JP2009063267A - Ground heat exchanger and its using method, and ground heat utilizing system and its operating method - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve comprehensive efficiency in a ground heat utilizing system. <P>SOLUTION: The ground heat utilizing system is provided with a plurality of the ground heat exchangers 100 buried in the ground, a heat pump 200 for carrying out heat collection and radiation via a heating medium circulated in each ground heat exchanger 100, an air conditioner 300 heating or cooling a room via a load side heating medium heated or cooled by the heat pump 200, a circulating pump 400 for circulating the heating medium in each ground heat exchanger 100, and a control panel 500 carrying out current transformation amount control of the circulating pump 400. The ground heat exchanger 100 is composed of a hollow pipe body 1, and plural sets of U-tubes 2 housed in the hollow pipe body 1. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention uses ground heat of the earth as a heat source, circulates a heat medium in the ground heat exchanger and collects and dissipates the heat, and a ground heat exchanger that supplies hot or cold to the load side and use thereof The present invention relates to a method, a geothermal heat utilization system, and an operation method thereof.

  Underground equipped with multiple underground heat exchangers buried in the ground, a heat pump for collecting and radiating heat through a heat medium that is circulated through the various regional heat exchangers, and a load machine connected to the heat pump A heat utilization system is known (see Patent Documents 1 to 3).

Such a geothermal heat utilization system uses the ground heat of the earth having a stable temperature as a heat source, collects and radiates heat from this heat source, and uses an air heat source system that collects and radiates heat from outside air. Compared with, it is possible to operate with high efficiency for both cooling and heating by using a stable underground temperature with little change throughout the year. Therefore, in the geothermal heat utilization system, in addition to effects such as energy saving, lower running cost, and carbon dioxide (CO ₂ ) emission suppression, the heat island phenomenon is also expected to be suppressed by not exhausting heat to the atmosphere. .

JP 2004-233031 A JP 2006-292313 A JP 2007-85675 A

  In this type of geothermal heat utilization system, as disclosed in Patent Documents 1 and 3, for example, it is proposed to use a foundation pile of a building or a civil engineering structure for a geothermal heat exchanger. A heat collecting and radiating pipe such as a U-shaped pipe is inserted into a hollow pipe body such as a steel pipe pile used as a foundation pile. In this underground heat exchanger, a heat medium (for example, an antifreeze containing ethylene glycol) flowing through the U-shaped tube exchanges heat with the soil via the tube wall and carries the heat. The amount of heat exchange at this time depends on the thermal characteristics of the ground and the underground heat exchanger, but is also greatly affected by the state of the heat medium flow.

  Here, there are a regular and orderly flow (laminar flow) and a temporally and spatially irregular flow (turbulent flow) in the state of the flow of fluid flowing in the pipe. Turbulent flow occurs when the flow velocity is high or the viscosity is low, and laminar flow is the opposite. The limit of transition of the flow in the tube from laminar flow to turbulent flow is given by the inner diameter d of the circular tube, the cross-sectional average flow velocity U (= flow rate / cross-sectional area), and the Reynolds number Re = Ud / ν using the kinematic viscosity coefficient ν. It is known. In the case of a flow in a circular pipe, for example, if the Ray nozzle number Re ≦ 2300, it may be determined as laminar flow, and if the Reynolds number Re> 2300, it may be determined as turbulent flow, but the transition from the actual laminar flow state to the turbulent flow state may occur. Occurs within a certain width of the number of lay nozzles Re and strongly depends on the shape of the inlet of the pipeline. When the pipe wall is smooth and a constriction is provided at the inlet and the flow path is gradually narrowed to smoothly guide the flow to the pipe, the Reynolds number Re for maintaining laminar flow increases, but the pipe inlet In a normal inlet shape where the angle is cut at a right angle, turbulence starts to occur from the number of lay nozzles Re = 2000, and depending on the flow path shape, a turbulent flow state substantially continuous when the number of lay nozzles Re = 2about 3,000 is exceeded. Become. That is, in the case of a flow in a circular pipe, in theory, it is generally handled as a laminar flow when Re <2000, a transition flow when 2000 ≦ Re <2700 to 3000, and a turbulent flow when Re ≧ 2700 to 3000. It is.

