This paper investigates how to increase the coefficient of performance (COP) of air- cooled chillers satisfying the cooling demand of an existing office building.
Low energy operating strategies for air-cooled chillers serving an office building F.W. Yu1,*, H.Y. Chu1, K.T. Chan1, Cary W.H. Chan2, Paul S.K. Sat2 1 Department of Building Services Engineering, The Hong Kong Polytechnic University, Hong Kong 2 Technical Services Department, Swire Properties Management Limited, Hong Kong ABSTRACT This paper investigates how to increase the coefficient of performance (COP) of aircooled chillers satisfying the cooling demand of an existing office building. An analysis on the operating data of an existing air-cooled reciprocating chiller showed that the differential pressure requirements of the thermostatic expansion valve can limit the drop of condensing temperature even at a low set point of 22oC. A thermodynamic model was developed for the chiller to ascertain how the condenser fans should be staged to control the condensing temperature. It is found that the reset of condensing temperature should be based on changes of chiller load and outdoor temperature. Under the optimum condensing temperature control (CTC), the chiller COP can increase by 1–49%, depending on the operating conditions. A chiller plant operating for the cooling load profile of the building was simulated in order to assess the potential electricity savings by use of CTC. With regard to the building requiring an annual cooling energy of 7,155,039 kWh, CTC enables the annual electricity consumption of chillers to drop by 156,740 kWh from 3,927,450 kWh. When variable speed fans and evaporative pre-coolers are applied to the air-cooled condensers, an electricity saving of up to 489,383 kWh could be achieved in the chiller plant. This shows how the low energy operating strategies for air-cooled chillers help decrease the electricity demand for the local commercial sector. Keywords: air-cooled chiller, condensing temperature control, high static fans
用于办公楼风冷冷水机的节能策略 摘要: 摘要 本文研究如何增加用于办公楼风冷冷水机的表现系数(COP)。根据一个风冷冷水 机操作资料的分析，冷凝温度的下降受制于膨胀阀的压差。本文利用一个冷水机模型 发展控制冷凝器风扇与冷凝温度的最佳策略，结果发现冷凝温度的设定应该根据制冷 量和室外温度的变化。在优化冷凝温度的控制下，COP 可增加 1.0–49%。当这优化控 制应用于办公楼风冷冷水机，制冷用电量可从 3,927,450 kWh 中省 156,740 kWh。若优 化冷凝温度的控制与变速冷凝器风扇和蒸发预冷器同时应用，制冷用电量更可省 489,383 kWh。这研究反映发展风冷冷水机的节能策略有助减省商业大厦的用电需求。 1.
INTRODUCTION
Air-cooled chillers are commonly used to provide cooling energy for commercial buildings in the subtropical region but with considerable electricity consumption (Chan 2000; Lam, 2000; Yik et al., 2001). They conventionally operate with constant speed condenser fans under head pressure control (HPC). To maintain the condensing temperature at a high level of 45oC under HPC, the number of staged condenser fans is kept minimal in almost all 1
operating conditions. This helps minimize the fan power, but the condensing temperature fluctuates widely and the chiller COP drops considerably, especially under light load conditions with moderate outdoor temperatures. COP is defined as the cooling capacity in kW over chiller power—the sum of compressor power and condenser fan power—in kW. To reduce the growing electricity demand of the commercial sector, it appears desirable to increase the COP of air-cooled chillers through improving the design and operation of their components. Some studies have discussed how the design of air-cooled condensers should be revised and how the condensing temperature can be better controlled in order to increase the chiller COP. Reducing the condensing temperature as low as possible has been considered as a pragmatic approach to increasing the chiller COP (Briley, 2003; Smith and King, 1998; Roper, 2000; Manske et al., 2001). According to experimental tests on an air-cooled reciprocating chiller (Chan and Yu, 2004; Yu and Chan, 2005a, 2005b), condensing temperature control (CTC) is a viable alternative to HPC to increase the COP of air-cooled chillers. Electronic expansion valves are a prerequisite for the successful implementation of CTC because they can control the refrigerant flow properly even when the condensing temperature drops to 20oC. All the condenser fans under CTC are staged in most operating conditions. This causes an increase in the fan power, but the condensing temperature can hover closely above the outdoor temperature to maximize the chiller COP. Based on the thermodynamic model developed for the experimental chiller, the algorithm for implementing CTC is to adjust the set point of condensing temperature in response to changes of outdoor temperature. The chiller COP can be increased by lowering the condensing temperature only when the extent to which the compressor power can drop always exceeds the increase in fan power resulting from staging more condenser fans. CTC is generally applicable to air-cooled chillers containing low static condenser fans typically with a rated total power of 22 W per kW