the explanation. This is because the pressure ratio is generally determined according to external conditions (such as the temperature of the cooling water CD and the temperature of the chilled water C during the operation of the screw chiller 1) and is not controlled by the control unit 91. During the operation of the screw chiller 1 under rated conditions, when the load of the heat demand facility (not shown) decreases, the refrigeration output of the screw chiller 1 exceeds the load of the heat demand facility (not shown), so the temperature of the chilled water C gradually decreases. Then, since the outlet temperature of the chilled water C is lower than the target temperature, the slide valve 13 controlled based on the outlet temperature of the chilled water C moves to the reducing side. Then, since the refrigeration output decreases, the rotational speed of the rotor 11 controlled to be proportional to the refrigeration output decreases in proportion to the decrease in the refrigeration output. Due to the decrease in the rotational speed of the rotor 11, the refrigeration output further decreases, so this time the temperature of the chilled water C exceeds the target temperature. For this reason, the position of the slide valve 13 moves to the increasing side and returns to the original position. By repeating this, the temperature of the chilled water C returns to the target value, the position of the slide valve 13 also returns to the original position, and the rotational speed of the rotor 11 converges to the rotational speed corresponding to the load.

[0044] Next, assume that the temperature conditions are different and the ideal compression ratio (pressure ratio) corresponds to a position other than the maximum position of the slide valve 13. As an example, assume that the load in this case is about 80% of the refrigeration output when the rotor 11 is rotated at the rated rotational speed at the position of this slide valve 13. In this case, the ideal rotational speed of the rotor 11 is about 80% of the rated rotational speed. Also, assuming that the suction volume of the rotor 11 at this time is 70% of the maximum suction volume, the compression ratio of the compressor 10 becomes 70% of the rated compression ratio. At this time, the load of the screw chiller 1 is 56% considering 70% regarding the slide valve 13 and 80% regarding the rotational speed of the rotor 11. In this situation, if the rotational speed of the rotor 11 is higher than 80%, the refrigeration output will exceed 56% and the temperature of the chilled water C will decrease. In response to the decrease in the temperature of the chilled water C, the slide valve 13 moves to the pressure-reducing side. Then, since the refrigeration output decreases, the rotational speed of the rotor 11 decreases in proportion to the decrease in the refrigeration output. Then, since the temperature of the chilled water C rises, next the slide valve 13 moves to the pressure-increasing side. By repeating this, the slide valve 13 converges to the optimal position and the rotational speed of the rotor 11 converges to the optimal value. On the other hand, assume that the position of the slide valve 13 is on the pressure-increasing side rather than the 70% position. In this case, if the rotational speed of the rotor 11 is 80%, the refrigeration output will be greater than 56%. Therefore, the rotor 11, which is controlled to be proportional to the refrigeration output, rotates at a higher speed, and accordingly the refrigeration output exceeds 56%. Then, since the temperature of the chilled water C decreases, the position of the slide valve 13, which is controlled based on the outlet temperature of the chilled water C, moves to the pressure-reducing side. Then, the refrigeration output decreases, and along with that, the rotational speed of the rotor 11 also changes, and again the slide valve 13 stabilizes at the optimal position and the rotational speed of the rotor 11 stabilizes at an appropriate value.

[0045] As described above, in a situation where there is no fluctuation in the pressure ratio, the screw chiller 1 according to the present embodiment controls the position of the slide valve 13 based on the outlet temperature of the chilled water C, and controls the rotational speed of the rotor 11 to be proportional to the refrigeration output. By controlling in this way, even if the position of the current slide valve 13 is unknown, the position of the slide valve 13 is controlled so as to obtain a desirable compression ratio with respect to the refrigeration output at that time. For this reason, in the screw chiller 1, while suppressing a decrease in the efficiency of the compressor 10 as the compression ratio of the compressor 10 approaches the pressure ratio, an appropriate refrigeration output can be obtained by adjusting the rotational speed of the rotor 11. When the pressure ratio changes due to some factor, the control unit 91 can grasp the change in the pressure ratio by receiving the signal of the pressure information from the condenser pressure gauge 33 and the evaporator pressure gauge 35. When the control unit 91 grasps the change in the pressure ratio, it is preferable to correct the rotational speed of the rotor 11 so as to be inversely proportional to the optimum volume ratio grasped by the correction amount grasping unit 95 or a physical quantity correlated therewith. This is to move the slide valve 13 to an optimum position with respect to the changed pressure ratio again when the pressure ratio changes due to some factor. For example, when the pressure ratio increases due to some factor, it is preferable to move the slide valve 13 to the increasing side so that the compression ratio of the compressor 10 approaches the changed pressure ratio. In the present embodiment, this is done by "lowering" the rotational speed of the rotor 11 to the optimum rotational speed at the pressure ratio. That is, when the position of the slide valve 13 is moved to an optimum position for the increased pressure ratio, the suction volume of the rotor 11 increases, so the flow rate of the refrigerant R vapor increases. Therefore, the rotational speed of the rotor 11 is decreased to the optimum rotational speed at this time. When the rotational speed of the rotor 11 decreases, the refrigeration output temporarily decreases and the temperature of the chilled water C rises. However, due to this rise in the temperature of the chilled water C, the slide valve 13 moves to the increasing side, the refrigeration output increases again, and stabilizes at the point when it balances with the load (the optimum position of the slide valve 13). In this way, in order to move the slide valve 13 to an optimum position with respect to the changed pressure ratio, the rotational speed of the rotor 11 is decreased (inversely proportional).Hereinafter, controlling the position of the slide valve 13 so that the temperature of the supplied heat medium (for example, cold water C) reaches the target temperature and controlling the rotational speed of the rotor 11 to be proportional to the output of the heat source device and inversely proportional to the optimum volume ratio or a physical quantity correlated thereto may be referred to as "optimum control".