  Heat transfer between the tube wall and the heat medium in the underground heat exchanger decreases when the flow state of the fluid heat medium transitions from turbulent flow to laminar flow, and the heat exchange performance of the underground heat exchanger decreases significantly. It is known. Therefore, the flow rate of the heat medium flowing through the U-shaped tube is required to be a flow rate that can maintain a turbulent state, and is theoretically set to a constant flow rate (constant flow rate) that can maintain a turbulent state. It is.

  On the other hand, a heat pump and a load machine connected to the heat pump, such as an air conditioner that heats or cools a room via a load-side heat medium heated or cooled by the heat pump, a ground heat exchanger, and a circulation pump The capacity of the heat utilization system is set according to the maximum load at the peak of summer and winter. However, the maximum load is generated only during a certain period of time before and after the peak, and considering the entire period, the period of operation with partial load is overwhelmingly longer. Even in a geothermal heat utilization system, there are cases where a restraint operation is performed for energy saving, such as a reduction in the number of compressor operation of a heat pump at a partial load, but as described above, when the flow velocity becomes small, the flow transitions to a laminar flow and the heat exchange performance is reduced. Therefore, it is general that only the circulation pump that conveys and circulates the heat medium is operated at the rated flow rate (always operated at a capacity of 100%) so as to maintain the flow rate.

  In this way, the circulation pump is operated at the rated flow rate even at partial loads, so the good heat source efficiency, which is an advantage of the geothermal heat utilization system, is lost due to the extra conveyance power of the circulation pump, and the overall efficiency is impaired. Currently. It is conceivable to control the circulation pump to change the flow rate and switch the number of underground heat exchangers and the number of piping systems. However, the control for opening and closing the valves is complicated, and a specific underground heat exchanger is also required. In order to prevent the bias toward equality, it is necessary to consider equal usage control.

  This invention is made | formed in view of the above points, and it aims at improving the total efficiency in an underground use system.

The underground heat exchanger of the present invention is a underground heat exchanger that circulates a heat medium in a heat collecting and radiating pipe, and has a flow velocity that can maintain a turbulent state theoretically when the heat medium is circulated in the heat collecting and radiating pipe. In other words, even at low flow rates that are theoretically in a laminar flow state, there is no significant decrease in heat exchange performance that should be in a laminar flow state, and the correlation between the flow rate or flow rate in that region and the heat exchange performance It is characterized by a low flow rate that is lower than the flow rate that can theoretically maintain a turbulent state.
The ground heat utilization system of the present invention is connected to a ground heat exchanger buried in the ground, a heat pump for collecting and radiating heat through a heat medium circulated in the ground heat exchanger, and the heat pump. A load pump, a circulation pump for circulating a heat medium in the underground heat exchanger, and a control means for controlling the flow rate of the circulation pump.
The use method of the underground heat exchanger of the present invention is a use method of the underground heat exchanger in which the heat medium is circulated in the heat collecting / radiating pipe, and theoretically turbulent when circulating the heat medium in the heat collecting / radiating pipe. There is no significant decrease in the heat exchange performance that should be in a laminar flow state even at a flow rate that is lower than the flow rate at which the state can be maintained, that is, in a theoretical laminar flow state, and It is characterized by a low flow velocity that is lower than the flow velocity that can theoretically maintain a turbulent state after grasping.
The operation method of the ground heat utilization system of the present invention includes a ground heat exchanger embedded in the ground, a heat pump for collecting and radiating heat through a heat medium circulated in the ground heat exchanger, and the heat pump. A ground heat utilization system comprising a load machine connected to the ground heat exchanger and a circulation pump for circulating a heat medium in the underground heat exchanger, wherein the circulation pump is subjected to variable flow rate control. Features.

  According to the present invention, it is possible to improve the overall efficiency of the underground use system by not damaging the good heat source efficiency, which is an advantage of the underground heat use system, with the extra conveyance power of the circulation pump.

