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Verification of the energy-saving effect of the district heating and cooling system—Simulation of an electric-driven heat pump system
Energy and Buildings 40 (2008) 732–741 www.elsevier.com/locate/enbuild
Verification of the energy-saving effect of the district heating and cooling system—Simulation of an electric-driven heat pump system Tomoji Nagota, Yoshiyuki Shimoda *, Minoru Mizuno Division of Sustainable Energy and Environmental Engineering, Osaka University, 2-1 Yamadaoka, Suita City, Osaka 565-0871, Japan Received 17 March 2007; received in revised form 9 May 2007; accepted 14 May 2007
Abstract A district heating and cooling system (DHC) is expected to be a promising energy-saving measure for high-density business areas in Japan. However, it has not been verified what advantages of the DHC are important for energy conservation. The clarification of this issue is supposed to contribute to improving the energy efficiency of the DHC. This paper focuses on the electric-driven heat-pump-type DHC, which uses only electricity as its energy source. An existing DHC plant has been selected for the case study, and its energy efficiency is examined by a simulation model that uses parameters derived from the measurement data. The simulation results for the plant reveal that the DHC exhibits an energy-saving effect of 29% for cooling when compared with the individual heat source system mainly due to the following two advantages: ‘‘economy of scale in chillers/heat pumps’’ and ‘‘thermal storage effect’’. Further, the energy-saving ratio for heating is only 5% since heat recovery chiller cannot be operated sufficiently due to the lack of cooling demand during winter. In addition, we have examined high-efficiency techniques such as the utilization of river water. As a result, potential energy-saving impact on the DHC is evaluated. # 2007 Elsevier B.V. All rights reserved. Keywords: District heating and cooling system; Simulation; Energy-saving effect; Electric-driven heat pump system
1. Introduction In Japan, the number of district heating and cooling systems (DHC) has constantly increased to over 150 since the first DHC commenced supplying heat in 1970. The DHC is expected to have various social advantages such as air-pollution abatement and intensive use of chillers and machinery space [1,2]. Above all, energy conservation is one of the most important advantages of the DHC with regard to the mitigation of global warming [3–5]. The DHC is classified into two categories according to the type of energy source (‘‘absorption chiller and boiler’’ and ‘‘electric-driven heat pump’’), and the factors of the DHC that influence energy conservation differ between the types of energy sources. The absorption-chiller-and-boiler-type DHC, which uses gas as its main energy source, conserves energy mainly due to the ‘‘co-generation system’’ and ‘‘concentration effect’’ [6,7]. In the electric-driven heatpump-type DHC, which uses electricity as its main energy
* Corresponding author. Tel.: +81 6 6879 7665; fax: +81 6 6879 7665. E-mail address: [email protected] (Y. Shimoda). 0378-7788/$ – see front matter # 2007 Elsevier B.V. All rights reserved. doi:10.1016/j.enbuild.2007.05.007
source, the following factors are considered to mainly contribute to energy saving [2,8–10]. High-efficiency equipment: Centrifugal refrigerators, which usually are not introduced for small-scale heat source system, produce cooling very efficiently. Thermal storage: Chillers/heat pumps can be operated at full load by utilizing the thermal storage capability. In addition, a heat recovery chiller with double-bundle condenser (referred to as a ‘‘heat recovery chiller’’ hereafter) can be introduced since a temporal difference between the heating and cooling loads can be adjusted. This type of DHC plant is expected to exhibit the large energy-saving effect in the place where there is a large amount of cooling demand. In Japan, 21% of the DHC plants adopt electric-driven heat-pump-type. However, a measurement survey on the DHC [11] reveals that the energy efficiency ratio (EER) of each electric-driven heat-pump-type DHC plant is widely distributed, as shown in Fig. 1. EER is defined as EER ¼
total amount of supplied heat ½GJ=year total primary energy consumption ½GJ=year
(1)
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Fig. 1. Distribution of EER in electric-driven heat-pump-type DHC plants adapted from a measurement data survey [11].
This indicates that the energy-saving effect from the abovementioned factors does not appear uniformly at the same level in all the plants. In other words, the potential of the energy-saving effect from some important factors is achieved in high-EER plants; however, it fails to be achieved in low-EER plants due to various issues. Thus far, any quantitative verification of the energy-saving effect of each factor of the electric-driven heat-pump-type DHC has not been conducted, and it is unclear what factors are important for the reduction in energy consumption. The clarification of this issue is considered to provide useful information about not only the modification of the existing plants and the planning of new DHC plants but also the necessity of the DHC in the future. In this study, an existing DHC plant is selected for a case study, and the simulation models of the DHC and an individual system (the case where every building has own heat source system) – both of which supply heat to the same heat-supply area of the case-study plant – are developed in order to perform a factorial analysis of the energy-saving effect of the DHC by comparing
Fig. 2. Flow chart for the verification of the energy-conservation effect of the DHC.
both the simulation results. The simulation parameters are derived from the measurement data of the existing heat source systems for the purpose of improving the accuracy of simulation. Fig. 2 shows a flow chart of this study. In addition, a highefficiency technique, such as the utilization of river water and the improvement in the heat load profile, is examined to reveal the potential of the DHC in improving the energy efficiency. 2. Simulation model An existing electric-driven heat-pump-type DHC plant (APlant), which achieves the second highest EER among the DHC plants in Japan and exhibits the energy-saving effect from advantages of the DHC properly, has been selected for the case study. The A-Plant comprises electric-driven chillers/heat
Fig. 3. System configuration of the A-Plant. The operation mode of cooling/heating tower, heating-tower mode (using brine) and cooling-tower mode (using cooling water), is changed corresponding to the operation mode of the centrifugal heat pump. The number of secondary hot/cold water pumps shown in the figure is different from that of the A-Plant. CR: centrifugal refrigerator, HRC: heat recovery chiller, HP: centrifugal heat pump, PCD: cooling water pump, BMP: brine/cooling water pump, PC: cold water pump, PH: hot water pump, S_PC: secondary cold water pump, S_PH: secondary hot water pump.
