JP2007292374A - Heat source variable flow control device and method - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a heat source variable flow control device and its method capable of reducing an operation cost of the total heat source machine. <P>SOLUTION: When a turbo refrigerating machine and an absorption-type refrigerating machine are simultaneously operated, much load flow rate (more than proportionate flow rate) is distributed to the turbo refrigerating machine. Thus the amount of treated heat in the absorption type refrigerating machine of low operation efficiency is decreased, the amount of treated heat in the turbo refrigerating machine of high operation efficiency is increased, and the operation cost of the total heat source machine can be reduced. When an INV turbo refrigerating machine and the turbo refrigerating machine are simultaneously operated, much load flow rate (more than proportionate flow rate) is distributed to the turbo refrigerating machine. Thus a loading factor of the turbo refrigerating machine of high operation efficiency in maximum load is increased, a loading factor of the INV turbo refrigerating machine of high operation efficiency in partial load is decreased, both of the turbo refrigerating machine and the IVN turbo refrigerating machine can exert their capacities in a state of high operation efficiency, and the operation cost of the total heat source machine can be reduced. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a heat source variable flow rate control device and method for controlling the rotational speed of a pump provided as an auxiliary device for a heat source machine based on the flow rate (load flow rate) of heat source water returned to a return header. .

[Conventional example 1]
FIG. 18 shows an instrumentation diagram of a conventional heat source system (secondary pump type heat source system) (see, for example, Patent Document 1). In the figure, 1-1 and 1-2 are heat source units, 2-1 and 2-2 are primary units provided as auxiliary units in the circulation path of the heat source water to the heat source units 1-1 and 1-2, respectively. Pump 3, a forward header that mixes heat source water from heat source devices 1-1, 1-2, 4 is a feed water line, and 5 is a supply of heat source water sent from the forward header 3 via a forward water line 4. An external load (heat load such as an air conditioner or a fan coil) that receives the electric power, 6 is a return water pipe, and 7 is a valve that adjusts the amount of heat source water supplied to the external load 5 according to the load state.

  Reference numeral 8 denotes a return header in which heat is exchanged in the external load 5 and the heat source water sent through the return water pipe 6 is returned. Reference numeral 9 denotes a bypass pipe that connects the forward header 3 and the return header 8. Reference numeral 10 denotes the forward header 3. A water temperature sensor that measures the temperature of the heat source water from the heat source water to the external load 5 as the water temperature TS, 11 is a return water temperature sensor that measures the temperature of the heat source water returned to the return header 8 as the return water temperature TR, and 12 A flow meter 13 for measuring the flow rate of the heat source water returned to the return header 8 as a load flow rate F, 13 is a control device.

  The forward header 3 is composed of a first forward header 3-1 and a second forward header 3-2, and a heat source from the forward header 3-1 is between the forward headers 3-1 and 3-2. Secondary pumps 14-1 to 14-3 for pumping water to the forward header 3-2 are provided. Further, a valve 15 and a differential pressure gauge 16 are provided between the forward header 3-1 and the forward header 3-2, and the secondary pumps 14-1 to 14-3, the valve 15 and the differential pressure gauge 16 are provided. In this case, a secondary pump control device 17 is provided. The secondary pump control device 17 receives the load flow rate F from the flow meter 12 and controls the number of operating secondary pumps 14-1 to 14-3 and forward headers 3-1 and 3 detected by the differential pressure gauge 16. The opening degree of the valve 15 is controlled so that the differential pressure ΔP between −2 is kept constant.

  In this heat source system, the water fed by the primary pumps 2-1 and 2-2 is converted into heat source water by the heat source devices 1-1 and 1-2, mixed in the forward header 3, and passed through the forward water pipeline 4. To the external load 5. Then, heat is exchanged in the external load 5, returned to the return header 8 through the return water pipe 6, and again pumped by the primary pumps 2-1 and 2-2 and circulates through the above path. For example, when the heat source units 1-1 and 1-2 are refrigerators, the heat source water is cold water and circulates through the above-described path. When the heat source devices 1-1 and 1-2 are heating devices, the heat source water is warm water and circulates through the above-described path.

  In this heat source system, the primary pumps 2-1 and 2-2 are turned on / off (operated / stopped) in conjunction with the heat source devices 1-1 and 1-2. When the heat source devices 1-1 and 1-2 are refrigerators, the heat source devices 1-1 and 1-2 are provided with cooling towers 18-1 and 18-2 as auxiliary devices. The cooling tower 18 (18-1, 18-2) includes a cooling water pump GP and a cooling tower fan GF, and the cooling water generated by the cooling tower fan GF is used as the heat source unit 1 (1-1). , 1-2).

  Further, in this heat source system, the control device 13 calculates F × (from the water flow temperature TS from the water flow temperature sensor 10, the water return temperature TR from the water return temperature sensor 11, and the load flow rate F from the flow meter 12. TR-TS) = Q, the current load heat quantity Q is obtained, and the number of operating heat source devices 1-1 and 1-2 is controlled based on the obtained current load heat quantity Q.

  For example, according to a predetermined operation order table, the control device 13 operates the heat source unit 1-1 of the designated rank 1 until the load heat quantity Q reaches a predetermined value Q1, and the load heat quantity Q has a predetermined value Q1. If exceeded, in addition to the heat source device 1-1, the operation of the heat source device 1-2 of the designated rank 2 is started. During the operation of the heat source units 1-1 and 1-2, the primary pumps 2-1 and 2-2 provided for the heat source units 1-1 and 1-2 are operated at a rated speed, Heat source water at a constant flow rate is recirculated to the heat source units 1-1 and 1-2.

  In this heat source system, the primary pump 2 circulates a constant flow rate of the heat source water to the heat source unit 1 regardless of the load flow rate F supplied to the external load 5, so that the excess heat source water flows through the bypass pipe 9. Waste of conveyance power by the primary pump 2 occurs.

[Conventional example 2]
Therefore, in Patent Document 2, as shown in FIG. 19, the inverters 19-1 and 19-2 are provided in the primary pumps 2-1 and 2-2, and the rotational speeds of the primary pumps 2-1 and 2-2 are set. The controllable configuration. In this heat source system, the control device 13 (13 ′) is a heat source that is currently operating the load flow rate F measured by the flow meter 12 when the heat source devices 1-1 and 1-2 are of equal capacity (equal capacity). The apportioned flow rate Fi is obtained by dividing by the number n of machines 1 to be operated, and the rotation speed of the primary pump 2 provided for the heat source device 1 currently operated is controlled according to the apportioned flow rate Fi.

For example, in FIG. 19, it is assumed that the heat source devices 1-1 and 1-2 are equal capacity heat source devices, and two units are operating. Further, the rated flow rate (rated pump capacity) of the primary pumps 2-1 and 2-2 is 100 m ³ / h, and the load flow rate F measured by the flow meter 12 is 150 m ³ / h.