Preferred embodiments of the present invention will be described below with reference to the accompanying drawings.
First, the basic concept of the present invention will be described. It is conceivable to reduce the flow rate of the heat medium in order to reduce the conveyance power by the circulation pump, but in that case, the heat exchange performance does not deteriorate even if the flow rate is reduced (the heat exchange performance can be maintained). A vessel is required.

  Generally used underground heat exchangers have holes (bore holes) with a diameter of about 100 mm and a depth of about 50-100 m in the ground, and pipes with a diameter (diameter) of about 25 mm are connected via a U-shaped joint. One set or two sets of heat collecting and radiating pipes are inserted into the borehole. In order to reduce soil excavation costs, it is not preferable to make the hole diameter of the borehole extremely larger than about 100 mm. For spatial connection with the heat extraction / radiation tube, the shape of the heat extraction / radiation tube is modified or the number of groups is increased. There is no room for ingenuity. Therefore, in order to maintain heat exchange performance, it is necessary to maintain a flow rate that can theoretically maintain a turbulent state as described above, and it is difficult to reduce the flow rate.

  On the other hand, when using foundation piles of buildings and civil engineering structures for underground heat exchangers, there are various pile construction methods, but there are also pile diameters of about 2 m, and generally there are boreholes. It becomes larger than the hole diameter. Therefore, when inserting a heat-dissipating pipe into a hollow pipe body such as a steel pipe pile used as a foundation pile, there is room for improvement in both the shape and the number of sets of the heat-dissipating pipe.

  Below, the correspondence example of the flow reduction of the underground heat exchanger using the foundation pile of a building or a civil engineering structure is demonstrated. As shown in FIGS. 1 (a) and 1 (b), the underground heat exchanger 100 has a flexible thin-diameter tube such as a cross-linked polyethylene in a hollow tube 1 such as a steel pipe pile used as a foundation pile. Three or more sets (5 sets in the illustrated example) of heat-dissipating tubes (U-tubes) 2 formed by bending the tube itself into a U shape are accommodated (hereinafter referred to as “small multiple tube system”).

  As a comparative example, although not shown in the drawings, an underground heat exchanger that accommodates two sets of heat-dissipating tubes formed by connecting a tube with a diameter (diameter) of 25 mm via a U-shaped joint to the hollow tube body. Think (hereinafter referred to as “standard method”).

  First, consider preparing a number of sets that can secure the same surface area per unit length as the standard-type heat collecting and radiating pipes in the small-diameter multi-tube system, and obtain the same heat exchange performance with a small flow rate. For example, if the U tube 2 has a diameter of 16 mm, the surface area per unit length can be ensured by three sets, and if the U tube 2 has a diameter of 13 mm, the number of sets is four. In addition, since the total cross-sectional area can be reduced to about 50 to 60%, the flow rate can be increased with the same flow rate. Thereby, the theoretical turbulent limit can be reduced, and the heat exchange performance equivalent to that of the standard system can be obtained at a low flow rate. The flow rate [L / min] is a flow rate per set of the hollow tube 1 or a single tube flow rate of the heat collecting and radiating tube.

FIG. 2 shows the correlation between the flow rate of the underground heat exchanger and the heat exchange performance when three sets of U tubes 2 having a diameter of 16 mm are used. The characteristic line 101 is an experimental result (experimental value) of the small-diameter multi-tube method, and the characteristic line 102 is a theoretical value that is theoretically obtained (by Reynolds number) in the small-diameter multi-tube method. The flow rate R ₃ is a theoretical turbulent limit, and theoretically, the heat exchange performance decreases only slightly until reaching this point, but if it exceeds this point, it significantly decreases.

However, in the small-diameter multi-tube system, the actual turbulent flow limit is not observed until at least the flow rate R ₁ . Therefore, even if it is below the theoretical turbulent limit, it is possible to reduce the flow rate to the flow rate R ₁ whose heat exchange performance has been grasped through experiments. Note that the theoretical turbulent limit value in FIG. 2 is a value obtained when the Reynolds number Re = 3000, but the Reynolds number at the flow rate R ₁ is smaller than Re = 2700, and Re = 2700. Even when the theoretical turbulent limit value is obtained, the same result is obtained.