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Table 1 Chillers and heat pumps in the A-Plant Equipment
Operating mode
Heating/cooling
Capacity (GJ/h)
CR1/CR2
Centrifugal refrigerator
HRC1/HRC2
Heat recovery chiller
HP1/HP2
Centrifugal heat pump
Cooling mode
Cooling
14.9
5.3
Heat recovery mode
Cooling Heating
5.0 6.8
6.7
Cooling mode Heating mode
Cooling Heating
18.3 12.7
4.7 3.2
pumps and thermal storage tanks, as shown in the system configuration (Fig. 3, Table 1), and it supplies cold water (the annual amount of supplied heat in fiscal 2002: 29.6 GWh) and hot water (11.6 GWh) to five buildings (floor areas: the smallest building, 31,610 m2; the largest building, 136,068 m2; total 415,000 m2) through a network of pipes (1108 m). Hourly data (flow rate demand of hot/cold water from each building, heat demand of each building, pipe heat loss, and wet bulb temperature), which were measured in the A-Plant from April 2002 to March 2003, are used as the input data for the simulation. The primary energy conversion factor of electricity is set to 9.83 [MJ of primary energy/kWh of electricity] in accordance with the energy conservation laws in Japan. 2.1. Model of the heat source system The A-Plant was modeled in the case of the DHC. The heat source system configuration and the capacity of all the chillers/ heat pumps are set to the same values as those for the A-Plant. In the case of the ‘‘individual system’’, the energy consumption is calculated by summing up the simulated energy consumption of the heat source system in each building (a total of five buildings). The heat source system comprises three air-source heat pumps (Fig. 4). The total cooling capacity of the air-source heat pumps is set to 1.2 times (safety factor) that of the maximum cooling demand of each building, and the capacity ratio of the three heat pumps is 1:1:1. The heating capacity of the air-source heat pumps is 1.15 times as large as cooling capacity. The simulation algorithm of both DHC and individual system is shown in Fig. 5. 2.2. Parameters for the simulation models The measurement data of 11 DHC plants and 18 individual systems are collected in order to derive the simulation parameters described below:
(1) The energy efficiencies of the chillers and heat pumps The coefficient of performance (COP) of the chillers/ heat pumps at the rated condition (load factor of 100% and cooling water inlet temperature of 32 8C/brine temperature of 7 8C) in the DHC model is set to the same value as that of the equipment in the A-Plant (Table 1), and the COP of the air-source heat pumps at the rated condition (load factor of 100% and dry-bulb temperature of 7 8C in heating mode/ 35 8C in cooling mode) in the individual system model is set to the default value used in a general simulation program for building-energy consumption in Japan (COP in cooling mode: 2.96 and COP in heating mode: 3.12) [12]. COP is defined as COP ¼
heat produced by chiller=heat pump ½kW power consumption of chiller=heat pump ½kW (2)
The change in the COP influenced by the load factor and cooling water/brine/outdoor air temperature is based on a product specification provided by the manufacturer of the chillers/heat pumps. For example, the COP change in a centrifugal refrigerator is shown in Fig. 11 (constant-speed driven chiller). (2) The power consumption of pumps and cooling/heating tower fans Pumps The power consumption of all the pumps in the DHC model is same to the pumps in the A-Plant and that in the individual system model is set to the average value of the measurement data (Table 2). The supply pumps of the DHC including the secondary pumps consume approximately 2.5 times more power than those of the individual system at the rated condition due to a pressure drop in the network of pipes. The power consumption of the variablespeed pumps is assumed to be represented by Eqs. (3) and (4): 3 F If F 0:5F rated : P ¼ Prated (3) F rated If F < 0:5F rated
Fig. 4. System configuration of the individual system. ASHP: air-source heat pump, PC: cold water pump, and PH: hot water pump.
COP at rated condition
3 1 : P ¼ Prated 2
(4)
where P is the power consumption of the pump, Prated the power consumption of the pump at the rated condition, F the flow rate, and F rated is the maximum flow rate.
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Fig. 5. Simulation algorithm: (a) DHC and (b) individual system.
Cooling/heating tower fans The power consumption of the cooling tower fans in the DHC model is calculated at a basic unit rate (0.0058 kW/ kW), which indicates the power consumption of fans for a
unit of heat discharge from the cooling tower. This basic unit rate is derived from the annual measurement data of the DHC plants on the assumption that the amount of heat discharge is equal to the sum of the power consumption of the chillers/heat pumps and the cooling load. The basic unit
Table 2 Power consumption per unit of chiller capacity of the heat supply pumps in the models DHC model Cold water pump Centrifugal refrigerator Heat recovery chiller Centrifugal heat pump Secondary pump Variable-speed pump (2) Constant-speed pump (5) Hot Water pump Heat recovery chiller Centrifugal heat pump Secondary pump Variable-speed pump (2) Constant-speed pump (3)
Individual system model 2.45 [kW/(GJ/h)] 2.05 [kW/(GJ/h)] 3.32 [kW/(GJ/h)] 336 [m3/h] 116 [kW] 672 [m3/h] 180 [kW] 2.21 [kW/(GJ/h)] 2.91 [kW/(GJ/h)] 252 [m3/h] 75 [kW] 504 [m3/h] 132 [kW]
Air-source heat pump (Cooling mode)
4.19 [kW/(GJ/h)]
–
Air-source heat pump (Heating mode)
–
Values within parentheses in the ‘‘secondary pump’’ columns indicate the number of pumps.
4.19 [kW/(GJ/h)]
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for the heating tower fans is assumed to be 0.0224 kW/kW based on the annual measurement data of the A-Plant. In the individual system model, the power consumption of the fans is included in that of the air-source heat pump. (3) Thermal storage tanks The DHC in this model has five large-capacity thermal storage tanks, as shown in Fig. 3. The charging storage begins at 22.00 h, and chillers/heat pumps are controlled such that the storage is fully charged at 08.00 h. The discharging storage is controlled during all the hours by a chiller-priority control strategy according to which the chillers/heat pumps are operated at full load and the heat load in excess of the chiller/heat pump capacity is met from the storage. However, during the electrical peak period (13.00–16.00 h), the heat load requirement is met from only the discharging storage when possible. The hourly heat loss from the thermal storage tanks is set at 0.12% of the amount of heat available for discharge. The heat loss rate is devised such that the simulated annual amount of heat loss is equal to the measured value in the A-Plant. (4) Multiple chiller/heat pump control sequence The chiller/heat pump control sequence determines the chiller/heat pump to be operated for meeting both the required heat load and flow rate demand. The control sequence of the DHC model is devised such that higherefficiency chillers/heat pumps have higher priority. For example, in a cooling heat source system, the operation of the heat recovery chiller is given top priority only in the case where both the cooling and heating loads exist. Otherwise, the centrifugal refrigerator is operated by priority. The centrifugal heat pump is operated as rarely as possible. The control sequence of the individual system model is simple, and it merely determines the number of heat pumps to be operated. In the operation of the chillers/heat pumps, the following two constraints exist: ‘‘dead band [13]’’ and ‘‘flow rate of the bypass pipe’’. However, both are not incorporated in the DHC model because the plant can handle the issues as mentioned below by the effect of thermal storage. The dead band constrains the chiller operation at a higher load factor than a
Fig. 6. Operation of chillers/heat pumps constrained by dead band in the case where the capacity of each chiller is 10 GJ. The constraint of the dead band gets lifted if the last chiller runs.
certain value [=1.0—the dead band margin] in order to respond to a sharp increase in the heat load and to avoid frequent switching of the chiller/heat pumps (Fig. 6). The measured value of the margin in the individual system is 0.3. The flow rate of the bypass pipe should exceed a certain value in order to respond to a sharp increase in the flow rate demand. Therefore, the total flow rate of chillers/heat pumps must always exceed the sum of the flow rate demand and the minimum flow rate of the bypass pipe. The minimum flow rate of the bypass pipe in an individual system is set at 50% of the designed flow rate of the smallest capacity chiller. 3. Simulation results 3.1. Accuracy verification of the simulation result A comparison between the simulation result and the measurement data of the A-Plant is performed in order to verify the accuracy of the DHC model. The simulated EER of the DHC is 1.25, and there is a 5% discrepancy when compared with the measured EER value of 1.19. From Fig. 7 (in which a comparison of the monthly power consumption of the chillers/heat pumps is shown), it can be seen that both the power consumption trends of the simulation result and the measurement data are similar except in the following regard: in the actual condition, the heat recovery chiller in the cooling mode (COP at the rated condition: 3.8) is operated for charging the storage in summer because the operation with the aim of shifting the power consumption to nighttime as much as possible is implemented from the economic point of view in 2002. However, the operation in the simulation model is oriented toward high-energy efficiency, and a heat recovery chiller is never operated in the cooling mode. This difference in operation is one of main reasons why the EER value obtained in the simulation result is higher. 3.2. Comparison in power consumption between DHC and individual system The simulation results of the power consumption shown in Table 3 reveal that the DHC consumes 22% less power than the
Fig. 7. A comparison of the monthly power consumption of the chillers and heat pumps between the simulation result and measurement data.