In this case, the control device 13 divides the load flow rate F by the number of operating heat source devices 1 to obtain a proportional flow rate Fi = 75 m ³ / h, and the flow rate of the heat source water to the heat source devices 1-1 and 1-2 is determined. The primary pump is given an inverter output (INV output) to the inverters 19-1 and 19-2 so that the load factor of the heat source devices 1-1 and 1-2 is the same so as to be 75 m ³ / h. 2-1 and 2-2 are controlled.

Thereby, the sum total of the flow volume of the heat source water output from the heat-source equipment 1-1 in operation | movement, and 1-2 becomes 150 m < ³ > / h, and becomes equal to the measured load flow volume F. FIG. Therefore, the flow rate of the heat source water flowing through the bypass pipe 9 becomes zero, so that the conveyance power is not wasted and energy saving is achieved.

JP 2002-89935 A JP 2004-101104 A

  However, in the heat source system of the above-described Conventional Example 2 (FIG. 19), when the operation efficiency of the heat source devices 1-1 and 1-2 is different, the amount of heat processed by the heat source device with poor operation efficiency increases, and the total heat source device There was a problem that the operating cost of the system increased.

  For example, as shown in FIG. 20, when the heat source device 1-1 is a turbo refrigerator and the heat source device 1-2 is an absorption refrigerator, the coefficient of performance (COP) indicating the operation efficiency of the turbo refrigerator is about 6. On the other hand, the COP of the absorption chiller is about 1, and the heat source unit 1-1 is much more efficient in operation than the heat source unit 1-2. In this case, in the heat source system of Conventional Example 2, the load flow rate is evenly distributed to the heat source units 1-1 and 1-2 so that the load factor is the same regardless of whether the operation efficiency is good or bad. However, the amount of heat treated by the heat source device 1-2 is increased, and the operation cost of the total heat source device is increased.

This problem occurs not only when the heat source devices 1-1 and 1-2 have the same capacity (equal capacity) but also when the heat source apparatuses 1-1 and 1-2 have different capacity (different capacity). For example, as shown in FIG. 21, it is assumed that the capacity ratio between the heat source device 1-1 and the heat source device 1-2 is 2: 1. The rated pump capacity of the primary pump 2-1 is 200 m ³ / h, the rated pump capacity of the primary pump 2-2 is 100 m ³ / h, and the load flow rate F measured by the flow meter 12 is 210 m ³ / h. To do. In this case, the control device 13 ′ distributes the load flow rate F at the capacity ratio between the heat source devices 1-1 and 1-2, the flow rate of the heat source water to the heat source device 1-1 is 140 m ³ / h, and the heat source device 1 The number of revolutions of the primary pumps 2-1 and 2-2 is controlled so that the flow rate of the heat source water to -2 is 70 m ³ / h. As a result, the load flow rate is evenly distributed to the heat source devices 1-1 and 1-2 so that the load factors are the same. Here, for example, if the heat source machine 1-1 is a turbo chiller and the heat source machine 1-2 is an absorption chiller, the amount of heat processed by the heat source machine 1-2 (absorption chiller) with poor operating efficiency increases. The operating cost of the total heat source machine increases.

  In addition, when one of the heat source devices 1-1 and 1-2 is a heat source device with good operation efficiency at partial load and the other is a heat source device with good operation efficiency at maximum load, The problem that the operating cost of a heat source machine increases arises. For example, an inverter-type turbo chiller (hereinafter referred to as an INV turbo chiller) in which an inverter is attached to a compressor generally has a characteristic (see FIG. 22) that operational efficiency at a partial load is improved. . On the other hand, a normal turbo refrigerator (a constant-speed turbo refrigerator (hereinafter simply referred to as a turbo refrigerator)) has a characteristic (see FIG. 23) that the operation efficiency at the maximum load is the best. . Here, for example, as shown in FIG. 24, the heat source unit 1-1 is an INV turbo chiller, the heat source unit 1-2 is a turbo chiller, and the heat source units 1-1, 1- 1 are set to have the same load factor. When the load flow rate F is evenly distributed to 2, the operation efficiency of the heat source device 1-1 is deteriorated when the load factor is high, and the operation efficiency of the heat source device 1-2 is deteriorated when the load factor is low. The operating cost of the heat source machine increases.

  The present invention has been made to solve such problems, and an object of the present invention is to provide a heat source variable flow rate control device and method capable of reducing the operation cost of the total heat source machine. It is in.

  In order to achieve such an object, at least one of the present invention has first to Nth (N ≧ 2) heat source units having different operating efficiencies, and auxiliary devices for the first to Nth heat source units. As for at least 1 unit provided in each unit, the first to Nth pumps capable of controlling the flow rate, the forward header for receiving the heat source water from the first to Nth heat source units, and the water fed from the forward header An external load that receives supply of heat source water, a return header that returns the heat source water heat-exchanged in the external load to the first to Nth heat source devices via the first to Nth pumps, a forward header, and a return header The first to Nth pumps are used in a heat source system including a bypass pipe that communicates with each other and a flow meter that measures the flow rate of the heat source water heat-exchanged in an external load as a load flow rate. Heat source for controlling the flow rate of the heat source water from the first to the Nth heat source machine In the flow rate control device, when at least two of the first to Nth heat source units are operated simultaneously, the flow rate of the heat source water to the heat source unit that operates the load flow rate according to the operation efficiency of the operated heat source unit As shown in FIG.

In the present invention, when at least two of the first to Nth heat source units are operated at the same time, the load flow rate is not equally distributed so that the load factor is the same, but the operation efficiency of the operated heat source unit Will be allocated according to.
For example, when the first and second heat source units having different operating efficiencies are operated simultaneously, the load flow rate is distributed to the heat source unit having the higher operating efficiency than the proportional flow rate according to the capacity of the heat source unit. As a result, the amount of heat treated by the heat source machine with poor operating efficiency is reduced, the amount of heat treated by the heat source machine with good operating efficiency is increased, and the operating cost of the total heat source machine is reduced.
In addition, when the first heat source unit having good operation efficiency at the partial load and the second heat source unit having high operation efficiency at the maximum load are operated simultaneously, the second heat source unit having high operation efficiency at the maximum load. In addition, the load flow rate is distributed more than the prorated flow rate according to the capacity of the heat source machine. As a result, the load factor of the second heat source unit having good operation efficiency at the maximum load is increased, the load factor of the first heat source unit having good operation efficiency at the partial load is decreased, and the first heat source unit and the second heat source unit are reduced. Both heat source machines demonstrate their capabilities with good operating efficiency, reducing the total operating cost of the heat source machines.