A characteristic line 103 is a theoretical value that is theoretically obtained (by the Reynolds number) by a standard method. Up to now, the flow rate (standard flow rate) R ₅ or more that allows for adjustment error, performance degradation, capability margin (ie, safety factor) to the theoretical flow rate limit flow rate (limit flow rate) R ₄ Although it was ensured, it can be seen that the heat exchange performance equivalent to the flow rate R ₂ smaller than the standard flow rate R ₅ can be obtained with the small-diameter multi-tube method.

In the illustrated example, the correlation with the heat exchange performance is grasped only up to the flow rate R ₁ , but even if the flow rate is less than that, if the heat exchange performance does not decrease as much as the theoretical value (substantially) (As long as a stable turbulent state can be maintained), it is possible to further reduce the flow rate in that region.

  The above-mentioned surface area securing type small-diameter multi-tube system was thought of on the premise that the flow rate could be reduced while maintaining the theoretical turbulent state, although the number of groups was limited, although many. In the experimental values, even if the theoretical turbulent limit is greatly exceeded, it does not actually transition to the laminar flow state, and it is possible to cope with a significant reduction in the flow rate. In other words, even if the number of sets is further increased and the surface area per unit length is increased as compared to the standard type heat collecting and radiating tube, as long as a substantial turbulent state can be maintained, the heat exchange performance is improved. It is considered that the flow rate can be further reduced.

  FIG. 3 shows the correlation between the flow rate of the underground heat exchanger and the heat exchange performance when the number of U tubes 2 having a diameter of 16 mm is five. In this case, the surface area per unit length is increased by about 60% and the total cross-sectional area can be made equal as compared with the standard type heat collecting and radiating tube, so that the flow rate is equivalent if the flow rate is the same. . The characteristic line 201 is an experimental result (experimental value) of the small-diameter multi-tube method, and the characteristic line 202 is a theoretical value theoretically obtained (by the Reynolds number) in the small-diameter multi-tube method. Further, the characteristic line 203 is a theoretical value that is theoretically obtained (by the Reynolds number) by the standard method.

In this case, theoretically, the turbulent flow limit R ₁₅ of the small-diameter multi-tube system is larger than the turbulent flow limit R ₁₃ of the standard system, but the actual turbulent flow is the same as the surface area securing type shown in FIG. No limit is seen at least until flow rate R ₁₁ . Therefore, even if it is below the theoretical turbulent limit, it is possible to reduce the flow rate up to the flow rate R ₁₁ whose heat exchange performance has been grasped by experiments.

It should be noted that the characteristic line 201 is gentler than the surface area ensuring type characteristic line 101 shown in FIG. Accordingly, the flow rate R ₁₂ is smaller than the flow rate R ₂ on the characteristic line 101, and is equivalent to the flow rate (standard flow rate) R ₁₄ that allows for a safety factor in the flow rate (limit flow rate) R ₁₃ that is the theoretical turbulent limit. Heat exchange performance is obtained. That is, it can be seen that further lower flow rate and lower flow rate are possible. When the flow rate at which heat exchange performance equivalent to the standard flow rate is obtained, the surface area securing type shown in FIG. 2 has a small flow rate of about 60 to 70% with respect to the standard flow rate R _5, whereas FIG. the surface area increasing type shown, can be a low flow rate of about 50% relative to the standard flow rate R _14.

Note that the Reynolds number at the flow rate R ₁₂ is Re = 1400 or less, which is much lower than the theoretical limit Re = 2000 in the transition flow region until reaching a complete laminar flow.

  Thus, the correlation between the underground heat exchanger flow rate and the heat exchange performance differs depending on the specifications of the underground heat exchanger. Experiments may be performed each time the specifications of the underground heat exchanger are different, but the analysis based on the experiment in one specification shows that the flow rate and heat exchange performance of the underground heat exchanger in another specification You may make it grasp | ascertain a correlation. As an analysis method, for example, Patent Document 3 discloses an analysis method based on an experimental result, and it may be used.