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Table 3 Simulated annual power consumption Power consumption (MWh) Cooling heat source system Chillers/heat pumps Cooling tower fans Cooling water pumps Supply pumps Cold water pumps Secondary pumps Subtotal energy consumption Heating heat source system Heat pumps Heating tower fans Brine pumps Supply pumps Hot water pumps Secondary pumps Subtotal energy consumption Total energy consumption EER
DHC
Individual system
Energy-saving ratio (%)
6,077 207 540
9,407 – –
35% – –
318 723
1,714 –
39%
7,865
11,121
29%
3,371 151 239
3,450 – –
2% – –
128 294
952 –
56%
4,183
4,402
5%
12,048 1.25
individual system. The results show that the energy-saving ratio of the DHC in the cooling heat source system reaches 29%, whereas in the heating heat source system, the energy-saving ratio is 5% when compared with the individual system. Detailed discussions on the results are as follows: (1) Cooling heat source system The COP of the chillers/heat pumps used in the DHC is approximately 1.5 times that of the air-source heat pumps used in the individual system. In addition, a decrease in the COP caused by the part-load operation can be prevented because the chillers/heat pumps are operated at full load at all the times by utilizing thermal storage, as shown in Fig. 8. Thus, the power consumption of the chillers/heat pumps in the DHC is considerably reduced when compared with that of the individual system. In addition, the DHC saves the power consumption of supply pumps, although at the rated condition the DHC consumes more power for supply pumps than the individual system. This is mainly caused by the following two factors: one is the utilization of variablespeed pumps in the secondary pump system, and the other is the chiller operation at full load that results in efficient heat transportation, or more specifically, less power for transporting a unit of heat. (2) Heating heat source system The operation of the heat recovery chiller is given top priority in the DHC. However, the heat load profile of the APlant is not preferable for the heat recovery chiller since the heating demand is much greater in winter than the cooling demand (Fig. 9). As a result, the heat recovery chiller produces only 28% of the total heating heat. The COP of the centrifugal heat pump, which produces 72% of the heating heat, nearly equals the COP of the air-source heat pump. However, the centrifugal heat pump is inferior to the
15,523 0.97
22% –
air-source heat pump in the system COP considering the power consumption of the pump and fan because the airsource heat pump does not need the heating-tower pump (brine pump). This is the main reason for not exhibiting a large energy-saving effect when compared with the result of the cooling heat source system.
Fig. 8. Chiller operation during the summer on August 14: (a) DHC and (b) individual system.
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(b)
Fig. 9. Monthly pattern of cooling/heating heat demand. Cooling (other plants) is the average cooling load profile of three other DHC plants.
(c)
3.3. Factorial analysis of the energy-saving effect of the DHC
(d)
The factorial analysis of the energy-saving effect of the DHC is carried out by performing optional simulations, as shown in Tables 4 and 5.
(e)
(a) The difference between RUN-0 and RUN-1 in Table 4 represents the concentration effect. The data combination of
the heat demand of each building enables the chillers/heat pumps to run at a high load factor. For example, when there is less heat demand during nighttime, at least one chiller must be operated for a small heat demand for each building in the individual system. On the other hand, the DHC can deal with a small heat demand by operating a small number of chillers, which leads to the chiller operation at a relatively high load factor. The difference between RUN-1 and RUN-2 represents the heat transportation loss, which includes the heat loss from the pipeline and the additional power consumption of the supply pumps. The difference between RUN-2 and RUN-3 represents the economy of scale. A high-efficiency chiller/heat pump can be utilized in the DHC because of the economy of scale. The difference between RUN-3 and RUN-4 represents the grade of operation. The chillers/heat pumps in the DHC are efficiently operated by well-trained operators. The difference between RUN-4 and RUN-5 represents the thermal storage effect. By utilizing thermal storage, the chillers/heat pumps can be operated at full load at all the times. In addition, the heat recovery chiller can be operated in the heat recovery mode since thermal storage enables the
Table 4 Parameter setting of optional simulations RUN-0
RUN-1
RUN-2
RUN-3
RUN-4
RUN-5
Heat source system configuration/heat load
Individual
DHC (without thermal storage: Type 1)
DHC (without thermal storage: Type 1)
DHC (without thermal storage: Type 2)
DHC (with thermal storage)
Pipe heat loss Power consumption of heat supply pump Grade of operation
Not included Individual
Not included Individual
Individual
Individual
Included DHC (without thermal storage) Individual
Included DHC (without thermal storage) Individual
DHC (without thermal storage: Type 2) Included DHC (without thermal storage) DHC (without thermal storage)
Included DHC (with thermal storage) DHC (with thermal storage)
Details of the parameter settings are shown in Table 5. Table 5 Details of the parameter settings Individual
DHC (without thermal storage) Type 1
Type 2 Seven centrifugal heat pumps
DHC (with thermal storage)
Heat source system configuration
Three air-source heat pumps
Seven air-source heat pumps
Power consumption of heat supply pump
4.19 [kW/(GJ/h)]
(Cooling) 3.32 [kW/(GJ/h)] plus secondary pumps (Heating) 2.91 [kW/(GJ/h)] plus secondary pumps
See ‘‘DHC model’’ in Table 2
30 50% of the designed flow rate of the smallest capacity chiller
10 20% of the designed flow rate of the smallest capacity chiller
No constraint No constraint
Grade of operation Dead band margin (%) Minimum flow rate of the bypass pipe
See Table 1
In the DHC (without thermal storage), parameters of the ‘‘grade of operation’’ are based on an actual operation of the DHC plant without thermal storage, and the number of chillers/heat pumps is set to the mean value of the existing electric-driven heat-pump-type DHC plants. The chiller capacity is determined to be equally divided as that in the individual system (Section 2.1).
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heat recovery chiller to produce both cooling and heating heat corresponding to the daily balance of the heating/ cooling demand. The factorial analysis result is shown in Fig. 10. Economy of scale and the thermal storage effect demonstrate considerable energy-saving effect in the cooling heat source system. On the other hand, in the heating heat source system, it shows that economy of scale is a key factor to lower the EER value. The centrifugal heat pump in the heating mode is inferior to the airsource heat pump in the system COP; therefore, it is very important that the heating heat should be produced by highefficiency equipment such as a heat recovery chiller to achieve a high-EER value. 4. Effect of the other potential parameters on the EER value The DHC is considered to exhibit energy-saving potential by introducing a high-efficiency technique. In addition, in practice, not all the heat source systems in the individual systems are equipped with air-source heat pumps. In this section, the effects of the other factors, which have not been verified in Section 3 (the base case), are discussed.
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RUN-A: Introduction of centrifugal refrigerators in the individual system An air-source heat pump is used for producing cooling heat in the individual system in the base case. However, centrifugal refrigerators can be introduced for large-scale buildings even in the individual system. The case where centrifugal refrigerators are used instead of air-source heat pumps is simulated. RUN-B: Increase in the cooling demand during winter In the base case of the DHC, the heat recovery chiller cannot produce considerable heating heat since the ratio of the cooling demand to the heating demand during winter (from November to March) is only 0.28. We have simulated the case in which the hourly cooling demand during winter doubles in order to enable the heat recovery chiller to produce more heating heat. In this condition, the ratio of the cooling demand during winter reaches 0.56; however, this is not unrealistic when compared with the average cooling load of three other plants, as shown in Fig. 9. RUN-C: Introduction of variable-speed driven chillers Variable-speed driven chillers are superior to conventional chillers (constant-speed driven chillers) in terms of their change in the COP caused by part-load operation and the cooling water temperature, as shown in Fig. 11. We have simulated the case in which the conventional centrifugal chillers/heat pumps (only in the cooling mode) in the DHC are changed into variable-speed driven chillers/heat pumps. The COP values shown in Table 1 are used in this simulation. RUN-D: Utilization of river water as the heat source/sink in the chiller and heat pump One of the remarkable advantages of the DHC is that it is possible to utilize natural and unused heat sinks/sources such as sewage water and seawater. However, the A-Plant does not currently utilize them. We have simulated the case in which river water is utilized as the heat source/sink in the DHC model.