[0046] In the previous description, it was assumed that the correction amount grasping unit 95 is configured to grasp a value proportional to the pressure ratio as a physical quantity correlated with the optimum volume ratio. However, it may be configured as follows. As a physical quantity correlated with the optimum volume ratio according to the first modification example, it may be configured to grasp a value proportional to the n-th root (n is the polytropic index) of the pressure ratio. Expressing this in a mathematical formula is as follows. First, in this embodiment, as described above, the discharge volume V2 is constant. Also, let the suction volume of the compressor 10 under rated conditions be V0 and the compression ratio be R0, and let the suction volume of the compressor 10 at an arbitrary position of the slide valve 13 be V1 and the compression ratio be Rx, which is the same as described above. And as described above, the internal compression ratio is V1 / V2. However, when the polytropic index is n, the actual compression ratio Rx is from the general formula of polytropic compression "PV n = constant", Rx = P2 / P1 = (V1 / V2) n Here, P1 is the suction pressure and P2 is the discharge pressure. From the above formula for the compression ratio Rx, V1 / V2 = (P2 / P1) 1 / n = Rx 1 / n is derived. From these and the definitions, the optimum volume ratio = V1 / V0 = (V1 / V2) / (V0 / V2) = (Rx / R0) 1 / n , and thus, the optimum volume ratio may be obtained by taking the 1 / n-th power of Rx / R0 when configured to grasp a value proportional to the aforementioned pressure ratio. Here, when the control device 90 does not have a function for exponential calculation, the value of the optimum volume ratio based on the pressure ratio may be calculated in advance and the optimum volume ratio may be calculated by performing calculations such as so-called table operations.

[0047] Also, as a physical quantity correlated with the optimum volume ratio according to the second modification example, it may be configured to grasp an experimentally obtained value corresponding to the pressure ratio. Examples of the experimentally obtained values are shown below. First, the screw chiller 1 is operated with the rotational speed of the rotor 11 at the maximum (100%). In this state, while fixing the position of the slide valve 13 at an arbitrary position, the chiller is operated while changing the temperature of the cooling water CD (or the discharge pressure of the refrigerant R). At this time, at each temperature of the cooling water CD (or the discharge pressure of the refrigerant R), the outlet temperature of the refrigerant R in the compressor 10 and the efficiency of the compressor 10 are calculated from the efficiency of the screw chiller 1 and the cycle efficiency of the refrigerant R. The efficiency of the compressor 10, as an example, can be obtained from the formula "Efficiency = (h ad -h in ) / (h out -h in )". Here, h in is the specific enthalpy of the refrigerant R vapor at the inlet of the compressor 10, which is a function of the temperature and pressure of the refrigerant R at the inlet of the compressor 10. Also, h out is the specific enthalpy of the refrigerant R vapor at the outlet of the compressor 10, which is a function of the temperature and pressure of the refrigerant R at the outlet of the compressor 10. Also, h ad is the specific enthalpy at the outlet when the refrigerant R vapor undergoes an isentropic change in the compressor 10, which is a function of the temperature and pressure of the refrigerant R at the inlet of the compressor 10. After obtaining the efficiency of the compressor 10 at each temperature of the cooling water CD (or the discharge pressure of the refrigerant R), the pressure ratio and the refrigeration output when the efficiency of the compressor 10 is maximized are obtained. Here, since the optimum volume ratio can be considered to be directly the ratio of the refrigeration output when the slide valve 13 is at the maximum position to the refrigeration output at the current position of the slide valve 13, a relational expression between the optimum volume ratio and the pressure ratio can be obtained in this way. In this case, it may be possible to obtain the optimum volume ratio from the pressure ratio of the screw chiller 1 during operation by the above-described table calculation.

[0048] In addition, when the control unit 91 grasps the change in the pressure ratio, making the rotation speed of the rotor 11 inversely proportional to the optimal volume ratio grasped by the correction amount grasping unit 95 or a physical quantity correlated therewith includes making it proportional to the reciprocal of the optimal volume ratio or a physical quantity correlated therewith. For example, it may be possible to perform a table operation on the reciprocal of the optimal volume ratio grasped by the correction amount grasping unit 95 or a physical quantity correlated therewith, and make the rotation speed of the rotor 11 proportional thereto. In this case, since the refrigeration output of the screw chiller 1 is proportional to the temperature difference between the inlet and outlet of the chilled water C in the evaporator 25, the rotation speed of the rotor 11 may be made proportional to the temperature difference between the inlet and outlet of the chilled water C. This can be applied not only to the above-described embodiment regarding the value grasped by the correction amount grasping unit 95, but also to the cases of the above-described first modification and second modification.