  In the present invention, it is preferable to determine whether or not to distribute a large load flow rate to the second heat source unit based on the temperature of the heat dissipating heat medium entering the first heat source unit. For example, the INV turbo chiller is a heat source machine with good operation efficiency at partial load, but when the temperature of the heat dissipation heat medium (for example, cooling water from the cooling tower) increases, the operation efficiency at partial load decreases. There is a tendency that the operating efficiency at the maximum load is the same as that of a turbo refrigerator. Therefore, when an INV turbo chiller is used as the first heat source machine, for example, a threshold value is set for the temperature of the radiant heat medium entering the first heat source machine, and the temperature of the radiant heat medium exceeds the threshold value. The distribution of the load flow rate to the second heat source unit is stopped, and the load flow rate is equally distributed to the first heat source unit and the second heat source unit so that the load factor is the same.

  In addition, when the threshold value with respect to the temperature of the heat-dissipating heat medium is finely defined and distribution of a large load flow rate to the second heat source unit is stopped, the first load factor is set to be the same according to the temperature of the heat-dissipating heat medium. The load flow rate may be evenly distributed to the heat source unit and the second heat source unit, or a large load flow rate may be allocated to the second heat source unit. In addition to INV turbo chillers, there are refrigerators whose operating efficiency varies greatly depending on the temperature of the heat-dissipating heat medium, and the first heat source device is not limited to a heat source device with good operating efficiency at partial load. Further, the heat radiation heat medium from the heat radiating device is not limited to the cooling water (water cooling) from the cooling tower, and the condenser in the heat source unit may be cooled (air cooled) with air. In the present invention, the temperature of the heat dissipation medium may be directly detected. However, the outside air enthalpy is obtained from the outside air temperature and the outside air humidity, and the outside air enthalpy is calculated from the outside air enthalpy. You may make it ask.

  Further, the present invention can be realized not only as a heat source variable flow rate control device but also as a heat source variable flow rate control method. Further, the present invention is not limited to application to a heat source system of a secondary pump system, and can be similarly applied to a heat source system of a primary pump system that does not use a secondary pump. In the present invention, the heat source machine is not limited to a turbo refrigerator, an absorption refrigerator, and an INV turbo refrigerator, and various other types of refrigerators can be used. Further, the heat source device may be a heater, a heat exchanger, or the like, and when a heater or a heat exchanger is used, various types of devices can be used like a refrigerator.

  Moreover, in this invention, the 1st-Nth pump provided as an auxiliary machine in the 1st-Nth heat-source equipment should just be a pump in which at least 1 unit | set can control rotation speed. That is, the first to Nth pumps may include a pump having a constant rotation speed that is not controlled by the inverter. For example, when the first heat source machine is a turbo refrigerator and the second heat source machine is an absorption refrigerator, the first pump provided as an auxiliary machine for the first heat source machine has a constant rotational speed. A pump is used, and a second pump provided as an auxiliary machine for the second heat source machine is a pump capable of controlling the rotation speed. In this case, when only the first heat source machine is operated, the first pump is turned on and heat source water having a constant flow rate is sent to the first heat source machine. Further, during simultaneous operation of the first heat source device and the second heat source device, the rotational speed of the second pump is controlled according to the load flow rate while the first pump is kept on. Even in this case, when the first heat source machine and the second heat source machine are operated simultaneously, a large load flow rate is distributed to the first heat source machine having good operation efficiency, and the total heat source machine operation is performed. Cost is reduced.

  According to the present invention, when at least two of the first to Nth heat source units are operated at the same time, the load flow rate is distributed according to the operation efficiency of the operated heat source units. When the first and second heat source units are operated at the same time, the load flow rate may be distributed to the heat source unit having better operating efficiency than the proportional flow rate according to the capacity of the heat source unit, or at the time of partial load When the first heat source unit with good operating efficiency and the second heat source unit with high operating efficiency at the maximum load are operated simultaneously, the capacity of the heat source unit is added to the second heat source unit with high operating efficiency at the maximum load. It is possible to reduce the total operating cost of the heat source unit by distributing the load flow rate more than the proportional flow rate according to the above.

Hereinafter, the present invention will be described in detail with reference to the drawings.
[Embodiment 1]
FIG. 1 is an instrumentation diagram showing an embodiment of a heat source system including a heat source variable flow rate control device according to the present invention. In this figure, the same reference numerals as those in FIG. 19 denote the same or equivalent components as those described with reference to FIG.

  The basic structure of this heat source system is not different from the conventional heat source system shown in FIG. Also in this embodiment, the inverters 19-1 and 19-2 are provided in the primary pumps 2-1 and 2-2, and an INV output is given from the control device 13 to the inverters 19-1 and 19-2. The rotational speeds of the pumps 2-1 and 2-2 are controlled. The difference between the two lies in the function of the control device 13. Hereinafter, in order to distinguish from the control device 13 ′ in the conventional heat source system, the description will be given assuming that the control device 13 in the present embodiment is the control device 13 </ b> A.

  The control device 13A is realized by hardware including a processor and a storage device, and a program that realizes various functions in cooperation with the hardware. Similarly to the conventional control device 13 ′, the control device 13A is provided with the water temperature TS from the water temperature sensor 10, the return water temperature TR from the return water temperature sensor 11, and the load flow rate F from the flow meter 12. It is done. In addition, as one of the characteristic functions of the control device 13A, the load flow rate F is converted into heat source water to the heat source devices 1-1 and 1-2 according to the operation efficiency of the heat source devices 1-1 and 1-2. It has an optimal load flow rate distribution function that distributes the flow rate.

  In the first embodiment, the heat source unit 1-1 is a turbo refrigerator, and the heat source unit 1-2 is an absorption refrigerator. As described above, the COP of the turbo chiller is about 6, whereas the COP of the absorption chiller is about 1, and the operation efficiency of the heat source 1-1 is much higher than that of the heat source 1-2. Is good.

Further, in the first embodiment, the heat source devices 1-1 and 1-2 are heat source devices of equal capacity (equal capacity), and the rated pump capacities of the primary pumps 2-1 and 2-2 are 100 m ³ / h.
In the first embodiment, it is assumed that a minimum passage flow rate (minimum flow rate) is determined for the heat source devices 1-1 and 1-2 in order to prevent an abnormal stop due to freezing or the like. In this example, it is assumed that the minimum flow rate of the heat source device 1-1 is α1, the minimum flow rate of the heat source device 1-2 is α2, and α1 = α2 = 50 m ³ / h.

  Further, in the first embodiment, the control device 13A has a function of controlling the number of operating heat source units based on the load heat quantity Q, but controls the number of operating heat source units based on the load flow rate F. Also good. Further, instead of the control device 13A, another control device (not shown) may be provided with the operation number control function, or the operator may manually indicate the operation number of the heat source units.