  As described above, when the heat medium is circulated through the U tube 2, the heat exchange performance should be in a laminar flow state even at a flow rate that is theoretically less than the flow rate at which a turbulent state can be maintained, that is, at a low flow rate in a theoretical laminar state. After grasping the fact that there is no significant decrease and the correlation between the flow rate and the heat exchange performance in that region, the flow rate is set to a low flow rate that is lower than the flow rate that can theoretically maintain the turbulent state.

  The above-described multi-tube method with small diameters makes it possible to reduce the flow rate, that is, maintain the same heat exchange performance with a small flow rate compared to the standard method. However, when the heat pumped up by the underground heat exchanger is carried into a heat pump that is a heat utilization device, there is a problem that the operation performance of the heat pump itself is lowered due to a small flow rate.

  Table 1 shows heat pump power consumption (heat source power consumption) and circulation pump power consumption (heating source power consumption) for each heating load factor when the circulation pump is operated at a constant flow rate that is approximately half the standard (rated) flow rate of a heat pump. An example of calculation of pump power consumption) is shown below.

  At the time of rating, the reduction in the power consumption (ie, conveyance power) of the circulation pump due to the reduction in flow rate is eroded by the increase in the efficiency of the heat pump (ie, heat source power), the total power consumption does not change, and the effect cannot be obtained. I understand that. In this case, it can be seen that reducing the flow rate in the heat medium constant flow operation, that is, reducing the flow rate in the full time, is less effective.

  However, when the partial load is 30% or 50%, the load factor is multiplied by the increase in the heat source power, and the deterioration rate is reduced. On the other hand, the reduction effect of the power consumption of the circulation pump can be obtained as it is. It can be seen that the more effective the driving. As described above, an air conditioner that heats or cools a room has an extremely large effect because the period of operation with a partial load is overwhelmingly long-term.

  Furthermore, Table 2 shows a calculation example when the circulation pump is inverter-controlled at a flow rate corresponding to the load factor at which the heat pump is operated sequentially. When a circulation pump with a capacity that matches the rated (standard) flow rate is selected and the flow rate is reduced to a flow rate that is suitable for partial load, the power consumption is proportional to the number of revolutions and the cube of the flow rate. It can be seen that the reduction can be further achieved. That is, by adding variable flow control, the effect of reducing the flow rate becomes very large.

  FIG. 4 is a graph showing an example of an air conditioning load fluctuation pattern in the cooling period of 6 months. This is based on the monthly load distribution ratio and the distribution ratio by time in the warm district office building shown in the simple air conditioning heat source study program published on the website of the heat pump and heat storage center website. It represents the load factor with respect to the peak load for each hour. It can be seen that in the initial two months and the final one month, the partial load operation is 50 to 70% even at the peak time. In this example, the data is representative data for each month in the same month, but in reality, the daily load pattern gradually increases and decreases.

  FIG. 5 shows the hourly data of FIG. 4 further rearranged in descending order. It can be seen that the time of partial load operation of 40 to 60% is the longest, and the partial load operation of 60% or less accounts for more than half of the operation time.

  From the graphs of FIGS. 4 and 5 and Table 1, it can be seen that at the time of partial load, it is very effective to operate the circulation pump in accordance with the circulation flow rate corresponding to the load.

  Table 3 summarizes the load factor average value, the operation time, and the total load equivalent time for each operation load factor in the descending order rearrangement data of FIG.

  This data is used to estimate the effect of variable flow rate control. The rated capacity of the heat pump in the trial calculation model is 560 kW, the cooling period COP (performance coefficient = “heat pump output ÷ heat pump power consumption”) = 6.0, and the circulation pump rated power consumption is 15 kW. From Table 3, the cooling pump output during the cooling period is 560 kW × 958 h = 536,480 kWh.

  First, as in the past, the overall efficiency when the heat medium circulation flow rate is operated at a constant flow rate, that is, the system performance coefficient SCOP (= “heat pump output / (heat pump power consumption + circulation pump power consumption)”) is obtained. Since COP = 6.0, the power is 536,480 kWh ÷ 6.0 = 89,413 kWh The circulation pump power consumption continues to consume the rated power consumption, so 15 kW × 1,722 h = 25,830 kWh Therefore, the total power consumption is 89,413 kWh + 25,830 kWh = 115,243 kWh, and the SCOP in the constant flow operation is 536,480 kWh ÷ 115,243 kWh = 4.7.