[0049] Here, mention will be made of the case where even when the rotor 11 of the compressor 10 is operated at the maximum rotation speed at a certain position of the slide valve 13, the load of the heat demand facility cannot be satisfied. For example, assuming that the suction volume (optimal volume ratio) of the rotor 11 when the compression ratio and the pressure ratio are equal is 50% instead of 70% described above, and the load of the screw chiller 1 is 60%. At this time, according to the above-described embodiment, if the rotation speed of the rotor 11 is set to 120%, the load can be satisfied, but in reality, it can only be increased to 100%, so the refrigeration output will remain at 50%. Therefore, since the temperature of the chilled water C rises, the slide valve 13 gradually moves toward the increasing side, and finally moves to the point where the volume ratio is 60% and stabilizes here. In this case, the compression ratio becomes approximately 60%, which is higher than the ideal 50%, but compared with the case where the slide valve 13 is fixed and only the rotation speed of the rotor 11 is changed for operation (rotation speed 60%, compression ratio 100%), the compression ratio approaches the actual pressure ratio. For this reason, the loss of the compressor 10 is reduced, and energy saving can be achieved.

[0050] Next, referring to FIGS. 3(A) and 3(B), the screw heat source machine 2 according to another embodiment will be described. FIGS. 3(A) and 3(B) are both schematic system diagrams showing the schematic configuration of the screw heat source machine 2. FIG. 3(A) represents the state during cooling, and FIG. 3(B) represents the state during heating. The screw heat source machine 2 has a function of cooling a heat medium, similar to the screw refrigerator 1 (see FIG. 1), and is also a device that can be used as a heat pump for heating the heat medium, and is a form of a heat source device. The screw heat source machine 2 is typically configured to be able to cool the heat medium during cooling and heat the heat medium during heating. The screw heat source machine 2 differs from the screw refrigerator 1 (see FIG. 1) in the following points. In the screw heat source machine 2, instead of the condenser 23 and the evaporator 25 in the screw refrigerator 1 (see FIG. 1), a heat source fluid passage 123 and a heat medium passage 125 are provided respectively. The heat source fluid passage 123 and the heat medium passage 125 have the same configuration as the condenser 23 (see FIG. 1) and the evaporator 25 (see FIG. 1) respectively, but are given different names due to differences in function. The heat medium passage 125 corresponding to the evaporator 25 (see FIG. 1) in the screw refrigerator 1 (see FIG. 1) in the screw heat source machine 2 is configured to cool or heat the heat medium supplied to the heat demand facility (not shown). Therefore, a cold water outflow pipe 41 and a cold water inflow pipe 42 are provided in the heat medium passage 125.

[0051] The heat source fluid passage device 123 is typically configured to give heat from the refrigerant R to the cooling water CD during cooling and to take heat from the cooling water CD by the refrigerant R during heating. That is, the heat source fluid passage device 123 is typically configured to function as a condensing unit during cooling and as an evaporating unit during heating. The heat medium passage device 125 is configured such that the chilled and warm water CH passes therethrough as a heat medium. The chilled and warm water CH is a medium for transporting cold heat or warm heat to a heat demand facility (not shown). The chilled and warm water CH is a general term for cold water as a cold heat medium and warm water as a warm heat medium, and typically becomes cold water during cooling and warm water during heating. Therefore, in the screw heat source machine 2, the chilled and warm water CH flows through the cold water outflow pipe 41 and the cold water outflow pipe 41. The heat medium passage device 125 is typically configured to take heat from the chilled and warm water CH by the refrigerant R during cooling and to give heat from the refrigerant R to the chilled and warm water CH during heating. That is, the heat medium passage device 125 is typically configured to function as an evaporating unit during cooling and as a condensing unit during heating.

[0052] Further, in the screw heat source machine 2, a four-way valve 29 is provided in the refrigerant flow path 28. The four-way valve 29 is provided across the refrigerant flow path 28 between the compressor 10 and the heat source fluid passage 123 and the refrigerant flow path 28 between the compressor 10 and the heat medium passage 125. The four-way valve 29 is configured to be able to switch the flow of the refrigerant R in the refrigerant flow path 28 between the following two states. One state of the four-way valve 29 is a state in which the refrigerant R discharged from the compressor 10 is guided to the heat source fluid passage 123 and the refrigerant R discharged from the heat medium passage 125 is guided to the compressor 10, as shown in Fig. 3(A). Typically, this state is set during cooling operation. The other state of the four-way valve 29 is a state in which the refrigerant R discharged from the compressor 10 is guided to the heat medium passage 125 and the refrigerant R discharged from the heat source fluid passage 123 is guided to the compressor 10, as shown in Fig. 3(B). Typically, this state is set during heating operation. Thus, the four-way valve 29 is configured to be able to switch between guiding the vapor of the refrigerant R compressed by the compressor 10 to the heat source fluid passage 123 and guiding it to the heat medium passage 125, and corresponds to the first switching valve. The four-way valve 29 is connected to the control unit 91 by a communication line and is configured to perform the above switching in response to a command from the control unit 91.

[0053] In the screw heat source machine 2 according to the present embodiment, the compressor 10 is capable of compressing to a compression ratio during heat pump operation (typically during heating) under the condition that the internal compression ratio is sufficiently large and the slide valve 13 is at the maximum position. Note that the screw heat source machine 2 can be provided with heat pump-specific devices such as a buffer tank for adjusting the circulation amount of the refrigerant R, as necessary. The configuration of the screw heat source machine 2 other than the above is the same as that of the screw refrigerator 1 (see Fig. 1).