[Optimal load flow distribution]
FIG. 2 shows a flowchart of the optimum load flow rate distribution function of the control device 13A. Hereinafter, the optimum load flow rate distribution function of the control device 13A will be described according to this flowchart. Note that the processing of this flowchart is repeatedly executed at regular intervals.

  In step 101, the control device 13 </ b> A takes in the load flow rate F measured by the flow meter 12. Then, the current operating device (heat source device) is determined (step 102).

[When operating one unit (turbo refrigerator)]
Here, it is assumed that, for example, an operation instruction is issued only to the turbo chiller 1-1 by controlling the number of operating units. In this case, the control device 13A compares the load flow rate F measured in step 101 with the minimum flow rate α1 of the turbo chiller (step 103), and if the load flow rate F is F ≦ α1 (YES in step 103), The flow rate of the heat source water to the turbo refrigerator 1-1 is set to α1 (step 104). Thereby, the INV output is given from the control device 13A to the primary pump 2-1, the rotational speed of the primary pump 2-1 is adjusted, and the minimum flow rate from the primary pump 2-1 to the turbo refrigerator 1-1. The heat source water of α1 is refluxed (points t1 to t2 shown in FIG. 3A).

  On the other hand, if the load flow rate F is F> α1 (NO in step 103), the load flow rate F is set as the flow rate of the heat source water to the turbo refrigerator 1-1 (step 105). Thereby, the INV output is given from the control device 13A to the primary pump 2-1, the rotational speed of the primary pump 2-1 is adjusted, and the load flow rate from the primary pump 2-1 to the turbo refrigerator 1-1. The heat source water of F is recirculated (points t2 to t3 shown in FIG. 3 (a)).

[When operating two units (turbo refrigerator + absorption refrigerator)]
On the other hand, it is assumed that an operation instruction is issued to both the turbo chiller 1-1 and the absorption chiller 1-2 by controlling the number of operating units. In this case, the control device 13A subtracts the total value of the minimum flow rate α1 of the turbo chiller 1-1 and the minimum flow rate α2 of the absorption chiller 1-2 from the load flow rate F measured in step 101 (F− (α1 + α2 )), A variable flow rate ΔF is obtained (step 109).

Then, the minimum flow rate α1 of the turbo chiller 1-1 is added to the variable flow rate ΔF to obtain a comparative flow rate F1 (F1 = ΔF + α1) (step 110), and this comparative flow rate F1 is the rated pump of the primary pump 2-1. Until the capacity exceeds 100 m ³ / h (NO in Step 111), the process proceeds to Steps 112 and 113, and the value obtained by subtracting the minimum flow rate α2 of the absorption chiller 1-2 from the load flow rate F is set as the turbo chiller 1-. The flow rate of the heat source water to 1 and the remainder are the flow rate of the heat source water to the absorption chiller 1-2.

  As a result, in a state where the minimum flow rate α2 to the absorption chiller 1-2 having poor operating efficiency is ensured (points t3 to t4 shown in FIG. 3B), a load is applied to the turbo chiller 1-2 having high operating efficiency. The flow rate F is distributed much (distributed more than the prorated flow rate) (points t3 to t4 shown in FIG. 3A), the amount of heat treated in the absorption chiller 1-2 having poor operating efficiency is reduced, and the operation is performed. The amount of heat processed by the efficient centrifugal chiller 1-1 is increased, and the operation cost of the total heat source unit is reduced.

When the comparative flow rate F1 exceeds 100 m ³ / h, the control device 13A proceeds to steps 114 and 115 in response to YES in step 111, and changes the flow rate of the heat source water to the turbo chiller 1-1 to the primary pump 2-1. Is the rated pump capacity of 100 m ³ / h, and the remainder obtained by subtracting 100 m ³ / h from the load flow rate F is the flow rate of the heat source water to the absorption chiller 1-2.

  Thereby, after the flow rate of the heat source water to the turbo chiller 1-1 reaches the rated pump capacity of the primary pump 2-1, the state is maintained (points t4 to t5 shown in FIG. 3A). ) As the load flow rate F increases, the flow rate of the heat source water to the absorption chiller 1-2 increases (points t4 to t5 shown in FIG. 3B). Even in this case, a large load flow rate F is distributed to the turbo chiller 1-1 having good operation efficiency, and the operation cost of the total heat source machine is reduced.

  In FIG. 3 (c), for reference, the transition (characteristic I) of the pump flow rate from the primary pump 2-1 to the turbo chiller 1-1 and the absorption chiller 1- from the primary pump 2-2. 2 shows the transition of the pump flow rate to 2 (characteristic II) and the transition of the combined pump flow rate (characteristic III) combining the primary pumps 2-1 and 2-2. Also from this figure, it can be seen that a large load flow rate F is distributed to the turbo chiller 1-1 having good operating efficiency during the simultaneous operation of the turbo chiller 1-1 and the absorption chiller 1-2. (T3 to t5 points shown in 3 (c)).

[When operating one unit (absorption type refrigerator)]
In the description of the single unit operation described above, it has been described that the operation instruction is issued only to the turbo chiller 1-1. However, depending on the control logic of the number of units operated, the operator may manually operate the heat source unit. When the number of units is instructed, an operation instruction may be issued only to the absorption chiller 1-2.

  In this case, the control device 13A compares the load flow rate F measured in step 101 with the minimum flow rate α2 of the absorption chiller 1-2 (step 106), and if the load flow rate F is F ≦ α2 (step 106). YES), the flow rate of the heat source water to the absorption refrigerator 1-2 is set to α2 (step 107). On the other hand, if the load flow rate F is F> α2 (NO in step 106), the load flow rate F is set as the flow rate of the heat source water to the absorption chiller 1-2 (step 108).

  FIG. 4 shows the transition of the pump flow rate when the absorption chiller 1-2 is operated as the first heat source machine. Even in such a case, when the turbo chiller 1-1 and the absorption chiller 1-2 are operated at the same time, a large load flow rate F is distributed to the turbo chiller 1-1 having good operating efficiency (FIG. 4C T3 to t5 points shown in FIG.

[Embodiment 2]
In the first embodiment, the heat source device 1-1 is a turbo chiller and the heat source device 1-2 is an absorption chiller. However, in the second embodiment, as shown in FIG. The turbo chiller and the heat source machine 1-2 are turbo chillers (constant speed turbo chillers). As described above, the INV turbo chiller generally has a characteristic that the operation efficiency at the time of partial load is improved. On the other hand, the centrifugal chiller has the characteristic that the operation efficiency at the maximum load is the best.