cooling capacity. For air-cooled chillers installed in an acoustic enclosure instead of an open space, high static condenser fans need to be used to cater for the additional pressure drop across the silencers. The rated total fan power then could be up to 77 W (350% of 22 W) per kW cooling capacity, which corresponds to 24% of the total rated compressor power. It is envisaged that the condensing temperature should be controlled at somewhere between its lower boundary and a high level of 45oC in order to maximize the chiller COP. At present there is no detailed investigation into how to modulate the set point of condensing temperature to achieve the maximum chiller COP when the condenser fan power is comparatively high. It remains to be analyzed whether variable speed condenser fans, instead of the constant speed ones, facilitate the control of condensing temperature at its set point with reduced power. To promote low energy operating strategies for air-cooled chillers, this paper investigates how the condensing temperature should be controlled to maximize their COP when the chillers with high-static condenser fans satisfy the cooling demand of an existing office building. An analysis was carried out on the operating data of an existing air-cooled reciprocating chiller to show how the differential pressure requirements of thermostatic expansion valves limit the drop of condensing temperature. A thermodynamic model was developed for the chiller to ascertain the optimum strategy for controlling the condensing temperature. The algorithm of resetting the condensing temperature is explained and the increase of chiller COP due to the optimum control is described. A chiller plant serving the building was simulated in order to assess the potential electricity savings by use of the optimum condensing temperature control and other advanced condenser features. 2
2.
DESCRIPTION OF THE BUILDING AND ITS CHILLER PLANT
2.1
Evaluation of Hourly Cooling Loads for the Office Building
The hourly cooling load of the building was computed by using a multi-zone building model within the simulation program TRNSYS (SEL, 2000). Details about the building shown in Table 1 was collected from the building operators and compiled into a building description file for the model. It was assumed that every piece of air-handling equipment was capable of delivering the cooling energy required to meet the cooling demand for the thermal conditions specified in each zone. The annual cooling energy was 7,155,039 kWh and the cooling hours accounted for 96% of the total 3182 office hours. Table 2 shows the frequency distribution of hourly building load ratios in different ranges of outdoor temperatures. The hourly building cooling load appeared to increase with the outdoor temperature and was concentrated mostly in a building load ratio of below 0.2. The building load ratio of 0.9–1 accounts for only 0.6% of the total cooling hours. 2.2
Details of the Chiller Plant
The building’s chiller plant was designed with nine air-cooled reciprocating chillers, each of which had a nominal cooling capacity of 943 kW. Seven chillers needed to operate to meet the peak cooling load of 5892 kW. Chiller sequencing was considered in the simulation so all the chillers were operating at the same load, and no additional chillers started to operate until each of the running chillers was operating at full load. Given seven steps of staging chillers to meet the changing building cooling load, the opportunity to operate the chillers at higher loads increased considerably. 56.3% of the total chiller load data were collected at a part load ratio of 0.8–1 at the entire range of outdoor temperatures. Furthermore, the chillers were able to work at a part load ratio of 0.9–1 with outdoor temperatures ranging between 15–34oC for over 26.5% of the operating time. Table 1 General information about the building and its HVAC systems General building details Gross floor area (GFA) (m2) Total air-conditioned area (m2) Orientation Window to wall ratio U-values of wall/window/roof (W/m2oC) Shading coefficient of glass Area per floor (m2); Number of floors Cooling temperature set point (oC) Relative humidity (%) Ventilation rate (L/s per person) Occupancy (m2 per person) Power density: equipment/lighting (W/m2) HVAC system operating hours Air side system details Type of air handling units Chiller plant details Total cooling capacity installed (kW) Number and type of chillers Nominal capacity of each chiller (kW) Chilled water distribution system Number and type of chilled water pumps Nominal power input to each pump (kW)
60568 48531 (80% GFA) N/E/S/W 0.53 2.3/4.6/0.7 0.4 2884; 21 23.5 50 10 9 25/15 0830-1730 (Mon-Fri); 0830-1300 (Sat) PAU and VAV AHU 8487 9 air-cooled reciprocating chillers (7 duty, 2 standby) 943 Two-loop pumping system with decoupling by-pass pipe 9 constant speed primary-loop pumps; 3 variable speed secondary-loop pumps. 3.6 for primary; 110 for secondary. 3