[0054] In the screw heat source machine 2 configured as described above, the operation and control during cooling (see Fig. 3(A)) are the same as those of the screw chiller 1 (see Fig. 1). On the other hand, the operation during heating (see Fig. 3(B)) is as follows. The vapor of the refrigerant R compressed by the compressor 10 flows into the heat medium passage 125, gives heat to the chilled and hot water CH passing through the inside of the heat medium passage 125, and itself condenses into a liquid of the refrigerant R. The chilled and hot water CH that has received heat from the vapor of the refrigerant R flows out of the heat medium passage 125 with an increased temperature and flows through the cold water outlet pipe 41. The liquid of the refrigerant R that has exited the heat medium passage 125 has its pressure reduced when passing through the expansion valve 24 and becomes a low-pressure liquid of the refrigerant R. The low-pressure liquid of the refrigerant R flows into the heat source fluid passage 123 and is heated and evaporated by the cooling water CD passing through the inside of the heat source fluid passage 123 to become a vapor of the refrigerant R. The vapor of the refrigerant R that has exited the heat source fluid passage 123 enters the compressor 10 again and repeats the above cycle.

[0055] The control during heating in the screw heat source machine 2 is different from that during cooling in the following aspects. First, regarding the control of the slide valve 13, when the temperature of the chilled and hot water CH exceeds the target temperature, the unloading valve 18 is opened, and when it is lower than the target temperature, the loading valve 16 is opened. This is because moving the slide valve 13 of the compressor 10 to the increasing force side will lower the temperature of the chilled and hot water CH during cooling, but will increase the temperature of the chilled and hot water CH during heating. That is, when the temperature of the chilled and hot water CH exceeds the target temperature during heating, opening the unloading valve 18 and when it is lower than the target temperature, opening the loading valve 16 will move the slide valve 13 in the direction to return the temperature of the chilled and hot water CH to the target temperature. Next, regarding the control of the rotational speed of the rotor 11, for the pressure ratio in the correction amount grasping unit 95, since the condenser part and the evaporator part are swapped compared to during cooling, the pressure ratio can be calculated by swapping them. That is, during heating, the evaporation pressure of the refrigerant R is detected by the condenser pressure gauge 33, and the condensation pressure of the refrigerant R is detected by the evaporator pressure gauge 35. Therefore, for the method of calculating the pressure ratio during cooling, the values detected by the condenser pressure gauge 33 and the evaporator pressure gauge 35 can be swapped. For the optimum volume ratio, it can be calculated in the same way as during the cooling operation. In addition, in the output calculation unit 93, the heating output is obtained from the temperature difference between the inlet and outlet of the chilled and hot water CH and the flow rate of the chilled and hot water CH in the same manner as during cooling. From these, the control of the rotational speed of the rotor 11 during heating can be corrected to be proportional to the heating output obtained by the output calculation unit 93 and inversely proportional to the optimum volume ratio grasped by the correction amount grasping unit 95 or a physical quantity correlated with this.

[0056] In the description of the screw heat source machine 2 above, it was assumed that the compressor 10 is capable of compressing up to the compression ratio during heat pump operation (typically during heating). However, it is also possible to use the compressor designed for cooling operation as it is during heat pump operation. In this case, generally during the heating operation of the heat pump machine, the pressure ratio between the condensing section and the evaporating section becomes larger than that during the cooling operation. Therefore, even when the position of the slide valve 13 is at the maximum position, the designed compression ratio of the compressor 10 is small relative to the pressure ratio during operation, and the efficiency of the compressor 10 decreases. In this regard, during heat pump operation, together with the fact that the increase in power (loss) due to the decrease in the efficiency of the compressor 10 is also effectively used as heating output, it is often used in such a state. In this case, in the method of controlling by the optimal control as described above, there may be a case where the position of the slide valve 13 and the rotational speed of the rotor 11 cannot be appropriately controlled. This is because, even though the rotational speed of the rotor 11 is controlled in an attempt to set the position of the slide valve 13 to the optimal position, the position of the slide valve 13 cannot move beyond the maximum position. In such a case, the control unit 91 shall perform the control shown below instead of performing the optimal control. In other words, the control unit 91 shall change the control mode (reset the optimal control). In the control that substitutes for the optimal control, the position of the slide valve 13 is always set to the maximum position, and the temperature of the chilled / hot water CH is controlled to the target temperature (desired temperature) by controlling the rotational speed of the rotor 11. Specifically, when the pressure ratio during operation is larger than the designed compression ratio of the compressor 10, the control unit 91 switches the control mode, keeps the load valve 16 always open, sets the position of the slide valve 13 to the maximum position at all times, and controls the rotational speed of the rotor 11 so that the temperature of the chilled / hot water CH becomes the target value. By doing so, even when the pressure ratio is larger than the designed compression ratio of the compressor 10, the position of the slide valve 13 and the rotational speed of the rotor 11 can be appropriately controlled. Note that by switching the control mode based on the pressure ratio in this way, not only during heating but also when the pressure ratio exceeds the designed compression ratio of the compressor 10 for some reason, the position of the slide valve 13 and the rotational speed of the rotor 11 can be appropriately controlled.