The heat source units 1-1 and 1-2 are heat source units having equal capacity (equal capacity), and the rated pump capacities of the primary pumps 2-1 and 2-2 are 100 m ³ / h, respectively. Further, it is assumed that the minimum flow rate of the heat source device 1-1 is β1, the minimum flow rate of the heat source device 1-2 is β2, and β1 = β2 = 50 m ³ / h. Further, in order to distinguish from the control device 13A in the first embodiment, the description will be given assuming that the control device 13 in the second embodiment is the control device 13B.

[Optimal load flow distribution]
FIG. 6 shows a flowchart of the optimum load flow rate distribution function of the control device 13B. In step 201, the control device 13 </ b> B takes in the load flow rate F measured by the flow meter 12. Then, the current operating device (heat source device) is determined (step 202).

[When operating one unit (turbo refrigerator)]
Here, it is assumed that, for example, only an INV turbo chiller 1-1 has been instructed to operate by controlling the number of operating units. In this case, the control device 13B compares the load flow rate F measured in step 201 with the minimum flow rate β1 of the INV turbo chiller 1-1 (step 203), and if the load flow rate F is F ≦ β1 (step 203). YES), the flow rate of the heat source water to the INV turbo chiller 1-1 is set to β1 (step 204). As a result, an INV output is given from the control device 13B to the primary pump 2-1, the rotation speed of the primary pump 2-1 is adjusted, and the minimum value from the primary pump 2-1 to the INV turbo chiller 1-1 is reached. Heat source water having a flow rate β1 is refluxed (points t1 to t2 shown in FIG. 7A).

  On the other hand, if the load flow rate F is F> β1 (NO in step 203), the load flow rate F is set as the flow rate of the heat source water to the INV turbo chiller 1-1 (step 205). As a result, an INV output is given from the control device 13B to the primary pump 2-1, the rotational speed of the primary pump 2-1 is adjusted, and a load is applied from the primary pump 2-1 to the INV turbo chiller 1-1. The heat source water at the flow rate F is recirculated (points t2 to t3 shown in FIG. 7A).

[When operating two units (INV turbo refrigerator + turbo refrigerator)]
On the other hand, it is assumed that an operation instruction is issued to both the INV turbo chiller 1-1 and the turbo chiller 1-2 by controlling the number of operating units. In this case, the control device 13B subtracts the total value of the minimum flow rate β1 of the INV turbo chiller 1-1 and the minimum flow rate β2 of the turbo chiller 1-2 from the load flow rate F measured in step 201 (F− (β1 + β2 )), A variable flow rate ΔF is obtained (step 209).

Then, the minimum flow rate β2 of the turbo chiller 1-2 is added to the variable flow rate ΔF to obtain a comparative flow rate F2 (F2 = ΔF + β2) (step 210), and this comparative flow rate F2 is the rated pump of the primary pump 2-2. Until the capacity exceeds 100 m ³ / h (NO in step 211), the process proceeds to steps 212 and 213, and the value obtained by subtracting the minimum flow rate β1 of the INV turbo chiller 1-1 from the load flow rate F is set as the turbo chiller 1−. 2 is the flow rate of the heat source water to 2, and the rest is the flow rate of the heat source water to the INV turbo chiller 1-1.

  As a result, in a state where the minimum flow rate β1 to the INV turbo chiller 1-1 is ensured (points t3 to t4 shown in FIG. 7A), the load flow rate F is largely distributed to the turbo chiller 1-2 (proportional flow rate). (T3 to t4 shown in FIG. 7 (b)), the load factor of the turbo chiller 1-2 having high efficiency at the maximum load is increased, and the efficiency at partial load is high. The load factor of the turbo chiller 1-1 decreases, and both the INV turbo chiller 1-1 and the turbo chiller 1-2 exhibit their capabilities with good operating efficiency, and the total operating cost of the heat source machine is reduced.

When the comparative flow rate F1 exceeds 100 m ³ / h, the control device 13B proceeds to steps 214 and 215 in response to YES in step 211 to change the flow rate of the heat source water to the turbo chiller 1-2 to the primary pump 2-2. Is the rated pump capacity of 100 m ³ / h, and the remainder obtained by subtracting 100 m ³ / h from the load flow rate F is the flow rate of the heat source water to the INV turbo chiller 1-1.

  Thereby, after the flow rate of the heat source water to the turbo chiller 1-2 reaches the rated pump capacity of the primary pump 2-2 (points t4 to t5 shown in FIG. 7B), the load flow rate F increases. Accordingly, the flow rate of the heat source water to the INV turbo chiller 1-1 increases (points t4 to t5 shown in FIG. 7A). Even in this case, a large load flow rate F is allocated to the turbo chiller 1-2 having good operation efficiency at the maximum load, and the operation cost of the total heat source unit is reduced.

  In FIG. 7 (c), for reference, the transition (characteristic I) of the pump flow rate from the primary pump 2-1 to the INV turbo chiller 1-1, and the turbo chiller 1- from the primary pump 2-2. 2 shows the transition of the pump flow rate to 2 (characteristic II) and the transition of the combined pump flow rate (characteristic III) combining the primary pumps 2-1 and 2-2. Also from this figure, when the INV turbo chiller 1-1 and the turbo chiller 1-2 are operated at the same time, a large load flow rate F is allocated to the turbo chiller 1-2 with good operating efficiency at the maximum load. (T3 to t5 shown in FIG. 7C).

[When operating one unit (turbo refrigerator)]
In the description of the single unit operation described above, it has been described that the operation instruction is issued only to the INV turbo chiller 1-1. However, depending on the control logic of the number of units operated, the operator may manually change the heat source unit. When the number of operating units is instructed, an operating instruction may be issued only to the turbo chiller 1-2.

  In this case, the control device 13B compares the load flow rate F measured in step 201 with the minimum flow rate β2 of the turbo refrigerator 1-2 (step 206), and if the load flow rate F is F ≦ β2 (step 206). YES), the flow rate of the heat source water to the turbo refrigerator 1-2 is set to β2 (step 207). On the other hand, if the load flow rate F is F> β2 (NO in step 206), the load flow rate F is set as the flow rate of the heat source water to the turbo refrigerator 1-2 (step 208).

  FIG. 8 shows the transition of the pump flow rate when the turbo chiller 1-2 is operated as the first heat source machine. Even in such a case, when the INV turbo chiller 1-1 and the turbo chiller 1-2 are operated simultaneously, a large load flow rate F is distributed to the turbo chiller 1-2 having a high operating efficiency at the maximum load. (T3 to t5 points shown in FIG. 8C).