Table 2 Frequency distribution of hourly building cooling loads Outdoor Building load ratio (Building cooling load expressed as a ratio of its peak value) temperature 0–0.1 0.1–0.2 0.2–0.3 0.3–0.4 0.4–0.5 0.5–0.6 0.6–0.7 0.7–0.8 0.8–0.9 0.9–1 Subtotal (oC) 297 63 0 0 0 0 0 0 0 0 360 15–17 17–19 133 121 7 0 0 0 0 0 0 0 261 19–21 82 146 94 23 0 0 0 0 0 0 345 21–23 36 70 116 68 43 1 0 0 0 0 334 23–25 34 11 41 82 66 46 6 2 0 0 288 25–27 17 24 4 22 105 150 63 22 1 0 408 27–29 18 45 0 2 46 93 113 123 22 1 469 29–31 2 2 0 0 6 16 50 232 92 12 412 31–33 0 0 0 0 0 0 12 72 92 4 180 33–35 0 0 0 0 0 0 0 1 3 1 5 Subtotal 619 482 262 197 266 306 244 452 210 18 3062
2.3
Features of the Chiller Model
The model is based on an existing chiller within the chiller plant. The chiller uses the refrigerant R134a and contains two modules, each of which has a nominal cooling capacity of 472 kW. The COP is 2.4 for the chiller operating at full load at an outdoor temperature of 35oC. Each chiller module consists of a shell-and-tube liquid evaporator, a thermostatic expansion valve (TXV), two constant speed reciprocating compressors and an air-cooled condenser. The evaporating temperature of the evaporator is designed to be 3oC. The temperature of supply chilled water is set at 7oC with a temperature rise of 5.5oC at full load. The chilled water flowing through the chiller is maintained at 40.2 kg/s. Each of the compressors has a nominal power of 81 kW and provides four capacity steps via unloading the cylinders. The condenser is equipped with 16 constant speed condenser fans, each of which is rated at 3.73 kW and delivers a heat rejection airflow of 3.9 m3/s. Each compressor operating is linked to the staging of two or four condenser fans. The operation of condenser fans is that the first two fans are staged with each operating compressor and the next two fans for that compressor are staged until the condensing temperature reaches its set point. The chiller has a low and fixed condensing temperature set point of 22oC, instead of 45 C used typically under HPC. This set point corresponds to the lower boundary of the condensing temperature based on the lowest outdoor temperature of 12oC for the chiller operation and based on the design of heat rejection capacity with a 10–15oC difference between the condensing temperature and outdoor temperature. With the low set point of condensing temperature, the condenser fans could be staged as many as possible in part load conditions in order to allow the condensing temperature to hover slightly above any given outdoor temperature. o
The operating variables of the chiller were collected continuously at 0.5 hr intervals throughout the year. Steady-state operating data were considered in evaluating the chiller performance and verifying the chiller model. The variables monitored included the temperature of supply chilled water (Tchws), the temperature of return chilled water (Tchwr), the number of steps of staged compressors, evaporating pressure (Pev), condensing pressure (Pcd), outdoor temperature (Tcdae), and compressor power (Ecc) in terms of the percentage of full load ampere (FLA). The condensing temperature (Tcd) and evaporating temperature (Tev) were computed based on the measured Pcd and Pev, respectively. The calculation of cooling capacity (Qcl) is based on a fixed chilled water flow rate of 40.2 kg/s, the specific heat 4
capacity of water of 4.19 kJ/kgoC, and the temperature rise (Tchwr–Tchws). A thermodynamic model for the chiller was developed using TRNSYS (SEL, 2000). Detailed information about the development of the model was given elsewhere (Chan and Yu, 2002; Yu and Chan, 2006a, 2006b, 2006c). The model considers the real process phenomena, including the capacity control of compressors and mechanistic relations between chiller components. In simulating the operation of a chiller, the compressors, expansion valve and condenser have to satisfy the mass balance of refrigerant and energy balance at the evaporator. An algorithm was introduced to compute the number and speed of staged condenser fans based on a set point of condensing temperature. The model is sophisticated enough to investigate how different condenser features influence the steady-state behaviour of chiller COP at various combinations of chiller loads and outdoor temperatures. This simulation analysis considered individual and mixed uses of three enhanced condenser features: CTC, variable speed condenser fans (VSF) and evaporative pre-coolers (EC). CTC refers to the optimum reset of condensing temperature in order to maximize the chiller COP. Compared with constant speed condenser fans, the use of VSF helps facilitate the control of condensing temperature at its set point with reduced power. An EC installed in front of an air-cooled condenser can cool the outdoor temperature from its dry bulb to wet bulb to further lower the condensing temperature. 3.