[0057] Next, referring to FIG. 4, the screw heat source machine 3 according to yet another embodiment will be described. FIG. 4 is a schematic system diagram showing the schematic configuration of the screw heat source machine 3. The screw heat source machine 3 has a function of cooling a heat medium, similar to the screw chiller 1 (see FIG. 1), and is also a device that can be used as a heat pump for heating the heat medium, and is one form of a heat source device. The screw heat source machine 3 is typically configured to be able to cool the heat medium during cooling and heat the heat medium during heating. The screw heat source machine 3 has a generally similar configuration compared to the screw chiller 1 (see FIG. 1), but the configuration around the cold water outflow pipe 41, the cold water inflow pipe 42, the cooling water outflow pipe 51, and the cooling water inflow pipe 52 is different. It should be noted that the fact that the heat medium in the screw heat source machine 3 is chilled water CH is also different from the screw chiller 1 (see FIG. 1). In the screw heat source machine 3, the cold water outflow pipe 41 and the cooling water outflow pipe 51 are connected via a switching forward pipe 61. An evaporation-side three-way valve 64 is provided at a portion where one end of the switching forward pipe 61 is connected to the cold water outflow pipe 41. A condensation-side three-way valve 65 is provided at a portion where the other end of the switching forward pipe 61 is connected to the cooling water outflow pipe 51. Also, in the screw heat source machine 3, the cold water inflow pipe 42 and the cooling water inflow pipe 52 are connected via a switching return pipe 62. A return three-way valve 66 is provided at a portion where one end of the switching return pipe 62 is connected to the cold water inflow pipe 42.

[0058] In addition, one end of the heat source return pipe 71 is connected to the cold water outflow pipe 41 of the screw heat source machine 3, and one end of the heat source supply pipe 72 is connected to the cold water inflow pipe 42. A heat source three-way valve 74 is provided at the portion where one end of the heat source return pipe 71 is connected to the cold water outflow pipe 41. The other end of the heat source return pipe 71 and the other end of the heat source supply pipe 72 are connected to the heat source equipment HS via the equipment-side piping. For the heat source equipment HS, for example, an air heat exchanger (radiator), a heating tower, river water, sewer heat, ground heat, etc. can be applied. Note that, depending on the type, a heat exchanger may be provided between these heat source equipment HS. Further, when an air heat exchanger is particularly used as the heat source equipment HS, since it is necessary to set the heat source fluid below the freezing point, an antifreeze liquid (brine) may be used instead of fresh water. In the screw heat source machine 3 according to the present embodiment, a heat source that does not drop below the freezing point such as ground heat is used, and fresh water is used as the heat source fluid HF. A heat source fluid pump (not shown) for flowing the heat source fluid HF is provided between the heat source return pipe 71 and the heat source equipment HS. The heat source fluid pump (not shown) is typically provided outside the screw heat source machine 3, but may be provided, for example, on the heat source return pipe 71 or the cold water outflow pipe 41 as a component of the screw heat source machine 3. Further, the screw heat source machine 3 according to the present embodiment is such that the compressor 10 can be compressed to the compression ratio during the heat pump operation (typically during heating) under the condition that the internal compression ratio is sufficiently large and the slide valve 13 is at the maximum position. The evaporation-side three-way valve 64, the condensation-side three-way valve 65, the return three-way valve 66, and the heat source three-way valve 74 are each connected to the control unit 91 by a communication line and are configured to be able to switch the flow path in response to a command from the control unit 91. The configuration of the screw heat source machine 3 other than the above is the same as that of the screw refrigerator 1 (see FIG. 1).

[0059] In the screw heat source machine 3 configured as described above, typically during cooling operation, the chilled and warm water CH flowing through the chilled water outflow pipe 41 is supplied to the heat demand facility (not shown) without flowing into the switching forward pipe 61 and the heat source return pipe 71. Thus, the flow paths of the evaporation-side three-way valve 64 and the heat source three-way valve 74 are determined. Also, the flow path of the return three-way valve 66 is determined such that the chilled and warm water CH flowing through the chilled water inflow pipe 42 flows into the evaporator 25 without flowing into the switching return pipe 62. Further, the flow path of the condenser-side three-way valve 65 is determined such that the cooling water CD flowing through the cooling water outflow pipe 51 is guided to the cooling tower (not shown) without flowing into the switching forward pipe 61. In such a state, the screw heat source machine 3 typically performs the same operation and control as the screw chiller 1 (see FIG. 1) during cooling operation.

[0060] In the screw heat source machine 3, typically during heating operation, the flow path of the heat source three-way valve 74 is determined such that the chilled and warm water CH flowing out of the evaporator 25 and through the chilled water outflow pipe 41 is guided to the heat source facility HS via the heat source return pipe 71. Also, the flow path of the return three-way valve 66 is determined such that the chilled and warm water CH returning from the heat demand facility (not shown) and entering the chilled water inflow pipe 42 flows into the condenser 23 via the switching return pipe 62 without being introduced into the evaporator 25. Further, the flow paths of the evaporation-side three-way valve 64 and the condenser-side three-way valve 65 are determined such that the chilled and warm water CH flowing through the cooling water outflow pipe 51 after passing through the condenser 23 is supplied to the heat demand facility (not shown) through the end of the chilled water outflow pipe 41 via the switching forward pipe 61 without being introduced into the cooling tower (not shown). The evaporation-side three-way valve 64, the condenser-side three-way valve 65, and the return three-way valve 66 are valves that switch whether the chilled and warm water CH passes through the evaporator 25 or the condenser 23 and correspond to the second switching valve. Note that some or all of the evaporation-side three-way valve 64, the condenser-side three-way valve 65, the return three-way valve 66, and the heat source three-way valve 74 may be configured to perform the same function by providing a plurality of two-way valves instead of the three-way valves. In such a state, the screw heat source machine 3 typically operates as follows during heating operation.