Fig. 10. Power-saving effect of the factors: (a) cooling heat source system and (b) heating heat source system.
Fig. 11. COP change by load factor and cooling water inlet temperature.
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Table 6 Simulation results with the additional parameters
5. Conclusions
Power consumption (MWh)
Individual (base case) Individual (RUN-A) DHC (base case) DHC (RUN-B) DHC (RUN-C) DHC (RUN-D)
EER
For cooling
For heating
Total
11,121 9,823 7,865 7,967 7,457 7,334
4402 4402 4183 3897 4183 3500
15523 14225 12048 11864 11640 10834
0.97 1.06 1.25 1.32 1.30 1.39
In RUN-B, the cooling demand is higher than the other cases. Therefore, the total power consumption of RUN-B is not comparable with the other cases.
The river water temperature and power consumption of pumps for transporting river water used in the model are derived from the measurement data of an existing DHC plant utilizing river water. The results of these simulations are shown in Table 6. The power consumption for cooling in RUN-A decreases by 12% when compared with the result of the base case of the individual system. However, when compared with the DHC, the power consumption is still greater. The major advantage of the economy of scale obtained by the DHC, as shown in Fig. 10, disappears in the case in which the centrifugal refrigerators are introduced into the individual system. However, due to the other advantages such as the thermal storage effect and the concentration effect, the DHC is superior to the individual system in RUN-A in terms of the EER value. In RUN-B, the ratio of the amount of heating heat produced by the heat recovery chiller increases to 39% from the result obtained in the base case (28%) by doubling the cooling demand during winter. As a result, the power consumption for heating is reduced by 6.8%. The result of RUN-C shows that the introduction of the variable-speed driven chillers reduces the power consumption for cooling by 5.2%. In the DHC, the chillers are operated at full load at all times due to thermal storage, and improvement in the part-load performance by introducing the variable-speed driven chiller is ineffective for saving energy. Thus, the introduction of the variable-speed driven chiller only provides the energysaving effect of improving the ratio of the COP change by cooling water temperature. The result of RUN-D shows that the utilization of river water is very effective in saving power. The river water temperature is distributed from 20 to 30 8C in summer and at intermediate time, and there is a slight discrepancy in the temperature between the river water and cooling water from the cooling tower. On the other hand, the river water temperature is distributed from 10 to 15 8C during winter; this temperature is considerably higher than the brine temperature from the heating tower (–5 0 8C). As a result, the COP of the centrifugal heat pump in the heating mode is considerably improved. Therefore, the energy-saving effect is evident, especially for heating. However, it should be noted that the energy-saving effect depends on various factors such as the temperature of river water and power consumption of the pumps for transporting the river water.
In this study, the energy-saving effect of the electric-driven heat-pump-type DHC is examined by using simulation models. In addition, the technique for improving the energy efficiency of the DHC is verified. The results are as follows: The simulation results show that the DHC provides an energy saving of 29% for cooling when compared with the individual system; meanwhile, the energy-saving ratio is 5% for heating. The difference in energy consumption between the DHC and the individual system is verified by means of a factorial analysis. The analysis result shows that the economy of scale (which indicates the effect of the introduction of the centrifugal refrigerators) and the thermal storage effect contribute significantly to saving energy for cooling. On the other hand, the disadvantage of economy of scale is shown for heating because most of the heating heat is produced by the centrifugal heat pump, which system COP is smaller than that of the air-source heat pump used in the individual system. The DHC can provide large energy-saving potential by improving the heat load profile or utilizing the natural and unused heat sinks/sources, both of which contribute to reducing the energy consumption particularly for heating. The heat load profile is proved to be one of the important factors for achieving a high-EER value. Therefore, it is necessary to thoroughly analyze the relationship between the EER value and the heat load profile. The utilization of the natural and unused heat sinks/sources is unique to the DHC, and the energy-saving effect is also significant. It is one of the most notable points of the DHC to be demonstrated in view of energy conservation from now on. A high-efficiency technique such as the variable-speed technique has the potential to reduce the energy consumption for an individual system as well as the DHC. When the highefficiency technique can be economically and technologically introduced into an individual system, there would be some changes in the energy conservation of the DHC. A simulation considering the progress of technology should be performed in order to verify the energy-saving effect of the DHC in the future. The result could be useful for verifying the validity of expanding the networks of pipes especially in Japan where the networks are not developed yet. The simulated EER of the individual system is 0.97, and this value exceeds the average EER value of 0.71 (obtained from the measurement survey) by 0.26 (The number of samples in this measurement survey is only 13, and the result could not be exact.) [14]. Thus, there could be some faults causing a decrease in the EER value such as aged deterioration, which are not observed in our analysis of the measurement data. A field survey on the individual system is required to reveal the actual condition of the individual system. Acknowledgements This work has been undertaken as a cooperative research with the Japan Heat Service Utilities Association. The authors
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wish to thank the members of the Special Committee on the Evaluation Method for Energy Efficiency of District Heating and Cooling, Society of Heating, Air-conditioning and Sanitary Engineers of Japan (SHASE) for their valuable advice. This research was supported by a Grant-in-Aid for Scientific Research from the Ministry of Education, Culture, Sports, Science and Technology, Japan, no. 18360273. References [1] M. Sakawa, K. Kato, S. Ushiro, M. Inaoka, Operation planning of district heating and cooling plants using genetic algorithms for mixed integer programming, Applied Soft Computing 1 (2) (2001) 139–150. [2] ASHRAE Research, 2004 ASHRAE Handbook, HVAC Systems and Equipment, 2004, pp.11.1–11.2. [3] R. Lazzarin, M. Noro, Local or district heating by natural gas: which is better from energetic, environmental and economic point of views? Applied Thermal Engineering 26 (2–3) (2006) 244–250. [4] T. Nagota, Y. Shimoda, N. Isayama, R. Kubara, M. Mizuno, Verification of energy saving advantage of co-generation incorporated district heating and cooling system, in: Proceedings of the IBPSA, Australia, (2006), pp. 141–148. [5] S. Sadohara, K. Nagano, et al., Study of diffusion of district heating and cooling in Japan and its effects on global environment preservation, Transactions of AIJ 510 (1998) 61–67 (in Japanese).