[Embodiment 3]
In the first embodiment (FIG. 1), inverters 19-1 and 19-2 are provided in the primary pumps 2-1 and 2-2, and an INV output is given to the inverters 19-2 and 19-2 from the control device 13A. The rotation speed of the primary pumps 2-1 and 2-2 was controlled. On the other hand, in the third embodiment, as shown in FIG. 9, the primary pump 2-1 is a pump having a constant rotational speed that does not use an inverter, and is turned on / off in conjunction with the turbo chiller 1-1. Run / stop). As with the first embodiment, the primary pump 2-2 is a pump capable of controlling the rotation speed.

In order to distinguish from the control device 13A in the first embodiment, the description will be continued assuming that the control device 13 in the third embodiment is the control device 13C. In the third embodiment, it is assumed that the minimum flow rate of the absorption chiller 1-2 is α, and α = 50 m ³ / h.

[Optimal load flow distribution]
FIG. 10 shows a flowchart of the optimum load flow rate distribution function of the control device 13C. In step 301, the control device 13 </ b> C takes in the load flow rate F measured by the flow meter 12. Then, the current operating device (heat source device) is determined (step 302).

[When operating one unit (turbo refrigerator)]
Here, it is assumed that, for example, an operation instruction is issued only to the turbo chiller 1-1 by controlling the number of operating units. In this case, the control device 13C maintains the state as it is. That is, in this case, the turbo chiller 1-1 is operated by controlling the number of operating units, and the primary pump 2-1 is turned on in conjunction with this operation. Thereby, 100 m < ³ > / h of heat source water is recirculated from the primary pump 2-1 to the turbo refrigerator 1-1, and this state is maintained (points t1 to t3 shown in FIG. 11A).

[When operating two units (turbo refrigerator + absorption refrigerator)]
On the other hand, it is assumed that an operation instruction is issued to both the turbo chiller 1-1 and the absorption chiller 1-2 by controlling the number of operating units. In this case, the control device 13C calculates the comparison flow rate F3 as “F3 = 100 m ³ / h + α” (step 306), and the absorption chiller 1 until the load flow rate F exceeds the comparison flow rate F3 (NO in step 307). The flow rate of the heat source water to -2 is α (step 308, t3 to t4 shown in FIG. 11B). When the load flow rate F exceeds the comparison flow rate F3 (YES in step 307), the flow rate of the heat source water to the absorption chiller 1-2 is set to “F-100 m ³ / h” (step 309, FIG. 11B). T4 to t5 points).

  As a result, when two turbo chillers 1-1 and absorption chillers 1-2 are operated, a large load flow rate F is distributed to the turbo chillers 1-1 (t3 shown in FIG. 11C). ~ T5 point), the amount of heat treated in the absorption chiller 1-2 having poor operating efficiency is reduced, the amount of heat treated in the turbo chiller 1-1 having good operating efficiency is increased, and the operating cost of the total heat source machine is reduced. .

[When operating one unit (absorption type refrigerator)]
When the operation instruction is issued only to the absorption chiller 1-2, the control device 13C compares the load flow rate F measured in step 301 with the minimum flow rate α of the absorption chiller 1-2 (step 303). ) If the load flow rate F is F ≦ α (YES in step 303), the flow rate of the heat source water to the absorption refrigerator 1-2 is set to α (step 304). On the other hand, if the load flow rate F is F> α (NO in step 303), the load flow rate F is set as the flow rate of the heat source water to the absorption chiller 1-2 (step 305).

  FIG. 12 shows the transition of the pump flow rate when the absorption chiller 1-2 is operated as the first heat source machine. Even in such a case, when the turbo chiller 1-1 and the absorption chiller 1-2 are operated at the same time, a large load flow rate F is distributed to the turbo chiller 1-1 having good operating efficiency (see FIG. T3 to t5 points shown in 12 (c)).

[Embodiment 4]
In the second embodiment (FIG. 5), the heat source unit 1-1 is an INV turbo chiller with good operating efficiency at partial load, and the heat source unit 1-2 is a turbo chiller with high operating efficiency at maximum load. During simultaneous operation of 1-1 and 1-2, a large load flow rate F is distributed to the heat source unit 1-2 with good operation efficiency at the maximum load.

  As described above, the INV turbo chiller generally has a characteristic that the operation efficiency at the time of partial load is improved. On the other hand, the centrifugal chiller has the characteristic that the operation efficiency at the maximum load is the best. The characteristics of the operating efficiency of this refrigerator vary depending on the temperature of cooling water from the cooling tower entering the refrigerator (cooling water inlet temperature).

  FIG. 13 shows changes in operating efficiency characteristics with the cooling water inlet temperature of the INV turbo refrigerator as a parameter. FIG. 14 shows the relationship between changes in operating efficiency characteristics using the cooling water inlet temperature of the centrifugal chiller as a parameter. As can be seen from the change in characteristics shown in FIG. 13, in the INV turbo chiller, when the cooling water inlet temperature increases, the operation efficiency at the partial load decreases, and the operation efficiency at the maximum load similar to the turbo chiller is good. Tend to be characteristic.

  Therefore, in the fourth embodiment, the cooling water inlet temperature Tcw from the cooling tower 18-1 entering the INV turbo chiller 1-1 is measured, and the INV turbo chiller 1-1, the turbo chiller 1-2, When the cooling water inlet temperature Tcw to the INV turbo chiller 1-1 exceeds a threshold value T0 (for example, T0 = 29 ° C.), the turbo chiller 1-2 has a high operating efficiency at the maximum load. The distribution of the load flow F is stopped, and the load flow F is equally distributed to the INV turbo chillers 1-1 and 1-2 so that the load factor is the same. That is, when Tcw ≦ T0, a large load flow rate F is distributed to the turbo chiller 1-2 having good operating efficiency at the maximum load (turbo chiller priority), and when T0 <Tcw, the turbo chiller The apportioning method in which priority is invalidated and the load flow rate F is evenly distributed is adopted.

  FIG. 15 shows an instrumentation diagram of the heat source system according to the fourth embodiment. This heat source system is the same as the heat source system shown in FIG. 5 in the basic configuration, but the cooling water inlet temperature Tcw from the cooling tower 18-2 entering the INV turbo chiller 1-1 is measured by the thermometer 20. Then, it is given to the control device 13 (13D). Further, the control device 13D has a function of determining whether or not to distribute a large load flow rate F to the turbo chiller 1-2 based on the cooling water inlet temperature Tcw to the INV turbo chiller 1-1. .

  FIG. 16 shows a flowchart of the optimum load flow rate distribution function of the control device 13D. The control device 13D checks whether or not the cooling water inlet temperature Tcw exceeds T0 during simultaneous operation of the INV turbo chiller 1-1 and the turbo chiller 1-2 (step 410). If the cooling water inlet temperature Tcw does not exceed T0 (NO in step 410), the operation efficiency at the maximum load is good by the processing of steps 411 to 417 corresponding to steps 209 to 215 shown in FIG. A large load flow rate F is distributed to the turbo refrigerator 1-2.