[0061] When a cold water pump (not shown) operates, the cold and hot water CH circulates between a heat demand facility (not shown) and the condenser 23. At this time, starting from the condenser 23, the cold and hot water CH flows through a part of the cooling water outflow pipe 51, the switching forward pipe 61, a part of the cold water outflow pipe 41, the heat demand facility (not shown), a part of the cold water inflow pipe 42, the switching return pipe 62, and a part of the cooling water inflow pipe 52 in this order and returns to the condenser 23 again. When the cold and hot water CH passes through the condenser 23, it is heated by the condensation heat released when the refrigerant R condenses, and the temperature rises. The cold and hot water CH with the increased temperature has its heat utilized in the heat demand facility (not shown), returns to the condenser 23 in a state where the temperature has decreased, is reheated in the condenser 23, and is supplied to the heat demand facility (not shown). On the other hand, when a heat source fluid pump (not shown) operates, the heat source fluid HF circulates between the heat source facility HS and the evaporator 25. At this time, starting from the evaporator 25, the heat source fluid HF flows through a part of the cold water outflow pipe 41, the heat source return pipe 71, the heat source facility HS, the heat source forward pipe 72, and a part of the cold water inflow pipe 42 in this order and returns to the evaporator 25 again. When the heat source fluid HF passes through the evaporator 25, it is cooled by having the latent heat of evaporation required for the refrigerant R to evaporate taken away, and the temperature decreases. The heat source fluid HF with the decreased temperature is heated in the heat source facility HS, returns to the evaporator 25 in a state where the temperature has increased, is cooled again in the evaporator 25, and is sent to the heat source facility HS. Note that the cooling water pump (not shown) does not operate during heating.

[0062] The control during heating in the screw heat source machine 3 according to this embodiment is typically performed as follows. First, the flow rate of the heat source fluid HF flowing in and out of the evaporator 25 is detected. In this embodiment, a flow meter 48 can be used. If there is no flow meter, the flow rate can be obtained from the differential pressure of the heat source fluid HF at the inlet and outlet of the evaporator 25. At this time, when the heat source fluid is a fluid different from that during cooling, such as brine, or when the temperature of the heat source fluid HF is significantly different from the temperature of the chilled warm water CH during the cooling operation and the viscosity is significantly different, a different proportionality constant will be used in the heating mode. This switching may be linked to the switching between the cooling mode and the heating mode. In this embodiment, as described above, the heat source fluid HF uses fresh water, just like the chilled warm water CH, and the temperature is approximately the same as the temperature of the chilled warm water CH during cooling, so the same proportionality constant can be used. After detecting the flow rate of the heat source fluid HF, based on this flow rate, the load on the evaporator 25, that is, the recovered heat quantity, is calculated. This calculation can be obtained by multiplying the temperature difference between the inlet and outlet of the heat source fluid HF to the evaporator 25 by the flow rate of the heat source fluid HF. This is the same as multiplying the temperature difference between the inlet and outlet of the chilled warm water CH by the flow rate of the chilled warm water CH during the cooling operation, so it can be easily calculated. Also, the rotational speed of the rotor 11 can be calculated in the same way by replacing the refrigeration output during the cooling operation with the recovered heat quantity. This is because there is no difference in the relationship between the suction volume of the compressor 10, the optimum compression ratio, and the optimum rotational speed of the rotor 11 in this case, whether it is a refrigerator or a heat pump. Here, regarding the movement of the slide valve 13, when the temperature at the outlet of the condenser 23 exceeds the target temperature of the chilled warm water CH, the unloader valve 18 is opened and it is moved to the reduced force side. When it is lower, the loader valve 16 is opened and it is moved to the increased force side.

[0063] In the description of the screw heat source machine 3 above, it was assumed that the compressor 10 can be compressed to the compression ratio during heat pump operation (typically during heating). However, it is also possible to use the compressor designed for cooling operation as it is during heat pump operation. In this case, when the pressure ratio during operation is greater than the designed compression ratio of the compressor 10, the control unit 91 may change the control mode (reset the optimal control). In the control that replaces the optimal control, the position of the slide valve 13 is always set to the maximum position, and the temperature of the chilled and heated water CH at the outlet of the condenser 23 may be controlled to the target temperature (desired temperature) by controlling the rotational speed of the rotor 11. By switching the control mode based on the pressure ratio in this way, it is possible to appropriately control the position of the slide valve 13 and the rotational speed of the rotor 11 not only during heating but also when the pressure ratio exceeds the designed compression ratio of the compressor 10 for some reason.