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[6] Y. Shimoda, T. Nagota, N. Isayama, M. Mizuno, Verification of energy efficiency of district heating and cooling system by simulation considering design and operation parameters, Build. Environ. 43 (2008) 569–577. [7] T. Nagota, Y. Shimoda, M. Mizuno, Comparison study on energy conservation effect between district heating and cooling system and individual heat source system—simulation evaluation of co-generation system incorporated absorption chiller & boiler system, Transactions of SHASE (2007) 37–47 (in Japanese). [8] M. Momota, T. Inoue, T. Ibamoto, R. Yanagihara, A. Okagaki, A study on reducing the environmental load of district heating and cooling using a large thermal storage tank, Journal of Environmental Engineering (Transactions of AIJ) (2003) 83–87 (in Japanese). [9] H. Tanaka, T. Tomita, M. Okumiya, Feasibility study of a district energy system with seasonal water thermal storage, Solar Energy 69 (6) (2000) 535–547. [10] S.M. Hasnain, N.M. Alabbadi, Need for thermal-storage air-conditioning in Saudi Arabia, Applied Energy 65 (1–4) (2000) 153–164. [11] Agency for Natural Resources and Energy, Report on the Infrastructure Development to Introduce Unutilized Heat Energy, 2003 (in Japanese). [12] Institute for Building Environment and Energy Conservation, Manual of Building Energy Consumption Simulator, 1999 (in Japanese). [13] G.A. Johnson, Optimization techniques for a centrifugal chiller plant using a programmable controller, ASHRAE Transactions 91 (2) (1985) 835–847. [14] Agency for Natural Resources and Energy, Report on the Promotion of Small-scale Area Energy Network, 2005 (in Japanese).
Verification of the energy-saving effect of the district heating and cooling system—Simulation of an electric-driven heat pump system Tomoji Nagota, Yoshiyuki Shimoda *, Minoru Mizuno Division of Sustainable Energy and Environmental Engineering, Osaka University, 2-1 Yamadaoka, Suita City, Osaka 565-0871, Japan Received 17 March 2007; received in revised form 9 May 2007; accepted 14 May 2007
Abstract A district heating and cooling system (DHC) is expected to be a promising energy-saving measure for high-density business areas in Japan. However, it has not been verified what advantages of the DHC are important for energy conservation. The clarification of this issue is supposed to contribute to improving the energy efficiency of the DHC. This paper focuses on the electric-driven heat-pump-type DHC, which uses only electricity as its energy source. An existing DHC plant has been selected for the case study, and its energy efficiency is examined by a simulation model that uses parameters derived from the measurement data. The simulation results for the plant reveal that the DHC exhibits an energy-saving effect of 29% for cooling when compared with the individual heat source system mainly due to the following two advantages: ‘‘economy of scale in chillers/heat pumps’’ and ‘‘thermal storage effect’’. Further, the energy-saving ratio for heating is only 5% since heat recovery chiller cannot be operated sufficiently due to the lack of cooling demand during winter. In addition, we have examined high-efficiency techniques such as the utilization of river water. As a result, potential energy-saving impact on the DHC is evaluated. # 2007 Elsevier B.V. All rights reserved. Keywords: District heating and cooling system; Simulation; Energy-saving effect; Electric-driven heat pump system
1. Introduction In Japan, the number of district heating and cooling systems (DHC) has constantly increased to over 150 since the first DHC commenced supplying heat in 1970. The DHC is expected to have various social advantages such as air-pollution abatement and intensive use of chillers and machinery space [1,2]. Above all, energy conservation is one of the most important advantages of the DHC with regard to the mitigation of global warming [3–5]. The DHC is classified into two categories according to the type of energy source (‘‘absorption chiller and boiler’’ and ‘‘electric-driven heat pump’’), and the factors of the DHC that influence energy conservation differ between the types of energy sources. The absorption-chiller-and-boiler-type DHC, which uses gas as its main energy source, conserves energy mainly due to the ‘‘co-generation system’’ and ‘‘concentration effect’’ [6,7]. In the electric-driven heatpump-type DHC, which uses electricity as its main energy
* Corresponding author. Tel.: +81 6 6879 7665; fax: +81 6 6879 7665. E-mail address: [email protected] (Y. Shimoda). 0378-7788/$ – see front matter # 2007 Elsevier B.V. All rights reserved. doi:10.1016/j.enbuild.2007.05.007
source, the following factors are considered to mainly contribute to energy saving [2,8–10]. High-efficiency equipment: Centrifugal refrigerators, which usually are not introduced for small-scale heat source system, produce cooling very efficiently. Thermal storage: Chillers/heat pumps can be operated at full load by utilizing the thermal storage capability. In addition, a heat recovery chiller with double-bundle condenser (referred to as a ‘‘heat recovery chiller’’ hereafter) can be introduced since a temporal difference between the heating and cooling loads can be adjusted. This type of DHC plant is expected to exhibit the large energy-saving effect in the place where there is a large amount of cooling demand. In Japan, 21% of the DHC plants adopt electric-driven heat-pump-type. However, a measurement survey on the DHC [11] reveals that the energy efficiency ratio (EER) of each electric-driven heat-pump-type DHC plant is widely distributed, as shown in Fig. 1. EER is defined as EER ¼
total amount of supplied heat ½GJ=year total primary energy consumption ½GJ=year
(1)
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Fig. 1. Distribution of EER in electric-driven heat-pump-type DHC plants adapted from a measurement data survey [11].
This indicates that the energy-saving effect from the abovementioned factors does not appear uniformly at the same level in all the plants. In other words, the potential of the energy-saving effect from some important factors is achieved in high-EER plants; however, it fails to be achieved in low-EER plants due to various issues. Thus far, any quantitative verification of the energy-saving effect of each factor of the electric-driven heat-pump-type DHC has not been conducted, and it is unclear what factors are important for the reduction in energy consumption. The clarification of this issue is considered to provide useful information about not only the modification of the existing plants and the planning of new DHC plants but also the necessity of the DHC in the future. In this study, an existing DHC plant is selected for a case study, and the simulation models of the DHC and an individual system (the case where every building has own heat source system) – both of which supply heat to the same heat-supply area of the case-study plant – are developed in order to perform a factorial analysis of the energy-saving effect of the DHC by comparing
Fig. 2. Flow chart for the verification of the energy-conservation effect of the DHC.
both the simulation results. The simulation parameters are derived from the measurement data of the existing heat source systems for the purpose of improving the accuracy of simulation. Fig. 2 shows a flow chart of this study. In addition, a highefficiency technique, such as the utilization of river water and the improvement in the heat load profile, is examined to reveal the potential of the DHC in improving the energy efficiency. 2. Simulation model An existing electric-driven heat-pump-type DHC plant (APlant), which achieves the second highest EER among the DHC plants in Japan and exhibits the energy-saving effect from advantages of the DHC properly, has been selected for the case study. The A-Plant comprises electric-driven chillers/heat
Fig. 3. System configuration of the A-Plant. The operation mode of cooling/heating tower, heating-tower mode (using brine) and cooling-tower mode (using cooling water), is changed corresponding to the operation mode of the centrifugal heat pump. The number of secondary hot/cold water pumps shown in the figure is different from that of the A-Plant. CR: centrifugal refrigerator, HRC: heat recovery chiller, HP: centrifugal heat pump, PCD: cooling water pump, BMP: brine/cooling water pump, PC: cold water pump, PH: hot water pump, S_PC: secondary cold water pump, S_PH: secondary hot water pump.