  If the cooling water inlet temperature Tcw exceeds T0 (YES in Step 410), it is checked whether or not the load flow rate F is (β1 + β2) or more (Step 418). If F <(β1 + β2), turbo The flow rate of the heat source water to the refrigerator 1-2 is set to β2 (step 419), and the flow rate of the heat source water to the INV turbo refrigerator 1-1 is set to β1 (step 420). If (β1 + β2) ≦ F, the flow rates of the heat source water to the INV turbo chiller 1-1 and the turbo chiller 1-2 are both F / 2 (steps 421 and 422).

  FIG. 17 shows the transition of the pump flow rate when the cooling water inlet temperature Tcw exceeds T0. As can be seen from this figure, in the fourth embodiment, when the cooling water inlet temperature Tcw exceeds T0, the INV turbo chiller 1-1 and the turbo chiller 1-2 are operated in a proportional manner during simultaneous operation. The heat source water is recirculated to the INV turbo chiller 1-1 and the turbo chiller 1-2. As a result, even if the cooling water inlet temperature Tcw increases and the operating efficiency characteristics of the INV turbo chiller 1-1 change, the total operating cost of the heat source machine can be reduced in response to the change in operating efficiency characteristics. It becomes possible to reduce.

  In the fourth embodiment described above, the turbo chiller has priority when Tcw ≦ T0, and the proportional distribution method when Tcw> T0. However, the first threshold value T1 and the second threshold value are used as the cooling water inlet temperature threshold values. T2 may be determined (T1 <T2), and when Tcw ≦ T1, turbo chiller priority may be given, when T1 ≦ Tcw <T2, the proportional distribution method may be used, and when T2 ≦ Tcw, INV turbo chiller priority may be given. In other words, when Tcw> T1, priority is given to the turbo chiller, and a proportional distribution method may be used when T1 ≦ Tcw <T2, and an INV turbo chiller may be given priority when T2 ≦ Tcw. Also in this case, it is still determined whether or not to allocate a large load flow rate F to the turbo chiller 1-2 based on the cooling water inlet temperature Tcw to the INV turbo chiller 1-1.

  In Embodiment 4 described above, the cooling water inlet temperature Tcw from the cooling tower 18 to the heat source unit 1 is measured by the thermometer 20, but the temperature of the outside air and the humidity of the outside air are measured, The outside air enthalpy may be obtained from the measured outside air temperature and the outside air humidity, and the cooling water inlet temperature Tcw may be obtained by calculation from the outside air enthalpy and the capacity of the fan of the cooling tower 18.

  Moreover, in Embodiment 4 mentioned above, although the cooling water from the cooling tower 18 was supplied to the heat source unit 1 (water cooling system (heat dissipation heat medium: water)), the condenser in the heat source unit 1 is air. In some cases, cooling (air cooling method (heat dissipating heat medium: air)) is performed. In the case of the air cooling method, the temperature of the air entering the heat source unit 1 is measured as the cooling air inlet temperature.

  Moreover, in Embodiment 1-4 mentioned above, although the inverter 19 was provided in the pump 2 and the rotation speed of the pump 2 was adjusted with INV output, the flow volume to the heat-source equipment 1 was controlled, For example, pump 2 A valve may be provided in the previous stage, and the flow rate to the heat source unit 1 may be controlled by controlling the valve.

  Moreover, in Embodiment 1-4 mentioned above, although heat source apparatus 1-1 and 1-2 were made into equal capacity (equal capacity), this invention is applied similarly when it is different capacity (different capacity). Is possible. For example, when the heat source devices 1-1 and 1-2 have different capacities in the first embodiment, when the heat source devices 1-1 and 1-2 are operated simultaneously, the heat source device 1-1 has a high operating efficiency. The load flow F is distributed more than the proportional flow according to the capacity of 1-1. When the heat source devices 1-1 and 1-2 have different capacities in the second embodiment, when the heat source devices 1-1 and 1-2 are operated simultaneously, the heat source device 1-2 has high operating efficiency at the maximum load. The load flow rate F is distributed more than the proportional flow rate corresponding to the capacity of the heat source device 1-2.

  Moreover, although Embodiment 1-4 mentioned above was demonstrated as an application example to the heat source system of a secondary pump system, it can apply similarly to the heat source system of a primary pump system which does not use a secondary pump. It is. Moreover, although Embodiment 1-4 mentioned above was made into the system which used 2 heat source units 1 for the simplification of description, it can apply similarly to the system which uses 3 or more heat source units. Is possible. In a system in which three or more heat source units 1 are used, when there are heat source units of the same type, the load flow rate may be apportioned.

Further, the heat source unit 1 is not limited to a turbo refrigerator, an absorption refrigerator, and an INV turbo refrigerator, and various other types of refrigerators may be used. Further, the heat source device 1 may be a heater, a heat exchanger, or the like, and when a heater or a heat exchanger is used, various types of devices can be used like a refrigerator.
Further, as the external load 5, in addition to a heat load such as an air conditioner or a fan coil unit, a consumer of district cooling and heating may be considered. However, when the district heating and cooling customer is an external load, the basic configuration of the heat source system shown in Embodiments 1 to 4 is slightly different.