[0064] In the above description, it was assumed that the outlet heat medium temperature detector is the outlet thermometer 45 that detects the temperature of the chilled water C flowing out of the evaporator 25. However, instead of directly detecting the temperature of the chilled water C at the outlet of the evaporator 25 with the outlet thermometer 45, it is also possible to indirectly detect the temperature of the chilled water C at the outlet of the evaporator 25 (detect a physical quantity correlated with the temperature of the chilled water C at the outlet of the evaporator 25). Examples of the physical quantity correlated with the temperature of the chilled water C at the outlet of the evaporator 25 include the temperature inside the supply header of the chilled water C on the side of the heat demand facility (not shown), the temperature of the chilled water C at the outlet of the manifold of the screw chiller where the module is installed, and the like. The same applies when detecting the temperature of the chilled warm water CH flowing out of the heat medium passage 125 in the screw heat source machine 2, and when detecting the temperature of the chilled warm water CH flowing out of the condenser 23 during the heating operation of the screw heat source machine 3. Also, it was assumed that the inlet heat medium temperature detector is the inlet thermometer 46 that detects the temperature of the chilled water C flowing into the evaporator 25. However, instead of directly detecting the temperature of the chilled water C at the inlet of the evaporator 25 with the inlet thermometer 46, it is also possible to indirectly detect the temperature of the chilled water C at the inlet of the evaporator 25 (detect a physical quantity correlated with the temperature of the chilled water C at the inlet of the evaporator 25). Examples of the physical quantity correlated with the temperature of the chilled water C at the inlet of the evaporator 25 include the temperature inside the return header of the chilled water C on the side of the heat demand facility (not shown), the temperature of the chilled water C at the inlet of the distribution pipe of the screw chiller where the module is installed, and the like. The same applies when detecting the temperature of the chilled warm water CH flowing into the heat medium passage 125 in the screw heat source machine 2, and when detecting the temperature of the chilled warm water CH flowing into the condenser 23 during the heating operation of the screw heat source machine 3.

[0065] In the above description, it is assumed that the heat medium flow rate detector includes the flow meter 48. However, instead of the flow meter 48 that directly measures the heat medium (cooling water C, chilled and warm water CH), a pressure gauge for obtaining the flow rate from the differential pressure may be provided. In this case, pressure gauges may be provided in each of the cooling water outflow pipe 41 and the cooling water inflow pipe 42, and the flow rate may be obtained from the differential pressure of the inflow and outflow of the cooling water C or the chilled and warm water CH with respect to the evaporator 25 (or the heat medium passage 125 or the condenser 23). Typically, the control unit 91 may be configured to receive the pressure values detected by each pressure gauge and calculate the flow rate. In this case, each pressure gauge and the control unit 91 constitute the heat medium flow rate detector, and the pressure values detected by each pressure gauge are physical quantities correlated with the flow rate of the heat medium.

[0066] In the above description, it is assumed that the heat source device output detector detects the output of the screw chiller 1. However, instead of this, it may be configured to detect a physical quantity correlated with the output of the screw chiller 1. Examples of the physical quantity correlated with the output of the screw chiller 1 include the temperature difference between the inside of the supply header and the return header of the cooling water C on the side of the heat demand facility (not shown), and the temperature difference between the inlet and outlet of the manifold (the output when the flow rate of the cooling water C is assumed to be the rated value). The same applies to the screw heat source machine 2 and the screw heat source machine 3.

[0067] In the above description, it is assumed that the control device 90 includes the control unit 91, the output calculation unit 93, and the correction amount grasping unit 95. The control unit 91, the output calculation unit 93, and the correction amount grasping unit 95 may be integrally configured within the control device 90, but a part or all of the output calculation unit 93 and / or the correction amount grasping unit 95 may be configured separately from the control unit 91. Further, those of the output calculation unit 93 and / or the correction amount grasping unit 95 that are configured separately from the control unit 91 may be arranged close to the control unit 91 inside or outside the control device 90, or may be arranged away from the control unit 91 and the control device 90.

Explanation of Signs

[0068] 1 Screw chiller 2 Screw heat source machine 3 Screw heat source machine 10 Compressor (Screw Compressor) 11 Rotor (Screw Rotor) 12 Electric Motor 13 Slide Valve 21 Inverter 23 Condenser 25 Evaporator 29 Four-way Valve (First Switching Valve) 33 Condensing Pressure Gauge (Condensing Pressure Detector) 35 Evaporating Pressure Gauge (Evaporating Pressure Detector) 45 Outlet Thermometer (Outflow Heat Medium Temperature Detector) 46 Inlet Thermometer (Inflow Heat Medium Temperature Detector) 48 Flow Meter (Heat Medium Flow Detector) 64 Three-way Valve on the Evaporation Side (Second Switching Valve) 65 Three-way Valve on the Condensing Side (Second Switching Valve) 66 Return Three-way Valve (Second Switching Valve) 90 Control Device 91 Control Unit 93 Output Calculation Unit (Heat Source Device Output Detector) 95 Correction Amount Grasping Unit 123 Heat Source Fluid Passer 125 Heat Medium Passer C Chilled Water CH Chilled Warm Water CD Cooling Water (Heat Source Fluid) R Refrigerant