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Table 1 Chillers and heat pumps in the A-Plant Equipment
Operating mode
Heating/cooling
Capacity (GJ/h)
CR1/CR2
Centrifugal refrigerator
HRC1/HRC2
Heat recovery chiller
HP1/HP2
Centrifugal heat pump
Cooling mode
Cooling
14.9
5.3
Heat recovery mode
Cooling Heating
5.0 6.8
6.7
Cooling mode Heating mode
Cooling Heating
18.3 12.7
4.7 3.2
pumps and thermal storage tanks, as shown in the system configuration (Fig. 3, Table 1), and it supplies cold water (the annual amount of supplied heat in fiscal 2002: 29.6 GWh) and hot water (11.6 GWh) to five buildings (floor areas: the smallest building, 31,610 m2; the largest building, 136,068 m2; total 415,000 m2) through a network of pipes (1108 m). Hourly data (flow rate demand of hot/cold water from each building, heat demand of each building, pipe heat loss, and wet bulb temperature), which were measured in the A-Plant from April 2002 to March 2003, are used as the input data for the simulation. The primary energy conversion factor of electricity is set to 9.83 [MJ of primary energy/kWh of electricity] in accordance with the energy conservation laws in Japan. 2.1. Model of the heat source system The A-Plant was modeled in the case of the DHC. The heat source system configuration and the capacity of all the chillers/ heat pumps are set to the same values as those for the A-Plant. In the case of the ‘‘individual system’’, the energy consumption is calculated by summing up the simulated energy consumption of the heat source system in each building (a total of five buildings). The heat source system comprises three air-source heat pumps (Fig. 4). The total cooling capacity of the air-source heat pumps is set to 1.2 times (safety factor) that of the maximum cooling demand of each building, and the capacity ratio of the three heat pumps is 1:1:1. The heating capacity of the air-source heat pumps is 1.15 times as large as cooling capacity. The simulation algorithm of both DHC and individual system is shown in Fig. 5. 2.2. Parameters for the simulation models The measurement data of 11 DHC plants and 18 individual systems are collected in order to derive the simulation parameters described below:
(1) The energy efficiencies of the chillers and heat pumps The coefficient of performance (COP) of the chillers/ heat pumps at the rated condition (load factor of 100% and cooling water inlet temperature of 32 8C/brine temperature of 7 8C) in the DHC model is set to the same value as that of the equipment in the A-Plant (Table 1), and the COP of the air-source heat pumps at the rated condition (load factor of 100% and dry-bulb temperature of 7 8C in heating mode/ 35 8C in cooling mode) in the individual system model is set to the default value used in a general simulation program for building-energy consumption in Japan (COP in cooling mode: 2.96 and COP in heating mode: 3.12) [12]. COP is defined as COP ¼
heat produced by chiller=heat pump ½kW power consumption of chiller=heat pump ½kW (2)
The change in the COP influenced by the load factor and cooling water/brine/outdoor air temperature is based on a product specification provided by the manufacturer of the chillers/heat pumps. For example, the COP change in a centrifugal refrigerator is shown in Fig. 11 (constant-speed driven chiller). (2) The power consumption of pumps and cooling/heating tower fans Pumps The power consumption of all the pumps in the DHC model is same to the pumps in the A-Plant and that in the individual system model is set to the average value of the measurement data (Table 2). The supply pumps of the DHC including the secondary pumps consume approximately 2.5 times more power than those of the individual system at the rated condition due to a pressure drop in the network of pipes. The power consumption of the variablespeed pumps is assumed to be represented by Eqs. (3) and (4): 3 F If F 0:5F rated : P ¼ Prated (3) F rated If F < 0:5F rated
Fig. 4. System configuration of the individual system. ASHP: air-source heat pump, PC: cold water pump, and PH: hot water pump.
COP at rated condition
3 1 : P ¼ Prated 2
(4)
where P is the power consumption of the pump, Prated the power consumption of the pump at the rated condition, F the flow rate, and F rated is the maximum flow rate.
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Fig. 5. Simulation algorithm: (a) DHC and (b) individual system.
Cooling/heating tower fans The power consumption of the cooling tower fans in the DHC model is calculated at a basic unit rate (0.0058 kW/ kW), which indicates the power consumption of fans for a
unit of heat discharge from the cooling tower. This basic unit rate is derived from the annual measurement data of the DHC plants on the assumption that the amount of heat discharge is equal to the sum of the power consumption of the chillers/heat pumps and the cooling load. The basic unit
Table 2 Power consumption per unit of chiller capacity of the heat supply pumps in the models DHC model Cold water pump Centrifugal refrigerator Heat recovery chiller Centrifugal heat pump Secondary pump Variable-speed pump (2) Constant-speed pump (5) Hot Water pump Heat recovery chiller Centrifugal heat pump Secondary pump Variable-speed pump (2) Constant-speed pump (3)
Individual system model 2.45 [kW/(GJ/h)] 2.05 [kW/(GJ/h)] 3.32 [kW/(GJ/h)] 336 [m3/h] 116 [kW] 672 [m3/h] 180 [kW] 2.21 [kW/(GJ/h)] 2.91 [kW/(GJ/h)] 252 [m3/h] 75 [kW] 504 [m3/h] 132 [kW]
Air-source heat pump (Cooling mode)
4.19 [kW/(GJ/h)]
–
Air-source heat pump (Heating mode)
–
Values within parentheses in the ‘‘secondary pump’’ columns indicate the number of pumps.
4.19 [kW/(GJ/h)]
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for the heating tower fans is assumed to be 0.0224 kW/kW based on the annual measurement data of the A-Plant. In the individual system model, the power consumption of the fans is included in that of the air-source heat pump. (3) Thermal storage tanks The DHC in this model has five large-capacity thermal storage tanks, as shown in Fig. 3. The charging storage begins at 22.00 h, and chillers/heat pumps are controlled such that the storage is fully charged at 08.00 h. The discharging storage is controlled during all the hours by a chiller-priority control strategy according to which the chillers/heat pumps are operated at full load and the heat load in excess of the chiller/heat pump capacity is met from the storage. However, during the electrical peak period (13.00–16.00 h), the heat load requirement is met from only the discharging storage when possible. The hourly heat loss from the thermal storage tanks is set at 0.12% of the amount of heat available for discharge. The heat loss rate is devised such that the simulated annual amount of heat loss is equal to the measured value in the A-Plant. (4) Multiple chiller/heat pump control sequence The chiller/heat pump control sequence determines the chiller/heat pump to be operated for meeting both the required heat load and flow rate demand. The control sequence of the DHC model is devised such that higherefficiency chillers/heat pumps have higher priority. For example, in a cooling heat source system, the operation of the heat recovery chiller is given top priority only in the case where both the cooling and heating loads exist. Otherwise, the centrifugal refrigerator is operated by priority. The centrifugal heat pump is operated as rarely as possible. The control sequence of the individual system model is simple, and it merely determines the number of heat pumps to be operated. In the operation of the chillers/heat pumps, the following two constraints exist: ‘‘dead band [13]’’ and ‘‘flow rate of the bypass pipe’’. However, both are not incorporated in the DHC model because the plant can handle the issues as mentioned below by the effect of thermal storage. The dead band constrains the chiller operation at a higher load factor than a
Fig. 6. Operation of chillers/heat pumps constrained by dead band in the case where the capacity of each chiller is 10 GJ. The constraint of the dead band gets lifted if the last chiller runs.
certain value [=1.0—the dead band margin] in order to respond to a sharp increase in the heat load and to avoid frequent switching of the chiller/heat pumps (Fig. 6). The measured value of the margin in the individual system is 0.3. The flow rate of the bypass pipe should exceed a certain value in order to respond to a sharp increase in the flow rate demand. Therefore, the total flow rate of chillers/heat pumps must always exceed the sum of the flow rate demand and the minimum flow rate of the bypass pipe. The minimum flow rate of the bypass pipe in an individual system is set at 50% of the designed flow rate of the smallest capacity chiller. 3. Simulation results 3.1. Accuracy verification of the simulation result A comparison between the simulation result and the measurement data of the A-Plant is performed in order to verify the accuracy of the DHC model. The simulated EER of the DHC is 1.25, and there is a 5% discrepancy when compared with the measured EER value of 1.19. From Fig. 7 (in which a comparison of the monthly power consumption of the chillers/heat pumps is shown), it can be seen that both the power consumption trends of the simulation result and the measurement data are similar except in the following regard: in the actual condition, the heat recovery chiller in the cooling mode (COP at the rated condition: 3.8) is operated for charging the storage in summer because the operation with the aim of shifting the power consumption to nighttime as much as possible is implemented from the economic point of view in 2002. However, the operation in the simulation model is oriented toward high-energy efficiency, and a heat recovery chiller is never operated in the cooling mode. This difference in operation is one of main reasons why the EER value obtained in the simulation result is higher. 3.2. Comparison in power consumption between DHC and individual system The simulation results of the power consumption shown in Table 3 reveal that the DHC consumes 22% less power than the
Fig. 7. A comparison of the monthly power consumption of the chillers and heat pumps between the simulation result and measurement data.