1 is an instrumentation diagram showing an embodiment (Embodiment 1) of a heat source system including a heat source variable flow rate control device according to the present invention. 4 is a flowchart illustrating an optimum load flow rate distribution function included in the control device in the heat source system according to the first embodiment. It is a figure explaining the optimal load flow distribution according to the flowchart shown in FIG. 2 (when the first heat source machine is a turbo refrigerator). It is a figure explaining the optimal load flow volume distribution according to the flowchart shown in FIG. 2 (when the heat-source machine operated 1st is an absorption refrigerator). It is an instrumentation figure which shows other embodiment (Embodiment 2) of the heat source system containing the heat source variable flow control apparatus which concerns on this invention. 6 is a flowchart illustrating an optimum load flow rate distribution function included in a control device in the heat source system according to the second embodiment. It is a figure explaining the optimal load flow distribution according to the flowchart shown in FIG. 6 (when the first heat source machine is an INV turbo chiller). It is a figure explaining the optimal load flow volume distribution according to the flowchart shown in FIG. 6 (when the heat-source machine operated 1st is a turbo refrigerator). It is an instrumentation figure which shows other embodiment (Embodiment 3) of the heat source system containing the heat source variable flow control apparatus which concerns on this invention. 10 is a flowchart illustrating an optimum load flow distribution function included in a control device in the heat source system according to the third embodiment. It is a figure explaining the optimal load flow distribution according to the flowchart shown in FIG. 10 (when the first heat source machine is a turbo refrigerator). It is a figure explaining the optimal load flow distribution according to the flowchart shown in FIG. 10 (when the first heat source machine is an absorption refrigeration machine). It is a figure which shows the change of the characteristic of the operation efficiency which makes the cooling water inlet temperature of an INV turbo refrigerator a parameter. It is a figure which shows the change of the characteristic of the operation efficiency which sets the cooling water inlet temperature of a turbo refrigerator as a parameter. It is an instrumentation figure which shows other embodiment (Embodiment 4) of the heat source system containing the heat source variable flow control apparatus which concerns on this invention. 10 is a flowchart illustrating an optimum load flow rate distribution function included in a control device in the heat source system according to the fourth embodiment. It is a figure explaining the optimal load flow distribution when the cooling water inlet_port | entrance temperature Tcw to an INV turbo refrigerator has exceeded T0 in the heat source system of Embodiment 4. FIG. It is an instrumentation figure (conventional example 1) of the conventional heat source system. It is an instrumentation figure (conventional example 2) of the conventional heat source system. It is a figure explaining the problem at the time of operating the heat source machine of the equal capacity from which the operation efficiency differs in the proportional distribution method in the heat source system of the prior art example 2. FIG. It is a figure explaining the problem at the time of operating the heat source machine of different ability from which the operation efficiency differs in the proportional distribution method in the heat source system of the prior art example 2. FIG. It is a figure which shows the change (change of operation efficiency) of COP with respect to the load factor of an INV turbo refrigerator. It is a figure which shows the change (change of operation efficiency) of COP with respect to the load factor of a turbo refrigerator. It is a figure explaining the problem at the time of drive | operating an INV turbo refrigerator and a turbo refrigerator simultaneously by the proportional distribution method.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 (1-1, 1-2) ... Heat source machine, 2 (2-1, 2-2) ... Primary pump, 3 (3-1, 3-2) ... Out header, 4 ... Outbound pipe line, 5 DESCRIPTION OF SYMBOLS ... External load, 6 ... Return water line, 7 ... Valve, 8 ... Return header, 9 ... Bypass line, 10 ... Outbound water temperature sensor, 11 ... Return water temperature sensor, 12 ... Flow meter, 13 (13A, 13B, 13C, 13D) ... Control device, 14 (14-1 to 14-3) ... Secondary pump, 15 ... Valve, 16 ... Differential pressure gauge, 17 ... Secondary pump control device, 18 (18-1, 18-2) ... cooling tower, 19 (19-1, 19-2) ... inverter, 20 ... thermometer.

Claims (8)

At least one of the first to Nth (N ≧ 2) heat source units having different operating efficiencies and at least one unit provided as an auxiliary to each of the first to Nth heat source units is a flow rate control. The first to Nth pumps that can perform the operation, the forward header that receives the heat source water from the first to Nth heat source machines, the external load that receives the supply of the heat source water sent from the forward header, and the external A return header that returns the heat source water heat-exchanged in the load to the first to N-th heat source devices via the first to N-th pumps, and a bypass pipe that communicates the forward header and the return header. The heat source system includes a flow meter that measures the flow rate of the heat source water heat-exchanged in the external load as a load flow rate, and the first to Nth pumps from the first to Nth pumps based on the load flow rate. Heat source current control that controls the flow rate of heat source water to the Nth heat source machine In the control device,
When at least two of the first to Nth heat source units are operated at the same time, the load flow rate is distributed as the flow rate of the heat source water to the operated heat source unit according to the operation efficiency of the operated heat source unit. A heat source variable flow rate control device comprising load flow rate distribution means. In the heat source variable flow rate control device according to claim 1,
The heat source system includes:
Comprising first and second heat source units having different operating efficiencies as the first to Nth heat source units;
The load flow distribution means is
When the first and second heat source devices are operated at the same time, the load flow rate is distributed more to the heat source device having better operating efficiency than the proportional flow rate corresponding to the capacity of the heat source device. Heat source variable flow rate control device. In the heat source variable flow rate control device according to claim 1,
The heat source system includes:
The first to Nth heat source units include a first heat source unit having good operation efficiency at partial load and a second heat source unit having good operation efficiency at maximum load,
The load flow distribution means is
When the first and second heat source devices are operated at the same time, the load flow rate is distributed more to the second heat source device with good operating efficiency at the maximum load than the prorated flow rate according to the capacity of the heat source device. A heat source variable flow rate control device. In the heat source variable flow rate control device according to claim 2 or 3,
The load flow distribution means is
A heat source variable flow rate control device that determines whether or not to distribute a large amount of the load flow rate to the second heat source unit based on the temperature of the radiant heat medium entering the first heat source unit. At least one of the first to Nth (N ≧ 2) heat source units having different operating efficiencies and at least one unit provided as an auxiliary to each of the first to Nth heat source units is a flow rate control. The first to Nth pumps that can perform the operation, the forward header that receives the heat source water from the first to Nth heat source machines, the external load that receives the supply of the heat source water sent from the forward header, and the external A return header that returns the heat source water heat-exchanged in the load to the first to N-th heat source devices via the first to N-th pumps, and a bypass pipe that communicates the forward header and the return header. And a heat source system including a flow meter for measuring a flow rate of heat source water heat-exchanged in the external load as a load flow rate, and the first from the first to Nth pumps based on the load flow rate. Heat source current control that controls the flow rate of heat source water to the Nth heat source machine In the control method,
When at least two of the first to Nth heat source units are operated at the same time, the load flow rate is distributed as the flow rate of the heat source water to the operated heat source unit according to the operation efficiency of the operated heat source unit. A heat source variable flow rate control method comprising a load flow rate distribution step. In the heat source variable flow rate control method according to claim 5,
The heat source system includes:
Comprising first and second heat source units having different operating efficiencies as the first to Nth heat source units;
The load flow distribution step includes
When the first and second heat source devices are operated at the same time, the load flow rate is distributed more to the heat source device having better operating efficiency than the proportional flow rate corresponding to the capacity of the heat source device. Heat source variable flow rate control method. In the heat source variable flow rate control method according to claim 5,
The heat source system includes:
The first to Nth heat source units include a first heat source unit having good operation efficiency at partial load and a second heat source unit having good operation efficiency at maximum load,
The load flow distribution step includes
When the first and second heat source devices are operated at the same time, the load flow rate is distributed more to the second heat source device with good operating efficiency at the maximum load than the prorated flow rate according to the capacity of the heat source device. A heat source variable flow rate control method characterized by: In the heat source variable flow rate control method according to claim 6 or 7,
The load flow distribution means is
A heat source variable flow rate control method characterized by determining whether or not to distribute a large amount of the load flow rate to the second heat source unit based on the temperature of the heat radiation heat medium entering the first heat source unit.

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