Claims

1. A heat source device for cooling or heating a heat medium supplied to a heat demand device, a screw compressor having a screw rotor that compresses vapor of a refrigerant that exchanges heat with the heat medium, a slide valve that variably adjusts the suction volume of the refrigerant vapor into the screw rotor, and an electric motor that rotates the screw rotor, an inverter that changes the rotational speed of the screw rotor by the electric motor, a condensing unit that condenses the refrigerant vapor, an evaporating unit that evaporates the refrigerant liquid, an outflow heat medium temperature detector that detects the temperature of the heat medium flowing out of the heat source device or a physical quantity correlated therewith, a heat source device output detector that detects the output of the heat source device or a physical quantity correlated therewith, a correction amount grasping unit that grasps an optimum volume ratio defined by a ratio of the suction volume when the ratio of the condensing pressure when the refrigerant vapor condenses in the condensing unit to the evaporation pressure when the refrigerant liquid evaporates in the evaporating unit and the compression ratio of the refrigerant vapor in the screw rotor becomes equal, to the suction volume when it is maximum, or a physical quantity correlated therewith, a control unit that controls the slide valve and the inverter, wherein the control unit moves the slide valve in a direction to return the value detected by the outflow heat medium temperature detector to a preset target value when the detected value deviates from the target value, and controls the inverter such that the rotational speed of the screw rotor changes in proportion to the value detected by the heat source device output detector and in inverse proportion to the value grasped by the correction amount grasping unit. Heat source device.

2. a condensing pressure detector that detects the condensing pressure, an evaporation pressure detector that detects the evaporation pressure, The heat source device according to claim 1.

3. The correction amount grasping unit is configured to grasp a value proportional to a pressure ratio that is a ratio of the condensing pressure detected by the condensing pressure detector to the evaporation pressure detected by the evaporation pressure detector as a physical quantity correlated with the optimum volume ratio. The heat source device according to claim 2.

4. The correction amount grasping unit is configured to grasp a value proportional to the n-th root of a pressure ratio that is a ratio of the condensing pressure detected by the condensing pressure detector to the evaporation pressure detected by the evaporation pressure detector as a physical quantity correlated with the optimum volume ratio when the polytropic index is n. The heat source device according to claim 2.

5. The correction amount grasping unit is configured to grasp an experimentally obtained value corresponding to a pressure ratio, which is a ratio of the condensation pressure detected by the condensation pressure detector to the evaporation pressure detected by the evaporation pressure detector, as a physical quantity correlated with the optimal volume ratio. The experimentally obtained value is a ratio of the output of the heat source device when the efficiency of the screw compressor is maximized at an arbitrary position of the slide valve to the output of the heat source device at the position of the slide valve where the suction volume of the refrigerant vapor is maximized. The heat source device according to claim 2.

6. When the pressure ratio, which is a ratio of the condensation pressure detected by the condensation pressure detector to the evaporation pressure detected by the evaporation pressure detector, exceeds the design pressure ratio, the control unit resets the control of the slide valve and the inverter, moves the slide valve to a position where the suction volume of the refrigerant vapor is maximized, and controls the inverter so that the value detected by the outflow heat medium temperature detector becomes the target value by adjusting the rotational speed of the screw rotor. The heat source device according to claim 2.

7. When the heat source device operates to heat the heat medium, the control unit resets the control of the slide valve and the inverter, moves the slide valve to a position where the suction volume of the refrigerant vapor is maximized, and controls the inverter so that the value detected by the outflow heat medium temperature detector becomes the target value by adjusting the rotational speed of the screw rotor. The heat source device according to claim 2.

8. A heat medium passage through which the heat medium passes; A heat source fluid passage through which a heat source fluid that exchanges heat with the refrigerant passes; And a first switching valve that switches between guiding the refrigerant vapor compressed by the screw compressor to the heat medium passage and guiding it to the heat source fluid passage. When the heat source device operates to cool the heat medium, the control unit controls the first switching valve so that the refrigerant vapor compressed by the screw compressor is guided to the heat source fluid passage, and when the heat source device operates to heat the heat medium, the refrigerant vapor compressed by the screw compressor is guided to the heat medium passage. When the heat source device operates to cool the heat medium, the heat medium passage functions as the evaporation section and the heat source fluid passage functions as the condensation section. When the heat source device operates to heat the heat medium, the heat medium passage device functions as the condensation part and the heat source fluid passage device functions as the evaporation part. The heat source device according to any one of claims 1 to 7.

9. It is provided with a second switching valve for switching between the heat medium passing through the evaporation part and the heat medium passing through the condensation part. When the heat source device operates to cool the heat medium, the control unit causes the heat medium to pass through the evaporation part, and when the heat source device operates to heat the heat medium, the heat medium passes through the condensation part. The second switching valve is controlled. The heat source device according to any one of claims 1 to 7.

10. An inflow heat medium temperature detector for detecting the temperature of the heat medium flowing into the heat source device or a physical quantity correlated therewith, A heat medium flow rate detector for detecting the flow rate of the heat medium passing through the heat source device or a physical quantity correlated therewith, The heat source device output detector is configured to obtain the output of the heat source device or a physical quantity correlated therewith by multiplying the difference between the value detected by the inflow heat medium temperature detector and the value detected by the outflow heat medium temperature detector by the value detected by the heat medium flow rate detector. The heat source device according to any one of claims 1 to 7.

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