  Next, the total efficiency, that is, the system performance coefficient SCOP when the heat medium circulation flow rate is operated at a flow rate (variable flow rate control) commensurate with the load is obtained. The circulation pump power consumption does not continue to consume the rated power consumption during the operation time, but consumes power corresponding to the flow rate corresponding to the load factor, as shown in Table 4. On the other hand, as shown in Table 4, the COP of the heat pump deteriorates due to a decrease in the circulation flow rate, and the heat source power consumption increases compared to the constant flow operation, but the circulation pump power consumption reduction is superior to this increase. As shown in Table 4, the power consumption is 105,478 kWh, which can be reduced by about 10% compared to the case of operating at a standard flow rate and a constant flow rate. And SCOP in variable flow operation can be improved to 536,480 kWh ÷ 105,478 kWh = 5.1.

  FIG. 6 shows a schematic configuration of a ground heat utilization system to which the present invention is applied. A plurality of underground heat exchangers 100 buried in the ground, a heat pump 200 for extracting and radiating heat through a heat medium that is circulated through the various local heat exchangers 100, and a load side heated or cooled by the heat pump 200 An air conditioner 300 that heats or cools the room via a heat medium, and a circulation pump 400 that circulates the heat medium in the heat exchangers 100 in various places are provided. Although simply illustrated in FIG. 6, the underground heat exchanger 100 is based on the above-described small-diameter multi-tube system, and is accommodated in the hollow tube 1 and the hollow tube 1. Preferably, three or more sets of U tubes 2 are used.

  The control panel 500 controls the flow rate of the circulation pump 400 based on the load output of the heat pump 200. The load output is a load output signal that is actually processed by the heat pump 200. For example, the heat pump 200 receives an operation status (load factor) signal for each load-side individual air conditioner 300 via the control wiring, and operates with the integrated value. Therefore, the load output at which the heat pump 300 is operating is received as an electrical signal, and the circulation pump 400 is controlled to change the flow rate accordingly.

In addition, when the circulation pump 400 is subjected to variable flow control, a lower limit value of the heat medium circulation flow rate is set. As described above, this lower limit value is based on the correlation between the flow rate obtained through experiments or analysis based on experiments and the heat exchange performance regardless of whether or not the flow rate can theoretically maintain a turbulent state. For example, the flow rates R ₁ and R _{11 in} FIGS. 2 and 3 correspond to this.

  In the example of FIG. 7, the temperature of the load-side heat medium (cold / warm water) circulated in the air conditioner 300 is measured by the temperature sensor 700. The control panel 500 controls the flow rate of the circulation pump 400 based on the temperature of the load-side heat medium measured by the temperature sensor 700. If the load increases, the temperature of the load-side heat medium decreases during heating (the reverse occurs during cooling), which is used as a simple method for determining the load.

  In the example of FIG. 8, the temperature of the load-side heat medium (cold / warm water) on the outlet side and the inlet side of the air conditioner 300 is measured by the temperature sensor 700 and the circulation amount is measured by the flow sensor 800. The control panel 500 calculates a load output (flow rate × temperature difference) using the temperature of the load-side heat medium measured by the temperature sensor 700 and the circulation amount measured by the flow sensor 800, and the circulation pump based on the load. 400 is subjected to variable flow control.

  In the example of FIG. 9, the temperature of the heat medium on the outlet side and the inlet side on the heat source side of the heat pump 200 is measured by the temperature sensor 900. The control panel 500 controls the flow rate of the circulation pump 400 so that the temperature difference becomes constant based on the temperature difference of the heat medium on the outlet side and the inlet side of the heat source 200 measured by the temperature sensor 900. To do.

  6 and 9, the load-side circulation pump 600 is indicated by phantom lines because the circulation pump 600 may be unnecessary. For example, as disclosed in Patent Document 2, a refrigerant (gas) is used as a load-side heat medium, the heat pump 200 and the air conditioner 300 are connected by a refrigerant pipe, and the secondary condenser / evaporator is connected to the air conditioner 300. By providing in, the circulation pump 600 by the side of a load can be made unnecessary.