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Table 3 Simulated annual power consumption Power consumption (MWh) Cooling heat source system Chillers/heat pumps Cooling tower fans Cooling water pumps Supply pumps Cold water pumps Secondary pumps Subtotal energy consumption Heating heat source system Heat pumps Heating tower fans Brine pumps Supply pumps Hot water pumps Secondary pumps Subtotal energy consumption Total energy consumption EER
DHC
Individual system
Energy-saving ratio (%)
6,077 207 540
9,407 – –
35% – –
318 723
1,714 –
39%
7,865
11,121
29%
3,371 151 239
3,450 – –
2% – –
128 294
952 –
56%
4,183
4,402
5%
12,048 1.25
individual system. The results show that the energy-saving ratio of the DHC in the cooling heat source system reaches 29%, whereas in the heating heat source system, the energy-saving ratio is 5% when compared with the individual system. Detailed discussions on the results are as follows: (1) Cooling heat source system The COP of the chillers/heat pumps used in the DHC is approximately 1.5 times that of the air-source heat pumps used in the individual system. In addition, a decrease in the COP caused by the part-load operation can be prevented because the chillers/heat pumps are operated at full load at all the times by utilizing thermal storage, as shown in Fig. 8. Thus, the power consumption of the chillers/heat pumps in the DHC is considerably reduced when compared with that of the individual system. In addition, the DHC saves the power consumption of supply pumps, although at the rated condition the DHC consumes more power for supply pumps than the individual system. This is mainly caused by the following two factors: one is the utilization of variablespeed pumps in the secondary pump system, and the other is the chiller operation at full load that results in efficient heat transportation, or more specifically, less power for transporting a unit of heat. (2) Heating heat source system The operation of the heat recovery chiller is given top priority in the DHC. However, the heat load profile of the APlant is not preferable for the heat recovery chiller since the heating demand is much greater in winter than the cooling demand (Fig. 9). As a result, the heat recovery chiller produces only 28% of the total heating heat. The COP of the centrifugal heat pump, which produces 72% of the heating heat, nearly equals the COP of the air-source heat pump. However, the centrifugal heat pump is inferior to the
15,523 0.97
22% –
air-source heat pump in the system COP considering the power consumption of the pump and fan because the airsource heat pump does not need the heating-tower pump (brine pump). This is the main reason for not exhibiting a large energy-saving effect when compared with the result of the cooling heat source system.
Fig. 8. Chiller operation during the summer on August 14: (a) DHC and (b) individual system.
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(b)
Fig. 9. Monthly pattern of cooling/heating heat demand. Cooling (other plants) is the average cooling load profile of three other DHC plants.
(c)
3.3. Factorial analysis of the energy-saving effect of the DHC
(d)
The factorial analysis of the energy-saving effect of the DHC is carried out by performing optional simulations, as shown in Tables 4 and 5.
(e)
(a) The difference between RUN-0 and RUN-1 in Table 4 represents the concentration effect. The data combination of
the heat demand of each building enables the chillers/heat pumps to run at a high load factor. For example, when there is less heat demand during nighttime, at least one chiller must be operated for a small heat demand for each building in the individual system. On the other hand, the DHC can deal with a small heat demand by operating a small number of chillers, which leads to the chiller operation at a relatively high load factor. The difference between RUN-1 and RUN-2 represents the heat transportation loss, which includes the heat loss from the pipeline and the additional power consumption of the supply pumps. The difference between RUN-2 and RUN-3 represents the economy of scale. A high-efficiency chiller/heat pump can be utilized in the DHC because of the economy of scale. The difference between RUN-3 and RUN-4 represents the grade of operation. The chillers/heat pumps in the DHC are efficiently operated by well-trained operators. The difference between RUN-4 and RUN-5 represents the thermal storage effect. By utilizing thermal storage, the chillers/heat pumps can be operated at full load at all the times. In addition, the heat recovery chiller can be operated in the heat recovery mode since thermal storage enables the
Table 4 Parameter setting of optional simulations RUN-0
RUN-1
RUN-2
RUN-3
RUN-4
RUN-5
Heat source system configuration/heat load
Individual
DHC (without thermal storage: Type 1)
DHC (without thermal storage: Type 1)
DHC (without thermal storage: Type 2)
DHC (with thermal storage)
Pipe heat loss Power consumption of heat supply pump Grade of operation
Not included Individual
Not included Individual
Individual
Individual
Included DHC (without thermal storage) Individual
Included DHC (without thermal storage) Individual
DHC (without thermal storage: Type 2) Included DHC (without thermal storage) DHC (without thermal storage)
Included DHC (with thermal storage) DHC (with thermal storage)
Details of the parameter settings are shown in Table 5. Table 5 Details of the parameter settings Individual
DHC (without thermal storage) Type 1
Type 2 Seven centrifugal heat pumps
DHC (with thermal storage)
Heat source system configuration
Three air-source heat pumps
Seven air-source heat pumps
Power consumption of heat supply pump
4.19 [kW/(GJ/h)]
(Cooling) 3.32 [kW/(GJ/h)] plus secondary pumps (Heating) 2.91 [kW/(GJ/h)] plus secondary pumps
See ‘‘DHC model’’ in Table 2
30 50% of the designed flow rate of the smallest capacity chiller
10 20% of the designed flow rate of the smallest capacity chiller
No constraint No constraint
Grade of operation Dead band margin (%) Minimum flow rate of the bypass pipe
See Table 1
In the DHC (without thermal storage), parameters of the ‘‘grade of operation’’ are based on an actual operation of the DHC plant without thermal storage, and the number of chillers/heat pumps is set to the mean value of the existing electric-driven heat-pump-type DHC plants. The chiller capacity is determined to be equally divided as that in the individual system (Section 2.1).
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heat recovery chiller to produce both cooling and heating heat corresponding to the daily balance of the heating/ cooling demand. The factorial analysis result is shown in Fig. 10. Economy of scale and the thermal storage effect demonstrate considerable energy-saving effect in the cooling heat source system. On the other hand, in the heating heat source system, it shows that economy of scale is a key factor to lower the EER value. The centrifugal heat pump in the heating mode is inferior to the airsource heat pump in the system COP; therefore, it is very important that the heating heat should be produced by highefficiency equipment such as a heat recovery chiller to achieve a high-EER value. 4. Effect of the other potential parameters on the EER value The DHC is considered to exhibit energy-saving potential by introducing a high-efficiency technique. In addition, in practice, not all the heat source systems in the individual systems are equipped with air-source heat pumps. In this section, the effects of the other factors, which have not been verified in Section 3 (the base case), are discussed.