RESULTS AND DISCUSSION
3.1
Operating Constraints of Thermostatic Expansion Valve (TXV)
The number of staged condenser fans (Ncf) depends on the operating conditions and how the set point of condensing temperature (Tcdsp) is controlled. Figure 1 gives a logic flow diagram for the control of the condenser fans. For any given heat rejection (Qcd), the minimum heat rejection airflow (Va,min) was determined based on Tcdsp and the outdoor temperature (Tcdae). Ncf was then calculated in conjunction with the required airflow (Va) based on Va,min. All the 16 condenser fans needed to be staged to provide the highest airflow when Tcdsp was very close to Tcdae. The condensing temperature (Tcd) floated above Tcdae in various degrees when the corresponding condensing pressure (Pcd) was above its lower limit based on the differential pressure requirements (∆P = Pcd–Pev) of the TXV. Fewer condenser fans were staged to raise Tcd in order to allow the coincident Pcd to comply with the differential pressure requirements. The reduction of Tcd was bounded by the minimum ∆P in such circumstances. The controllability of condensing temperature (Tcd) at various outdoor temperatures (Tcdae) was studied, given that the chiller operated with TXVs at a low Tcdsp of 22oC. Table 3 summarizes the Tcd measured at various combinations of chiller part load ratios and Tcdae. Instead of using HPC with a high Tcdsp of 45oC, the low set point of 22oC helped narrow the difference (Tcd–Tcdae). Yet Tcd varied widely from 22.0 to 51.4oC rather than hovering around its set point. The difference between Tcd and the set point of 22oC tended to be significant when the chiller load was high. At an outdoor temperature of 31oC, the condensing temperature deviated from its set point by 19.7oC at a part load ratio of 0.5 and by 29.4oC at full load. This indicates that a considerable differential pressure is required for the TXV to throttle higher flow rate of refrigerant at higher chiller loads, based on the typical expansion valve model: mr = Kex ρex0.5(Pcd–Pev)0.5, where mr is the refrigerant flow rate, ρex is the density of the liquid refrigerant before expansion and Kex is a proportionality constant and is changed 5
as required to maintain the superheat in the evaporator. Given this situation, not all the condenser fans were staged for most operating conditions to maximize the heat rejection airflow required to increase the chiller COP. Unlike TXVs requiring a ∆P of at least 700 kPa, electronic expansion valves (EEVs) can work at a small ∆P of about 100 kPa while controlling the refrigerant flow to meet the varying cooling capacity. The use of EEVs enables the condensing temperature to hover closely to its set point. INPUTS Tcdae, Tcdsp
Tcdae < Tcdsp?
N
Calculate Tcdsp,min based on ∆P requirements of TXV at any given Qcl
Y
Va,min =
Qcd ρ a C pa (Tcdsp − Tcdae )
Tcdsp = Tcdsp,min List of symbols: Cpa specific heat capacity of air (1.02 kJ/kgoC) Ncf number of condenser fans staged heat rejection, the sum of Qcl Qcd and compressor power Ecc (kW) Qcl cooling capacity (kW) Tcdae outdoor temperature (oC) Tcdsp condensing temperature set point (oC) Va heat rejection airflow delivered by the staged condenser fans (m3/s) ∆P differential pressure across TXV (kPa) ρa air density (1.2 kg/m3)
Va,min Integer N cf,tot ≤ N cf V a, tot Ncf