  10 and 11 show another example of a small-diameter multi-tube type underground heat exchanger. In FIG. 1, three or more sets of U tubes 2 are arranged in parallel and accommodated in the hollow tube 1. However, as shown in FIG. The bottoms may be stacked and accommodated so that they are coaxially arranged.

  Further, as shown in FIG. 11, a heat-radiating tube (spiral tube) 3 is formed by spiraling a flexible small-diameter tube, and three or more sets (three sets in the illustrated example) of the spiral tubes 3 are used. May be arranged one above the other. As shown in the figure, the heat-dissipating tube goes all the way to the bottom of the hollow tube, but actually it is folded back from the bottom and rises up to the top of the hollow tube inside or outside the spiral. This is omitted.

  Also, as shown in FIG. 12, a ladder-type heat collecting / dissipating structure in which a large number of small-diameter tubes 4b are arranged at an appropriate pitch between a pair of header pipes 4a (a pair in this example) extending vertically. A tube 4 may be used. In this case, if only one of the pair of header pipes 4a is used as a supply pipe and the other is used as a discharge pipe, the number of the small-diameter pipes 4b arranged in parallel is so large that laminar flow is likely to occur. Therefore, a partition 4c is appropriately installed in the header pipe 4a, and as shown by an arrow in the figure, the heat medium flows from one header pipe 4a to the other header pipe 4a via the plurality of small diameter tubes 4b, and below that Then, it is made to flow alternately in units of a plurality of small diameter pipes 4b such as flowing from the other header pipe 4a to one header pipe 4a via a plurality of small diameter pipes 4b (referred to as S-shaped flow). Is desirable. In addition, as shown in FIGS. 12B and 12C, only one set of the heat-radiating tubes 4 may be accommodated in the hollow tube 1 or when the diameter of the hollow tube 1 is large. As shown in FIG. 12 (d), two sets of heat collecting and radiating tubes 4 may be arranged and accommodated so as to face each other.

  Although the concept is different from the small-diameter multi-tube method described so far, as shown in FIGS. 13A and 13B, for example, a corrosion-resistant resin chain, a torsion bar, It is also effective to insert a turbulent flow inducing material 52 such as a screw rod and forcibly induce turbulent flow in the heat collecting and radiating pipe 51.

  As mentioned above, although this invention was demonstrated with embodiment, this invention is not limited only to these embodiment, A change etc. are possible within the scope of the present invention. For example, in the above-described embodiment, the underground heat exchanger has been described as having a structure in which the heat-dissipating tube is accommodated in the hollow tube 1. However, for example, a heat-dissipating tube (U tube 2, spiral tube 3, ladder in concrete piles). It may have a structure in which a mold heat-radiating pipe 4 or the like is embedded.

It is a figure which shows the structure of an underground heat exchanger of a small diameter multi-tube system. It is a characteristic view which shows the correlation of a underground heat exchanger flow volume and heat exchange performance. It is a characteristic view which shows the correlation of a underground heat exchanger flow volume and heat exchange performance. It is a graph which shows an example of an air-conditioning load fluctuation pattern. It is the graph which rearranged the characteristic view of FIG. It is a figure which shows schematic structure of the underground heat utilization system to which this invention is applied. It is a figure which shows schematic structure of the geothermal heat utilization system of the other example to which this invention is applied. It is a figure which shows schematic structure of the geothermal heat utilization system of the other example to which this invention is applied. It is a figure which shows schematic structure of the geothermal heat utilization system of the other example to which this invention is applied. It is a figure which shows the structure of the underground heat exchanger of the other example of a small diameter multi-tube system. It is a figure which shows the structure of the underground heat exchanger of the other example of a small diameter multi-tube system. It is a figure which shows the structure of the underground heat exchanger of the other example of a small diameter multi-tube system. It is a figure which shows the structure of the underground heat exchanger of the system which induces a turbulent flow.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Hollow tube 2 U tube 3 Spiral tube 4 Ladder type heat-radiating pipe 4a Header pipe 4b Tube 4c Partition 100 Underground heat exchanger 200 Heat pump 300 Air conditioner 400 Circulation pump 500 Control panel 600 Circulation pump 700 Temperature sensor 800 Flow sensor 900 Temperature sensor