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RUN-A: Introduction of centrifugal refrigerators in the individual system An air-source heat pump is used for producing cooling heat in the individual system in the base case. However, centrifugal refrigerators can be introduced for large-scale buildings even in the individual system. The case where centrifugal refrigerators are used instead of air-source heat pumps is simulated. RUN-B: Increase in the cooling demand during winter In the base case of the DHC, the heat recovery chiller cannot produce considerable heating heat since the ratio of the cooling demand to the heating demand during winter (from November to March) is only 0.28. We have simulated the case in which the hourly cooling demand during winter doubles in order to enable the heat recovery chiller to produce more heating heat. In this condition, the ratio of the cooling demand during winter reaches 0.56; however, this is not unrealistic when compared with the average cooling load of three other plants, as shown in Fig. 9. RUN-C: Introduction of variable-speed driven chillers Variable-speed driven chillers are superior to conventional chillers (constant-speed driven chillers) in terms of their change in the COP caused by part-load operation and the cooling water temperature, as shown in Fig. 11. We have simulated the case in which the conventional centrifugal chillers/heat pumps (only in the cooling mode) in the DHC are changed into variable-speed driven chillers/heat pumps. The COP values shown in Table 1 are used in this simulation. RUN-D: Utilization of river water as the heat source/sink in the chiller and heat pump One of the remarkable advantages of the DHC is that it is possible to utilize natural and unused heat sinks/sources such as sewage water and seawater. However, the A-Plant does not currently utilize them. We have simulated the case in which river water is utilized as the heat source/sink in the DHC model.
Fig. 10. Power-saving effect of the factors: (a) cooling heat source system and (b) heating heat source system.
Fig. 11. COP change by load factor and cooling water inlet temperature.
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Table 6 Simulation results with the additional parameters
5. Conclusions
Power consumption (MWh)
Individual (base case) Individual (RUN-A) DHC (base case) DHC (RUN-B) DHC (RUN-C) DHC (RUN-D)
EER
For cooling
For heating
Total
11,121 9,823 7,865 7,967 7,457 7,334
4402 4402 4183 3897 4183 3500
15523 14225 12048 11864 11640 10834
0.97 1.06 1.25 1.32 1.30 1.39
In RUN-B, the cooling demand is higher than the other cases. Therefore, the total power consumption of RUN-B is not comparable with the other cases.
The river water temperature and power consumption of pumps for transporting river water used in the model are derived from the measurement data of an existing DHC plant utilizing river water. The results of these simulations are shown in Table 6. The power consumption for cooling in RUN-A decreases by 12% when compared with the result of the base case of the individual system. However, when compared with the DHC, the power consumption is still greater. The major advantage of the economy of scale obtained by the DHC, as shown in Fig. 10, disappears in the case in which the centrifugal refrigerators are introduced into the individual system. However, due to the other advantages such as the thermal storage effect and the concentration effect, the DHC is superior to the individual system in RUN-A in terms of the EER value. In RUN-B, the ratio of the amount of heating heat produced by the heat recovery chiller increases to 39% from the result obtained in the base case (28%) by doubling the cooling demand during winter. As a result, the power consumption for heating is reduced by 6.8%. The result of RUN-C shows that the introduction of the variable-speed driven chillers reduces the power consumption for cooling by 5.2%. In the DHC, the chillers are operated at full load at all times due to thermal storage, and improvement in the part-load performance by introducing the variable-speed driven chiller is ineffective for saving energy. Thus, the introduction of the variable-speed driven chiller only provides the energysaving effect of improving the ratio of the COP change by cooling water temperature. The result of RUN-D shows that the utilization of river water is very effective in saving power. The river water temperature is distributed from 20 to 30 8C in summer and at intermediate time, and there is a slight discrepancy in the temperature between the river water and cooling water from the cooling tower. On the other hand, the river water temperature is distributed from 10 to 15 8C during winter; this temperature is considerably higher than the brine temperature from the heating tower (–5 0 8C). As a result, the COP of the centrifugal heat pump in the heating mode is considerably improved. Therefore, the energy-saving effect is evident, especially for heating. However, it should be noted that the energy-saving effect depends on various factors such as the temperature of river water and power consumption of the pumps for transporting the river water.
In this study, the energy-saving effect of the electric-driven heat-pump-type DHC is examined by using simulation models. In addition, the technique for improving the energy efficiency of the DHC is verified. The results are as follows: The simulation results show that the DHC provides an energy saving of 29% for cooling when compared with the individual system; meanwhile, the energy-saving ratio is 5% for heating. The difference in energy consumption between the DHC and the individual system is verified by means of a factorial analysis. The analysis result shows that the economy of scale (which indicates the effect of the introduction of the centrifugal refrigerators) and the thermal storage effect contribute significantly to saving energy for cooling. On the other hand, the disadvantage of economy of scale is shown for heating because most of the heating heat is produced by the centrifugal heat pump, which system COP is smaller than that of the air-source heat pump used in the individual system. The DHC can provide large energy-saving potential by improving the heat load profile or utilizing the natural and unused heat sinks/sources, both of which contribute to reducing the energy consumption particularly for heating. The heat load profile is proved to be one of the important factors for achieving a high-EER value. Therefore, it is necessary to thoroughly analyze the relationship between the EER value and the heat load profile. The utilization of the natural and unused heat sinks/sources is unique to the DHC, and the energy-saving effect is also significant. It is one of the most notable points of the DHC to be demonstrated in view of energy conservation from now on. A high-efficiency technique such as the variable-speed technique has the potential to reduce the energy consumption for an individual system as well as the DHC. When the highefficiency technique can be economically and technologically introduced into an individual system, there would be some changes in the energy conservation of the DHC. A simulation considering the progress of technology should be performed in order to verify the energy-saving effect of the DHC in the future. The result could be useful for verifying the validity of expanding the networks of pipes especially in Japan where the networks are not developed yet. The simulated EER of the individual system is 0.97, and this value exceeds the average EER value of 0.71 (obtained from the measurement survey) by 0.26 (The number of samples in this measurement survey is only 13, and the result could not be exact.) [14]. Thus, there could be some faults causing a decrease in the EER value such as aged deterioration, which are not observed in our analysis of the measurement data. A field survey on the individual system is required to reveal the actual condition of the individual system. Acknowledgements This work has been undertaken as a cooperative research with the Japan Heat Service Utilities Association. The authors
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wish to thank the members of the Special Committee on the Evaluation Method for Energy Efficiency of District Heating and Cooling, Society of Heating, Air-conditioning and Sanitary Engineers of Japan (SHASE) for their valuable advice. This research was supported by a Grant-in-Aid for Scientific Research from the Ministry of Education, Culture, Sports, Science and Technology, Japan, no. 18360273. References [1] M. Sakawa, K. Kato, S. Ushiro, M. Inaoka, Operation planning of district heating and cooling plants using genetic algorithms for mixed integer programming, Applied Soft Computing 1 (2) (2001) 139–150. [2] ASHRAE Research, 2004 ASHRAE Handbook, HVAC Systems and Equipment, 2004, pp.11.1–11.2. [3] R. Lazzarin, M. Noro, Local or district heating by natural gas: which is better from energetic, environmental and economic point of views? Applied Thermal Engineering 26 (2–3) (2006) 244–250. [4] T. Nagota, Y. Shimoda, N. Isayama, R. Kubara, M. Mizuno, Verification of energy saving advantage of co-generation incorporated district heating and cooling system, in: Proceedings of the IBPSA, Australia, (2006), pp. 141–148. [5] S. Sadohara, K. Nagano, et al., Study of diffusion of district heating and cooling in Japan and its effects on global environment preservation, Transactions of AIJ 510 (1998) 61–67 (in Japanese).
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