Claims (16)

An underground heat exchanger that circulates a heat medium in the heat collection and radiating pipe,
When the heat medium is circulated through the heat collecting and radiating pipe, the heat exchange performance that should be in a laminar flow state should not be significantly reduced even at a low flow rate that can theoretically maintain a turbulent state, that is, at a low flow rate that is theoretically in a laminar flow state. A ground heat exchanger characterized by having a low flow velocity equal to or lower than the flow velocity that can theoretically maintain a turbulent state after grasping the correlation between the flow rate or flow velocity in that region and the heat exchange performance.   The underground heat exchanger according to claim 1, wherein a correlation between the flow rate or the flow velocity and the heat exchange performance is grasped by an experiment or an analysis based on the experiment.   The underground heat exchanger according to claim 1 or 2, comprising a hollow tube body and a plurality of sets of heat collecting and radiating tubes housed in the hollow tube body.   The underground heat exchanger according to claim 3, wherein the heat collecting and radiating pipes are three or more sets.   It is constituted by a hollow tube body and a ladder-type heat collecting / radiating tube housed in the hollow tube body and constructed by laying a plurality of small-diameter tubes between a pair of header pipes. The underground heat exchanger according to claim 1 or 2.   The underground heat exchanger according to any one of claims 1 to 5, wherein a turbulent flow inducing material forcibly inducing turbulent flow is inserted into the heat collecting and radiating pipe. An underground heat exchanger buried in the ground,
A heat pump for collecting and radiating heat through a heat medium circulating in the underground heat exchanger;
A load machine connected to the heat pump;
A circulation pump for circulating a heat medium in the underground heat exchanger;
A ground heat utilization system comprising control means for controlling the flow rate of the circulation pump.   The geothermal heat utilization system according to claim 7, wherein the control means controls the flow rate of the circulating pump based on a load output of the heat pump.   The geothermal heat utilization system according to claim 8, wherein the load output is a load output signal that is actually processed by the heat pump.   The geothermal heat utilization system according to claim 8, wherein the load output is a load output calculated using a temperature and a circulation flow rate of the load-side heat medium.   The geothermal heat utilization system according to claim 7, wherein the control means controls the flow rate of the circulation pump based on the temperature of a load-side heat medium circulated through the load machine.   The geothermal heat utilization according to claim 7, wherein the control means controls the flow rate of the circulation pump based on a temperature difference between the heat medium on the outlet side and the inlet side on the heat source side of the heat pump. system.   The underground heat exchanger according to any one of claims 1 to 6, wherein the control means controls the flow rate of the circulation pump to a flow rate range where the flow velocity is low or lower than a flow velocity that can theoretically maintain a turbulent state. The geothermal heat utilization system according to any one of claims 7 to 12, wherein:   14. The underground according to claim 7, wherein a lower limit value is set for the heat medium circulation flow rate in a region in which the correlation between the flow rate or the flow velocity and the heat exchange performance is grasped. Heat utilization system. It is a method of using an underground heat exchanger that circulates a heat medium in a heat-dissipating pipe,
When the heat medium is circulated through the heat collecting and radiating pipe, the heat exchange performance that should be in a laminar flow state should not be significantly reduced even at a low flow rate that can theoretically maintain a turbulent state, that is, at a low flow rate that is theoretically in a laminar flow state. And the use method of the underground heat exchanger characterized by setting it as the low flow velocity below the flow velocity which can maintain a turbulent flow state theoretically, after grasping | ascertaining the correlation between the flow volume and the heat exchange performance in the area | region. An underground heat exchanger buried in the ground,
A heat pump for collecting and radiating heat through a heat medium circulating in the underground heat exchanger;
A load machine connected to the heat pump;
A method for operating a geothermal heat utilization system comprising a circulation pump for circulating a heat medium in the geothermal heat exchanger,
A method for operating a geothermal heat utilization system, wherein the circulation pump is subjected to variable flow control.

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